diff --git "a/designv6-24.json" "b/designv6-24.json" new file mode 100644--- /dev/null +++ "b/designv6-24.json" @@ -0,0 +1,13809 @@ +[ + { + "image_filename": "designv6_24_0001650_pcicon.2015.7435110-Figure17-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001650_pcicon.2015.7435110-Figure17-1.png", + "caption": "Fig. 17: Blower pedestal between motor and blower", + "texts": [ + " TROUBLESHOOTING CASE STUDIES The following case studies illustrate how easily high motor vibration can initially appear to come from one source but in reality is resulting from a different excitation source. Electrically induced vibration can excite selective substructures that are not immediately part of the motor itself. This introduces some confusion to the drive train operator. For example, a 6 pole motor running a blower in a power station ran quite smoothly with very low vibration on the housing and shaft. There was no significant vibration on the motor frame. The inboard blower bearing pedestal (shaft extension end) experienced a significant 120 Hz vibration, see Fig. 17. The spectrum in Fig. 18 shows the vibration on the blower inboard bearing pedestal in the horizontal direction. The running speed component (1X) is very small but the 7200 cpm (120 HZ) component is 0.18 ips (4.57 mm/sec) which is quite significant. In this case the motor inherent twice line excitation is small enough (.02 ips or .5 mm/sec) not to excite the motor frame, but found a sympathizing fan-bearing-pedestal structure that is tuned to 120 HZ and excited it. The vibration is electrical by definition since it will disappear immediately as soon as the power is removed" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001137_icelmach.2008.4800194-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001137_icelmach.2008.4800194-Figure1-1.png", + "caption": "Fig. 1. EMSM geometry for 1 pole.", + "texts": [ + " n ncuth T TIIPRT \u0394 \u0394=\u21d2\u22c5=\u0394 (1) Where delta \u2206T represents the temperature difference between the winding and the ambient air, Pcu the copper losses, Rth the thermal resistance, and I and In are the nominal current for \u201340 \u00b0C and 100 \u00b0C respectively. The EMSM modeled is a 12-pole, three-phase machine with 180-degrees distributed winding in the stator and salient poles in the rotor. The length of the machine is 60 mm and the outer radius of the stator is 85 mm. Besides, a slotting factor of 0.5 is considered for the stator core. The resultant geometry is illustrated in Fig. 1. In order to find the optimal relation between the air gap radius and the inner radius of the rotor with respect to the outer radius of the machine several alternatives are analyzed. For each combination, first a thermal FE simulation is conducted in order to find the maximum current that fulfils the temperature limitation by means of an iterative process. Subsequently, a series of magnetostatic FE simulations is made in order to find the torque produced for different armature current loading cases" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001291_bf03378615-Figure3-4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001291_bf03378615-Figure3-4-1.png", + "caption": "Fig. 3-4-high rolling mill threaded for C-B-S rolling.", + "texts": [], + "surrounding_texts": [ + "The General Electric Co. has been studying C-B-S rolling since 1958. During this time it has carried out a variety of experimental programs on several experimental mills. A considerable amount of roll ing has been undertaken on a 12-in., 4-high, cold, reversing tension strip mill threaded as shown in Fig. 3. Here the 2 work rolls are driven in speed ratio. Strip 12 in. wide and up to 0.150 in. in thick ness has been C-B-S rolled on this mill. Materials rolled include mild steel, 18-8 stainless steel, Rene 41, silicon-iron, and others. Of particular interest is the EMERGING STRIP PRODUCT INTERSTAND AREA ~ A-ENTRY CONTACT ROLL B-MIDDLE CONTACT ROLL C-EXIT CONTACT ROLL 1- FIRST STAND BEND ROLL 2-SECOND STAND BEND ROLL ai-FIRST STAND TOGGLE ANGLE a2-SECOND STAND TOGGLE ANGLE I i I.---------.J ISTANDI #1 ENTERING STRIP Fig. 2-Basic configuration of C-B-S tandem rolling mill. fact that stainless steel can be rolled from 0.050 to 0.004 in. in 8 passes without an intermediate anneal." + ] + }, + { + "image_filename": "designv6_24_0002245_aim.2005.1511046-Figure9-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002245_aim.2005.1511046-Figure9-1.png", + "caption": "Fig. 9 Bode plots of the position loop.", + "texts": [ + " The controller gains can therefore be obtained by the following equations, and the design results are listed in In addition, the constraints can be formulated as: 222 2 2 12211 2 2 2 1 2 )4( allmecha )()( 2 1 2 22 22 211 mechmech After the speed controller design is accomplished, design for the position loop gain can be preceded. As shown in Fig. 8, the only proportional gain for the position loop is ppK . Consequently, the design objective RI can be expressed as an ITAE (Integral Time Absolute Error) criterion, and constraints can be summarized as below and the resulting design parameters for X axis and Y axis are shown in Table 2. 0 4.14)(. )(min e tTts dttetI R , (12) where mpmppi mmr KdtKeKtT e )()( From Fig. 9(b), it appears that the position loop bandwidth of the Y axis, 12 Hz, is unable to achieve the design specification, which is 15 Hz. Therefore, a redesign process needs to be performed. According to (7) and (8), the nature frequency for the feed drive system is determined by two parameters iK and mech . Therefore, enhancement of the system\u2019s performance relies on modification of those two parameters. However, the control variable iK is chosen based on the mechanical characteristics. In order to improve response of the system, a new design has to be carried on" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001452_j.isprsjprs.2008.10.001-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001452_j.isprsjprs.2008.10.001-Figure3-1.png", + "caption": "Fig. 3. Possible tilt of the camera in direction of optical system.", + "texts": [ + " After this, the maximum length of residual vector was still over four pixels and residuals in y-direction were still slightly bigger than in x-direction. Also, the comparison to the reference showed that the 3D points from computation were systematically farther away than in the reference data. Conclusion of this discrepancy was that the camera must have tilted during imaging. This indicated that the assumption of the direction of optical axis to be staticwith respect to the planewas violated. Fig. 3 demonstrates the suspected phenomenon. The tilting can be considered as a vertical plane rotation. Some vertical and horizontal displacement of a projection center may have occurred as a consequence of this tilt. This, however, can be neglected since the maximum value of the tilt \u03bd was less than 1\u25e6 and the maximum effect in that case could have been less than a millimeter. So the effect of the tilt can be appended as an additional angle into calculation of the rotation matrix. R\u03c9i,\u03c6i,\u03bai = R\u03c90,\u03c60,\u03ba0 \u00b7 R\u03b1i,\u03bdi,\u03b8i " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001026_tap.2015.2487512-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001026_tap.2015.2487512-Figure1-1.png", + "caption": "Fig. 1. Dielectrically loaded parallel plate waveguide (PPW) array made mostly of metal for ease of fabrication where a rectangular dielectric plunger will be positioned within the PPW feedline to control the phase delivered to each slot element (and thus scan angle). The final array dimensions will be approximately 12\u201d \u00d7 2\u201d \u00d7 0.5\u201d, with its height increasing to 0.6\u201d as the plunger position is adjusted.", + "texts": [ + " Nevertheless, lumped elements can be lossy and therefore may not be suited for spaceborne applications. Ferrites have also been proposed to induce beam steering by controlling the substrate\u2019s permeability [25]\u2013[29]. However, ferrite solutions tend to be heavy due to using a biasing magnet and may also be lossy. Another approach is to use multiple ports to create beam steering [30], but many ports are needed for fine beam resolution. In this paper, we avoid the aforementioned issues by instead using a movable dielectric plunger within the feedline of the TWA. Figure 1 depicts the proposed TWA configuration. This TWA employs a parallel plate waveguide (PPW) feed supporting a propagating wave that excites slot antennas to the right in a sequential manner. The propagation constant within PPW is controlled using a small, linear mechanical movement (Figure 2). This motion changes the feedline\u2019s air to dielectric ratio and has several advantages: 1) low cost, 2) low complexity, 3) large total bandwidth, 4) large scan ranges (compared to frequency scanned arrays). In [31] we employed a parallel plate waveguide (PPW) transmission line (TL) feed by placing a dielectric coating at the inner side of the planar strips forming the TL", + " Specifically, the rectangular plunger and the solid metal plates can be more precisely fabricated for Ka and Ku band applications. Below we present the design of the new TWA concept, its excitation, beam steering performance, and validation using a fabricated prototype operating at 13GHz. As noted, beam steering in a TWA is achieved by controlling the propagation constant of the feedline. Here, we employ a dielectrically loaded parallel plate waveguide (PPW) TL to feed an array of cavity-backed slots as depicted in Figure 1. A rectangular dielectric is positioned within the vertical PPW plates. Specifically, a section of the PPW transmission line is depicted in Figure 2 and as the dielectric plunger moves up and down along the z-axis, the propagation constant changes. The key parameters of the dielectrically loaded PPW are depicted in Figure 3. As expected, the maximum and minimum effective dielectric constant (\u03b5eff ) of the PPW depend on the position of the dielectric plunger. Also, the dielectric composition of the plunger plays a crucial role in determining \u03b5eff ", + " A second coax port was also placed at the far end of the PPW to measure the remaining/unradiated signal with the PPW (viz S21). Initially, this is of a width SR = 30mil and then tapers to 120mil, the width of the PPW. The other Figure parameters are LR = 50mils, LE = 100mils, and LT = 50mils. We remark that, the feed exhibits only 1 to 1.5dB of insertion loss, implying proper operation. Combining the dielectrically loaded PPW feedline, nonrectangular cavity-backed slot, and coaxial probe feedline excitation, a 20 element linear array was created (Figure 1). Aluminum is chosen for the metal substructure to ensure both machinability and high conductivity. The slot width taper 0018-926X (c) 2015 IEEE. Personal use is permitted, but republication/redistribution requires IEEE permission. See http://www.ieee.org/publications_standards/publications/rights/index.html for more information. (Wi for i = 1, 2, ..., 20) was determined through the design procedure outlined in Figure 12. Here, the desired pattern at some scanning angle is determined by the Kaiser array weighting coefficients [33]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000940_s1474-6670(17)66076-2-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000940_s1474-6670(17)66076-2-Figure6-1.png", + "caption": "Fig. 6 Basic Geometry Modelled in the Simulation", + "texts": [ + " After allowing perturbations to the vehicle motion to damp to a suitable level, vehicle control was transferred to the Agena gravity gradient CMG Control System. This experiment is a solar pointing array controlled by applying torques to two orthog onal gimbal axes. The \"inner\" axis is ori ented along the Agena longitudinal (X) axis. This is referred to as the support axis, while the remaining axis is called the drum axis. Input signals to the gimbal controller are derived from a 30 degree field-of-view (conical) sun sensor, see Fig. 6, a (2 deg x 360 deg) field-of-view (wedge) sun sensor and two tachometers which define the rotational rates relative to the Agena axes. 1218 J. J. Rodden The basic requirement placed upon the RTD 806 experiment was to track the sun whenever the satellite was not in the earth's shadow. This implied keeping the sun vector within a 30 deg \u00b1 15 deg (half angle) cone about the normal to the sensitive side of the array. This is reflected in a \u00b1 15 deg attitude excursion bound on all axes placed upon the Agena gravity gradient CMG control system" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001472_9781118354179.auto089-Figure31-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001472_9781118354179.auto089-Figure31-1.png", + "caption": "Figure 31. Basic clutch schematic and dimensions.", + "texts": [ + " In addition to the above-mentioned modern pneumatic clutch actuation system (Figure 28), which can be automated, two other systems shown earlier, centrifugal clutch (Figure 6) and electromagnetic clutch (Figure 7), can also be relatively easily automated. Both of these are older designs not currently used. Instead, sophisticated hydraulic and electric systems have been developed, as shown in Figures 29 and 30. Obviously, such systems require outside power to operate, that is, hydraulic pressure for the (LuK) system shown in Figure 29 and electric current for the (ZF) system shown in Figure 30. Fenton (1996) gives more detailed insight into the evolution of automated clutch control systems. Figure 31 shows a clutch schematic and basic dimensions. The outer diameter of the friction disc is OD, inner ID, and total spring force, normal on friction surfaces, is FN. Engine angular velocity is \u03c9E and clutch friction disc velocity is \u03c9C. It should be noted that engine velocity \u03c9E is identical to the flywheel velocity and all clutch components fastened to the flywheel also have the same velocity. The only exception is the friction disc, whose velocity \u03c9C can be different, as a result of clutch being disengaged or slipping" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000083_ir-03-2018-0042-Figure12-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000083_ir-03-2018-0042-Figure12-1.png", + "caption": "Figure 12 Forces acting on the revolute joint", + "texts": [ + " The knee and ankle joints are fixed at 0\u00b0. For the knee joint, the worst case when u 2 rotates to peak at 70\u00b0, and reverse the direction as the wearer walks to reach up to 0\u00b0 again. The ankle joint is fixed at 0\u00b0. The ankle joint worst case when u 3 rotates to All-terrains wearable vehicle B.M. Sayed et al. Industrial Robot: the international journal of robotics research and application Volume 46 \u00b7 Number 6 \u00b7 2019 \u00b7 740\u2013762 peak at 30\u00b0, and reverse the direction as the human walk to reach at 30\u00b0. Figure 12 describes the force acting on the revolute joint. The revolute joint worst case when u 4 rotates to peak at 30\u00b0 from the vertical, and reverse the direction to obtain the stability while in walking mode. The torque required for each joint can be determined by using equations (16)-(19). The simulation analysis of the torque required is applied by using SolidWorks package software. The simulation results are shown in Figure 13. For hip joint, the torque pattern of each joint obtained from the simulation for the exoskeleton to perform one leg swing motion" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001238_ias.1998.732255-Figure18-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001238_ias.1998.732255-Figure18-1.png", + "caption": "Fig. 18. Circle diagram of saturated IPM motor", + "texts": [ + " The first three solutions, at very low current, allow Am, Ld, Lq to be estimated. They are Am=O.O7Vs and, referring to Figs. 14- 15, L~0.29mH, Lq=0.59mH. Two more field :;elutions, with high q-axis current components, allow I\\ps and L, to be determined. If &=O. 12Vs at ic250A and &=o. 14Vs at iq=350A (see Fig.15) are imposed itwice in the second of (17), one can achieve A,=O.O7Vs and L,=0.2&. Then, even in saturation region (iqXqS): an easy representation can be still obtained in the id-iq plane. A drawn of the characteristics in the h i q plane is reported in Fig. 18. In fact, the constant torque curves are modified hyperbolae, and the voltage limits are adjusted ellipses, given by For design purpose, the simple model characterised by (16) (17) Ad(&) = Am + Ldid = A, + Lq& t = - p Amiq - A&) + ( Ld - ~ J i ~ i , ~ ] 2 [( [ = (A, + Ldid)2 + (Aq + Lqsiqy IX. CONCLUSIONS While in unsaturated conditions the inductive parameters and PM flux linkage of an IPM synchronous motor can be univocally determined, under saturated operation they may be defined in dflerent ways" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002582_j.neunet.2005.07.014-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002582_j.neunet.2005.07.014-Figure7-1.png", + "caption": "Fig. 7. Graphical evolution of two prehension gestures. (a) Power grasp. (b) Precision grip. Movement of the arm is only in the transverse plane (XY plane). Dotted line represents the inverse kinematics computations carried out by DIRECT model at each time step.", + "texts": [ + " It is also important to note that systems that use directional mappings as the system presented in this paper, can successfully reach targets even if the directional mapping contains a large amount of error. Therefore, any residual error that might exist; e.g. from assuming linearity over too large a region of the workspace, will not prevent the system from reaching targets, but will instead only lead to curvature in the movement trajectories. We have carried out systematic computer simulations of the neural model for coordination of hand gestures during prehension. In Fig. 7 are shown three key instants of two different prehension tasks. On the left column are shown the initial instant of the movement (upper left figure), the instant of maximum grip aperture at w60% of movement time (middle left figure) and final grasp configuration (lower left figure) for a power grasp of a cylinder. On the right column are shown the initial instant of the movement (upper right figure), the instant of maximum grip aperture at w60% of movement time (middle right figure) and final grasp configuration (lower right figure) for a precision grasp (thumb opposed to one finger) of the same cylinder", + " In all simulations initial hand shape is described by {0, 7} and initial palm orientation is aZ0, bZ0, gZ0. In Fig. 8 are shown the kinematic curves for velocity and acceleration of wrist (left) and grip aperture and grip aperture velocity (right) for the Precision Grip task carried out under three different movement velocities. The movement velocity is controlled by varying the G0 parameter. In our simulations the lowest movement velocity is related with a value of G0Z15. Medium movement velocity was simulated using G0Z20 (this is the G0 value used in simulations shown in Fig. 7) and high velocity movements were simulated using G0Z25. Transport velocity exhibits a bell shaped but asymmetrical velocity profile typical of point to point arm movements. The plot of hand aperture shows the opening of the hand until it gets to maximum peak aperture, then it shows a decreasing of grip aperture until it reaches object size. Due to the computational structure of the model, time of maximum grip aperture is always correlated with tpdec, and the spatio-temporal coordination of transport and grasp channels remains unchanged through variations of speed performance (Fig. 8) or task demands (Fig. 7). Some emergent properties of the neural model match stereotyped kinematic profiles in humans. As shown in Fig. 8, as velocity of transport movement increases, the maximum grip aperture (the maximum measured distance between thumb and index finger tips along the movement) also increases. As a result of the GO signal shared by reaching and grasping components, this well-known kinematic feature in humans is achieved in the model without any explicit information transfer between reaching and grasping components. In Fig. 9 it is explicitly showed the temporal correlation between the kinematic variables of transport channel and finger pre-shaping channel in the task showed by Fig. 7(a). It is important to note that the grip aperture profiles observed are driven by the evolution of temporal weightings of hand synergies. Fig. 10 shows the temporal evolution in the parametric space of the model described in Section 3.1.1, of four simulated tasks related to the grasping of four different objects (a sphere G1Z{K7, 5}, G2Z{K3, 7}, an egg G1Z{K 5, K3}, G2Z{K1, 0}, a block G1Z{K10, K7}, G2Z{K9, 2} and a cylinder G1Z{K7, 4}, G2Z{K3, 6}). The VITE model related with finger pre-shaping generates the bidimensional continuous trajectory in the weightings of hand synergies space (Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003898_iccasm.2010.5620754-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003898_iccasm.2010.5620754-Figure4-1.png", + "caption": "Figure 4. Poincare Section ofHenon-Heiles System at Level E=0.07625.", + "texts": [ + " Consider Henon-Heils system, this is a two freedom Hamilton system with Hamiltonian 1 ( z z z Z ) z 1 3 H =- PI + pz +ql +qz +qlqz --q, 2 3 (12) For a given total energy constant E, motion of the system (12) are restricted in a three-dimensional manifold. Henon and Heiles had made a research in 1964. They found, when the energy level become higher, the torus of KAM become distorted and get destructed gradually, as well as the regular motion of the system become chaos. The numerical results give a precious material for understanding of KAM theory. Here we make a more detailed computation using our method, different Poincare sections are defined at different energy levels. See Fig 4, 5, 6. 2010 International Conforence on Computer Application and System Modeling (ICCASM 2010) V6-382 For Fig 4(a), Fig 5(a) and Fig 6(a), \ufffd is defined by {XE R4 : X20 = o} , here x = {QI'Q2'pl'pzl . They are similar to the pictures given by Henon and Heiles previously, but the energy levels are slightly different. For Fig 4(b), Fig 5(b) and Fig6(b), \ufffd is defined by {x E R4 : XIO = o} . They present a helical symmetry structure different from the formers obviously. All of these figures demonstrate the process that the system evolves from regular motion to chaos. ACKNOWLEDGMENT The support of Natural science foundation of ShanXi (No: 2007011018) is gratefully acknowledged. REFERENCES [I] M. Henon, On the numerical computation of Poincare maps, Physica, D 5, (1982), 412-414. [2] Warwick Tucker, Computing accurate Poincare maps, Physica, D 171, (2002),127-137" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001472_9781118354179.auto089-Figure24-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001472_9781118354179.auto089-Figure24-1.png", + "caption": "Figure 24. Hydraulic clutch control system. (Reproduced with permission from ZF Friedrichshafen AG.)", + "texts": [ + " Efficiency is typically low because of friction in contacts and joints, the cable stretches, with cable ends (connections) prone to snapping off. Such a system is relatively easy to install if the cable is short and access clear, not requiring sharp bends. However, in more demanding installations and when servo assistance and/or automation is required, hydraulic and other systems are used. Hydraulic clutch control systems are widely used on passenger cars, commercial, and other vehicles, owing to the ease of routing and installation, high efficiency, reliability, ease of servo assistance, and automation. Figure 24 (from ZF Sachs) shows such a modern system, together with the clutch assembly. Components present: dual mass flywheel (1), clutch cover (2), thrust bearing (3), pedal vibration damping device (4), master cylinder (5), pedal box (6), slave cylinder (7), and friction disc (8). Pedal vibration device (4) shown in Figure 24, is not always used, but can be very effective in suppressing axial crankshaft vibrations transfer to the pedal. Different designs are possible, with the solution shown using two diaphragms, with multiple fluid re-directional passages. Initial and service bleeding of the hydraulic system from air presence can be a complex process; therefore, special systems and procedures have been developed to ease the installation and servicing. Modern plastics have found application for manufacture of all components, including cylinders and the entire pedal box, as shown in Figure 24. This can drastically reduce both the price and mass of the system. No doubt, a supply fluid bottle is necessary (not shown in Figure 24), which is positioned above the master cylinder (5) and connected via a tube attached to the angular input shown in Figure 24. Encyclopedia of Automotive Engineering, Online \u00a9 2014 John Wiley & Sons, Ltd. This article is \u00a9 2014 John Wiley & Sons, Ltd. DOI: 10.1002/9781118354179.auto089 Also published in the Encyclopedia of Automotive Engineering (print edition) ISBN: 978-0-470-97402-5 Clutch hydraulic systems use brake fluid, which has low viscosity and can operate well in wide temperature range. Sometimes, instead of using a separate reservoir, brake (twin) reservoir is used for fluid supply. The clutch is always connected toward the higher fluid level, ensuring that if the brake fluid level drops, the clutch control function is lost prior to brake system function being endangered. There is also a low fluid level light warning to the driver. Instead of using a separate slave cylinder (item 7 in Figure 24), fork, and thrust bearing (3), it is becoming more common for passenger cars to use a simple assembly, concentric slave cylinders shown in Figure 25, which replaces all three components. Placed at the position of the thrust bearing (3 in Figure 24), the output shaft runs through the middle of the unit. This slave cylinder and its piston have large central holes. Obviously, such a hollow cylinder requires two seals\u2014one on the inner and one on the outer diameter, as well as careful guiding. The effective piston area can be quite large, multiplying the input pedal force. Concentric slave cylinders reduce the number of components, and load the housing symmetrically and centrally, making the entire transmission system lighter and simpler. Typically, passenger cars do not require assistance for clutch control, as the clutch spring, and therefore pedal forces are relatively low. However, heavier commercial vehicle engines have much high torques, requiring much greater normal spring force in the clutch assembly. Consequently, the pedal forces are considerably increased even becoming impossible for the driver to operate the pedal. Assistance is typically provided by using compressed air. Figure 26 (Nunney, 1998) shows a principle of the operation of such an assembly. It is important to bear in mind that the actuation is principally still as shown in Figure 24, with hydraulic master and slave cylinders, and compressed air assistance. Various actual solutions are possible, with Figure 27 showing an actual production unit (from Wabco). More recently, in order to make heavy commercial vehicles even more efficient, a purely pneumatic actuator has been developed to control clutch engagement\u2014ConAct release cylinder by ZF, as shown in Figure 28a. Control of the cylinder (Figure 28b) is achieved with electro-pneumatic valve, using a travel (position) sensor and vehicle operating data, such as engine speed and load (torque), vehicle speed, clutch pedal position, and accelerator pedal position" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001268_2013.25362-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001268_2013.25362-Figure1-1.png", + "caption": "Figure 1. The FR\u2010SCPS shank of 2.7\u2010cm thickness contains octagonal ring force transducers and both active (1) and dummy (2) cutting elements.", + "texts": [ + " The new design incorporated significant changes in the geometry of the sensor. The new concept aimed for a thinner unit with sufficient structural rigidity combined with a shape that could induce self\u2010penetration of the sensor into the ground. Thin octagonal ring\u2010type load sensing units were custom\u2010designed. Moreover, the top 8 cm of the blade were not instrumented since the analysis of data from this layer did not provide useful information in the case of the first prototype of the SCPS (Andrade\u2010Sanchez et al., 2007). Figure 1 displays a 3\u2010D rendering of the design. The main features of this new design were the 90\u2010degree rake angle, reduced shank width of only 2.7 cm, and the use of five customized octagonal ring load sensing units, which resulted in an effective sensing depth that ranged from 7.5 to 45.7 cm. These load\u2010sensing units were designed based on their relative location along the depth and expected load in order to maintain similar sensitivity levels among all five sensing units. Although the primary reason for using the octagonal ring type load sensing unit was to reduce the thickness of the compaction profile sensor, they are relatively inexpensive compared to commercial load cells" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000957_1.1559581-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000957_1.1559581-Figure8-1.png", + "caption": "Fig. 8 Trimming operation for a curve", + "texts": [ + " In case of the modification of trimmed surfaces, trimming curves are handled in the same way as the indicated curves described in the previous subsection, and surfaces are constrained by these curves. We use Type 3 correction function in Fig. 2 by which a Cardinal spline is multiplied. The parameter t1 in Fig. 2d is changed to the value at the trimming curve from that of the boundary, and the correction function is extended to the boundary keeping its value the same. First, we explain the process of modifying a curve with Fig. 8. The curve is trimmed at parameter value tT and modified smoothly to pass through Q1 by moving P1 ~at parameter value t1) on the original curve in the normal direction. The distance between P1 and Q1 is L1 ~see Fig. 8a!. The correction function for point indication is used at parameter value t1 , and that for trimming is used at parameter value tT ~see Fig. 8b!. The correction values l1 , lT are calculated by Eq. ~2!. The correction functions multiplied by these values are shown in Fig. 8c. The center curve is the total of the functions and pass through L1 at t1 and 0 at tT . The region where a surface is trimmed is defined by parameter function C(u ,v), which represents the curve specified as a boundary. When the curve is given by a designer, we find the parameter function with a spline as described in Subsection 2.2.2. An example is shown in Fig. 9. On the left are the correction functions, which are extended to the outside of the trimmed curve. We ring DECEMBER 2002, Vol. 2 \u00d5 269 ess" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000706_1.1479690-Figure9-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000706_1.1479690-Figure9-1.png", + "caption": "Fig. 9 Photograph of test bed", + "texts": [ + ", we obtain the torque generated by the TBC and the final input torque depicted in Fig. 8. According to the criterion expressed in Eq. ~18!, we can find that the TBC system substantially reduces the peak value of the original input torque up to 50%. To verify the effectiveness of the counterbalancing torque, according to the working conditions, dimensional parameters, and the design results obtained above, a prototype system having a spring-loaded TBC mechanism attached to a GCIM photographed in Fig. 9 is constructed. To measure the indexing accuracy of the GCIM, the applied instruments are tabulated in Table 3. A schematic illustration of this measuring system is plotted in Fig. 10. The measured results for the indexing accuracy corresponding to cases with and without the TBC system are revealed in Fig. 11. In Fig. 11, at the designed speed, we can see that the indexing accu- 446 \u00d5 Vol. 124, SEPTEMBER 2002 rom: http://mechanicaldesign.asmedigitalcollection.asme.org/ on 05/05/20 racy is substantially increased from 300 sec" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003987_j.1525-1594.1999.06315.x-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003987_j.1525-1594.1999.06315.x-Figure1-1.png", + "caption": "FIG. 1. The schematic drawing is of the electrohydraulic total artificial heart system. The system is composed of an intrathoracic pumping unit consisting of blood pump and energy converter subunits and an electronics unit consisting of an internal controller subunit, internal and external batteries, and transcutaneous energy transmission and telemetry subunits.", + "texts": [ + " Address correspondence and reprint requests to Dr. Eisuke Tatsumi, Department of Artificial Organs, National Cardiovascular Center, Research Institute, 5-7-1 Fujishiro-dai, Suita, Osaka 565-8565, Japan. 242 been evaluated in a series of animal implantations, and 1 animal survived successfully for over 10 days. This paper describes the results of these animal implantations together with the present status of development. The EHTAH system comprises an intrathoracic pumping unit and an electronics unit (Fig. 1). The electronics unit is composed of internal and external controller subunits, transcutaneous energy transfer and optical telemetry subunits, and internal and external battery subunits. A report of the present development status and a detailed description of the electronics unit have been published elsewhere (5\u20137). The intrathoracic pumping unit consists of a blood pump subunit and an energy converter subunit (Fig. 2). The blood pump subunit is composed of diaphragm type left and right blood pumps, left and right atrial cuffs, and aortic and pulmonary artery grafts" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001083_978-81-322-2638-3_72-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001083_978-81-322-2638-3_72-Figure2-1.png", + "caption": "Fig. 2 Fabricated prototype antenna. a Top view. b Back view [27]", + "texts": [ + " This paper also shows that approach of EBG structure to create tunable notch is better over CSRR. Here authors substantiate their own work presented in [27] with measured radiation patterns and group delay. Following Table 1 provides some useful information about the EBG structures. All the simulation and optimization of the proposed antenna has been done with the Ansoft HFSS 13. Configuration and geometry of suggested antenna is presented in Fig. 1. Fabricated prototype antenna is presented in Fig. 2. This antenna is designed on the FR-4 dielectric material, which has thickness of 1.6 mm, dielectric constant \u03b5r = 4.4 and loss tangent of 0.02. (a) CSRR Antenna Design (WiMAX Notched Band) A CSRR provides filtering characteristic so, we have used a CSRR slot on radiating patch to create notch in WiMAX band. Figure 1a shows the antenna with CSRR slot, and its length is approximately \u03bbg/2. Proposed length of circular split ring resonator can be intended from the Eqs. (1) and (2). Leq \u00bc 2p R1 g \u00f01\u00de fc \u00bc C 2 Leq ffiffiffiffiffiffiffiffiffi er \u00fe 1 2 q \u00f02\u00de where g is 5" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002177_tap.2007.915427-Figure13-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002177_tap.2007.915427-Figure13-1.png", + "caption": "Fig. 13. Geometry of the proposed CP quadruple L-probe circular patch antenna.", + "texts": [ + " The slight asymmetry observed in the H- and V-polarization patterns can be attributed to the asymmetrical feed orientation of the dual L-probe antenna configuration. These results reveal significant enhancements in the impedance and axial ratio bandwidths over the dual L-probe antenna presented in [12]. In terms of the common frequency coverage of , axial ratio 3 dB, and 3-dB gain (gain 5.53 dBi), the proposed CP antenna exhibits a measured CP bandwidth of 28.04% from 1.38 to 1.83 GHz. The geometry of the CP quadruple L-probe circular patch antenna is shown in Fig. 13. The quadruple L-probe antenna shares the same antenna parameters with the dual L-probe antenna shown in Fig. 7. The feed network, comprising a pair of the proposed 90 broadband baluns connected by a 180 transformer, was printed on the RO4003 substrate. To provide 180 phase shifting, the lengths of the microstrip branches must differ by , where refers to the guide wavelength at the center operating frequency, say, 1.8 GHz, in this work. The input transmission line is connected to the two microstrip branches by a quarter-wavelength transformer with characteristic impedance given by " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000084_6.1992-1082-Figure9-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000084_6.1992-1082-Figure9-1.png", + "caption": "Figure 9. Finite-Element Model", + "texts": [], + "surrounding_texts": [ + "A NASTRAN finite-element model was created from the vehicle geometry to study the effects of thermal loading. By varying the internal rib and spar arrangement between a rigidly fixed concept and the thermally compliant concept, the effects on the magnitude of the skin panel running loads could be asscssed. The model of the winglfusclagc structurc (Figurc 9) is apolymcrcompositc fuselage with an all-titanium wing. This combination of materials represents a compromise between the most extreme condition (aluminum1 polymer composite) and the least sensitive combination (polymcr cornpositelpolymer composite). 'd 31 -1 57 Both models were analyzed by applying thermal loads only. The temperaturc data were gencrated from a Mach 2.2 cruisc mission using adiabatic wall temperatures. Temperature data from 125 points were calculated and then interpolated over thc remainder of the aircraft. The skin panel thicknesses and internal substructure were sized based on strength requirements from previously run mechanical load cases. By proceeding in this manner, the running loads generated by the model wcrc strictly a function of material selection and structural gcomctry. Thc location of rcprcscntativc skin panels and the running loads gcneratcd arc shown in Figurc 10 and Table 1, respectively. 49 -1 75" + ] + }, + { + "image_filename": "designv6_24_0000777_smasis2010-3636-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000777_smasis2010-3636-Figure1-1.png", + "caption": "Figure 1: Proposed deployable ISR asset scheme12", + "texts": [], + "surrounding_texts": [ + "Morphing and reconfigurable aircraft in recent years have been the topic of much research due to their adaptability both from efficiency and capability perspectives. One such mission that would greatly benefit from reconfigurability is the high speed delivery of low speed intelligence, surveillance, and reconnaissance (ISR) air vehicles. The ideal ISR asset is one that does not require constant maintenance to maintain system readiness, such as keeping onboard batteries constantly charged while the vehicle is in storage, and has as small as possible delay between storage and deployment, such as waiting for batteries to charge once the need for the vehicle arises. Weighing these two objectives, taking advantage of the benefits of reconfigurability while maintaining system readiness, energy harvesting quickly seems like an optimum solution. Current ISR vehicles, namely those falling into the broad class of unmanned aerial vehicles (UAVs), are optimized for low speed/high endurance flight. As a result, these aircraft require long flight times to reach destinations that are significant distances from their take off points. Large ISR UAVs are deployed at high altitude servicing a large general area under standing orders with smaller ISR UAVs deployed by ground troops supporting more direct mission objectives. While the smaller UAVs deployed in the field serve to support missions with little advanced notice, they also suffer from limited capabilities due to size and weight restrictions associated with ground transportation. Large UAVs have significantly greater capabilities, however they are not very portable and thus must fly from either their current location or a central base to the desired target, adding time before the asset is in the desired position. To overcome these two challenges, a medium sized ISR UAV transported to the desired location and deployed by a high speed aircraft has been proposed. Such a vehicle would maintain the functionality afforded by size while having a low \u201ctime on target\u201d time. Such a design, however, requires that the vehicle be packaged in a way as to allow transportation by another vehicle and then deploy upon release from the transport vehicle. The most notable challenges associated with this are the folding of the aircraft wings. Deployment of folded wings, in turn, require purposely designed mechanical linkages and actuators and an energy source to drive them. To supply the energy required, direct energy harvesting is proposed. Above Mach 1.0, a significant amount of thermal energy is produced at the nose of the vehicle due to compression of the atmosphere. Since conversion of thermal energy to other forms, such as electric through the use of thermoelectric or pyroelectric materials, necessarily involves high losses due to conversion efficiencies of the materials, it is desirable to investigate a way to utilize the thermal energy directly, directing the thermal energy to desired locations within the vehicle rather than converting it to other forms. Presented are the results of ongoing efforts in the feasibility and design of such a system." + ] + }, + { + "image_filename": "designv6_24_0000535_0954406214544726-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000535_0954406214544726-Figure3-1.png", + "caption": "Figure 3. Goldak double ellipsoid heat source model.", + "texts": [ + " As the surface is exposed to the surrounding environment, the primary mode of heat transfer would be convection after neglecting the effect of radiation, and therefore, the boundary condition21 on the surface can be written as K T\u00f0 \u00de~rT:~n \u00bc h\u00f0Ts T0\u00de \u00f02\u00de where ~n is the normal vector of the surface, Ts is the surface temperature, T0 is ambient temperature and h is the convective heat transfer coefficient. The net line energy is given by Q \u00bc UI = \u00f03\u00de where U is the voltage, I is the current, is the arc efficiency and is circumferential weld speed. Generally for modeling, the heat source configuration of a conventional double ellipsoidal heat source model suggested by Goldak et al.22 can be used for TIG welding. Goldak heat source model is defined spatially by a double ellipsoid as shown in Figure 3. The front half of the source is the quadrant of one ellipsoidal source, and the rear half is the quadrant of at Middle East Technical Univ on March 1, 2016pic.sagepub.comDownloaded from a second ellipsoidal source. The power density distribution is assumed to be Gaussian along the weld path, or the axis of the work piece. The relation between the new reference frame on the heat source and the coordinate fixed on the work piece is given by \u00bc z\u00fe v\u00f0 t\u00de \u00f04\u00de where v is the welding speed and is a lag time necessary to define the position of the heat source at time t\u00bc 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000284_s10846-005-0932-y-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000284_s10846-005-0932-y-Figure2-1.png", + "caption": "Figure 2. Calculation of mirror dimensions.", + "texts": [ + " In PSVS, to avoid probable shades of parts of complex images because of the improper size or of the location of mirrors and the appearance of ghost images, the calculation of mirrors dimensions is required. The equation providing these dimensions is the following: AC = AB + BC \u21d2 AC = ( OB \u00b7 tan a \u00b7 cos \u03c91 tan(\u03c91 \u2212 a) + OB \u00b7 tan a \u00b7 sin \u03c91 ) + OB \u00b7 tan a sin \u03c91 + cos \u03c91 \u00b7 tan a . (2) Angle 2\u03b1 represents the angular field of view of the lens while \u03c91 is the angle the optical axis forms with a mirror plane (Figure 2). OB is the path a light beam follows along the optical axis from the optical center of the real camera to a mirror. Then from Equation (2) the partial cases for \u03c91 = 45\u25e6 and \u03c91 = 90\u25e6 are derived. Proof. For the calculation of the dimension AC, the segments AB and BC are independently calculated (Figure 2). This way, the intersection of the optical axis with a mirror is determined. Knowing the intersection, a mirror can correctly mounted on the PSVS-base during the construction. The segment AB is calculated as follows: AB = AF + FB = cos \u03c91 \u00b7 tan a \u00b7 OB tan(\u03c91 \u2212 a) + sin \u03c91 \u00b7 tan a \u00b7 OB. (3) The segment BC is calculated as follows: OB = OD + DB = sin \u03c91 \u00b7 BC tan a + cos \u03c91\u00b7BC \u21d2 OB = BC \u00b7 sin \u03c91 + cos \u03c91\u00b7 \u00b7 tan a tan a \u21d2 BC = OB \u00b7 tan a sin \u03c91 + cos \u03c91\u00b7 \u00b7 tan a . (4) Adding the segments AB and BC Equation (2) is derived" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002808_cp.2010.0978-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002808_cp.2010.0978-Figure8-1.png", + "caption": "Figure 8: Electronic Commutator stack", + "texts": [ + " The current (leading the voltage) is still square wave in nature, but contains harmonics generated from network bridge and electronic commutator switching. The voltage waveform illustrates (motoring) armature reaction and a slight phase (delay) shift relative to the current waveform. The phase shift is necessary to provide volts for commutation. 6 Implementation Figure 7 below illustrates a direct drive Active Stator Permanent Magnet Generator (PMG) with the electronic commutator stacks mounted on the periphery of the stator. Figure 8 below illustrates the liquid cooled electronic commutator stack. The combination of air cooled machine and liquid cooled electronic commutator is ideal for direct drive PMGs. When power density is a key requirement, the machine can also be liquid cooled. [3] 1. Active Stator is an innovative variable speed drive topology that exploits and extends the benefits of DC machine technology, enabling significant power density improvements. Its ability to operate with permanent magnet rotors and in the longer term high temperature superconducting rotors makes it an ideal candidate for offshore wind turbines" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003331_9781119546924.ch7-Figure7.8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003331_9781119546924.ch7-Figure7.8-1.png", + "caption": "Figure 7.8 d-axis electrical circuit.", + "texts": [ + " The voltage equations, which involve time derivatives of the flux linkages, represent dynamic equations used for time simulations. The three d-axis voltage equations in reciprocal per-unit bases are ed = d\ud835\udf13d dt \u2212 \ud835\udf14\ud835\udf13q \u2212 Raid = \u2212(L\ud835\udcc1 + Lad) did dt + Lad difd dt + Lad dikd dt \u2212 \ud835\udf14\ud835\udf13q \u2212 Raid (7.145) efd = d\ud835\udf13 fd dt + Rfdifd = \u2212Lad did dt + (L\ud835\udcc1fd + Lad) difd dt + Lad dikd dt + Rfdifd (7.146) ekd = d\ud835\udf13kd dt + Rkdikd = \u2212Lad did dt + Lad difd dt + (L\ud835\udcc1kd + Lad) dikd dt + Rkdikd = 0 (7.147) Like the flux-linkage equations, these three per-unit equations can also be combined into a single d-axis equivalent voltage circuit as shown in Figure 7.8. The per-unitized reciprocal q-axis voltage equations are eq = d\ud835\udf13q dt + \ud835\udf14\ud835\udf13d \u2212 Raiq (7.148) ekq = d\ud835\udf13kq dt + Rkqikq = 0 (7.149) which can be combined into the q-axis circuit shown in Figure 7.9. The 0-axis voltage equation is given by e0 = d\ud835\udf130 dt \u2212 Rai0 (7.150) which can be modeled by a separate circuit, also shown in Figure 7.9. The stator flux and voltage equations in the reciprocal, equal mutual-reactance per-unit system are summarized below, with t in radians \ud835\udf13d = \u2212Ldid + Lafdifd + Lakdikd (7" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000008_icelmach.2018.8506886-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000008_icelmach.2018.8506886-Figure5-1.png", + "caption": "Fig. 5. 3D calculation model with material parameters", + "texts": [ + " In order to conduct a thermal analysis, based on the finite element method, a simplified three-dimensional model of the stator of the engine was developed (Fig.4). The model has been prepared in such a way as to simplify the geometry that does not affect the efficiency of the cooling system and the thermal state of the machine. The applied model includes: an aluminium support element with a water jacket (1), a simplified stator core (2), a simplified winding model (3), a thermally conductive resin filling the space between the winding and the supporting structure (4). In the CFD analysis program, the model (shown in Fig.5) was additionally supplemented with a cooling medium in the water jacket channels. The thermal resistance substitute parameters have also been assumed: Rs - thermal resistance corresponding to the pressure between the core and the water jacket construction, R\u017c - the thermal resistance corresponding to the groove insulation. Then the model was discretized. The discreet model is presented in Fig.6. All models and calculations were performed in Autodesk Inventor and Autodesk Simulation CFD software" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002239_amm.103.417-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002239_amm.103.417-Figure5-1.png", + "caption": "Fig. 5 Shaft Sensor Detection Circuit", + "texts": [ + " Three of the relays are used for the multi velocity signal output, 23 combinations signal can output. Meanwhile, the relay output status through the opto-coupler feedback to STC controller to judge, and software interlock to prevent relay malfunction [5]. Shaft sensor is an electrical bridge sensor, the output voltage signal, the average voltage ratio about 1mV/1V . Input voltage generally 5~10V, so it is necessary to meet the load of 0 ~ 2t measurement accuracy. In this paper we choose AD's high-precision differential amplifier AD620. The gain range is 0 to 1000, only adjust R3 in Fig.5, the voltage follower OP07 provides a reference voltage to AD620. Using serial communication speech decode chip to complete real-time broadcast voice messages. Voice chip uses PM66, which is belonging to CY Century Technology Companie [6]. This chip is an integrated audio recording circuit, flash memory, ADPCM encoder and decoder, amplifier, voltage regulator circuit, and others fully functional. This paper designs a elevator Automatic leveling system, by using STC controller based on fuzzy logic control" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003263_1.4031902-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003263_1.4031902-Figure3-1.png", + "caption": "Fig. 3 von Mises equivalent stress contours of seven different sheaves: (a) double web, (b) double web with holes, (c) straight web, (d) straight web with holes, (e) thin web with stiffeners, (f) thin web with stiffeners and holes, and (g) web with decreasing thickness", + "texts": [ + " However, explicit formulas proposed in the literature [11,12] were applied to two proposed sheave geometries, such as straight web and web with decreasing thickness and compared some of the results with finite element (FE) analysis results as shown in the end of this section. The FEM-employed ANSYS is used to analyze the half portion of the sheave geometry by considering symmetricity of the loadings, geometry, and boundary conditions [13,14]. The design pressure on the grove which is calculated in Sec. 2.1 is applied and the design side load calculated in Sec. 2.2 is also applied. The obtained von Mises equivalent stress contours for all seven design geometries are shown in Fig. 3. The obtained values of maximum von Mises equivalent stresses are shown in Table 1. The calculated values of the utilization ratios are shown in Table 2. The utilization ratios for buckling capacities were obtained by the proposed guidelines in Sec. 2.4. The first modes of buckling were obtained by FEM-employed ANSYS tool as shown in Fig. 4 and corresponding buckling load multipliers a are shown in Table 1. While increasing the out-of-plane side load, the stresses at the web to hub joint versus applied out-of-plane moments were plotted" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001151_antem.2004.7860707-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001151_antem.2004.7860707-Figure3-1.png", + "caption": "Fig 3 Ka-band tracking/comms feed", + "texts": [ + " The feed horn was designed for axially symmetric, low side lobe pattern (TE11 mode) and good return loss for all three modes. In the final design step all three components (horn, TE21-coupler and TE11/ TM01 coupler) have been combined and fineoptimized for optimum performance. The entire feed chain was designed using advanced modelling tools (Mode Matching and FEM based) in a single iteration without resorting to any experimental development. The simulated return loss performance of the feed chain is shown in Fig.2. Fig.2 . Tracking/Comms Feed Simulated Return Loss The Ka-band tracking/comms feed is shown in Fig.3 Examples of the tracking/comms feed measured performance are shown in Figs. 4-5. 28000 28500 29000 29500 30000 frequency [MHz] -50 -45 -40 -35 -30 -25 -20 -15 -10 -5 0 R et ur n Lo ss [d B ] Fig.4 . Tracking/Comms Feed Measured Return Loss TE11 Port TE21 Port TM01 Port TE11 TE21 TM01 As Table 3 demonstrates, very good RF performances of the Ka-band comms/tracking feed have been achieved overall. Measured return loss of better than 25dB allows to maintain good impedance match and low amplitude ripple of the entire tracking network over required temperature range" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003378_bf02765177-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003378_bf02765177-Figure3-1.png", + "caption": "Fig. 3. General view of ShchK6 shield unit: 1) Ceiling; 2) brace; 3) footing; 4) body support; 5) hydraulic cylinder for shift; 6)slider; 7) guide; 8)suspensionmechani.~m; 9) scraper-conveyor.", + "texts": [ + " The geometric parameters of the ShchK6 support were chosen within the limits of rational values: ll/L = 0.4; c/L = 0.37; ~ = 72 \u00b0, -r = 38 \u00b0. With these values, the support can be shifted either as a ~kick ~ or with separate movements of the footings and ceilings. For nkick n movement, the following parameters are preferred: ll/L = 0.3-0.35; ~5 = 60-65\u00b0; 3' = 45 \u00b0. This means that the parameters we have selected are more suitable for separate movement of footings and ceilings. The ShchK6 shield unit (Fig. 3) consists of a powered shield support of the guard type, a scraper-conveyor for breaking and delivering coal to the coal discharge bord, suspension mechanisms to tie the rollers of the scraper-conveyor to sections of the support, hydraulic equipment to control the scraper-conveyor and move the support, electrical equipment, and a dust suppression system. The support consists of sections that are kinematically connected and that move frontally along the dip of the bed; these are interconnected along the strike by rigid and flexible connections along the floor and roof, respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002323_antem.2012.6262422-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002323_antem.2012.6262422-Figure2-1.png", + "caption": "Figure 2. Patch antenna power radiation pattern at 1.71, 1.905 and 2.05 GHz", + "texts": [], + "surrounding_texts": [ + "Keywords- statistical modeling ; antennas ; radio channel\nI. INTRODUCTION With the development of new generation wireless telecommunication standards and services, the contribution of antennas to the global system performance is getting more and more difficult to assess. Indeed, future standards such as LTE will put access points (AP) closer to the user, in a street or even within premises (femtocells) [1-3], where they will be far from ideal conditions. Unfortunately, antennas are commonly designed, simulated and measured isolated from anything else as much as possible, which is not the real situation for most cases where they are often deployed in proximity to various disturbers. As an example, if the antenna is placed close to a wall, the interaction between both of them will dramatically affect the antenna characteristics and e.g. the coverage, especially in the case of an omnidirectional radiator. Although many works have been done to evaluate the effect of e.g. a human body on antennas [4-6], the same effort has not been done for wall interaction, in spite of its relevance.\nIn this context, we focus on the concept of ideal antenna sector, then define three kinds of performance indicators relevant for quantifying the deviation between true antenna characteristics and the ideal sector. However, since the objects in proximity of a sensor are highly variable, a statistical approach is needed. An exhaustive study is of course impossible, whether experimentally or from deterministic simulation tools and a trade-off between complexity and accuracy is needed, allowing to condense this complexity into a small number of statistical parameters, once the statistical distributions are known [7].\nThis approach is applied to the influence of a wall on radiation patterns. Three antenna-wall interaction scenarios (patch-wall perpendicular, patch-wall parallel and dipole-wall) are investigated through electromagnetic simulations and analyzed to see whether the normal distribution can be acceptable to fit the observed distributions,\nII. SECTOR DEFINITION AND PERFORMANCE PARAMETERS As regards the placement and performance of an antenna such as an access point, telecom operators are mostly concerned about whether it can cover a targeted area, where coverage generally means a received power over a certain threshold [8-9]. The orientation and tilt angle need be chosen adequately for omnidirectional as well as directional antennas. In the present paper, we address both cases and attempt to evaluate the impact brought by environmental perturbations to the radiated power, taking into account the considered \u201csector\u201d (whether it was a part or the totality of the azimuth circle). Ideally, a radiation sector is thus defined vs. an intended use, i.e. as a specification of the antenna. However, in practice real antennas already cannot perfectly obey the specification. The deviation with respect to ideality occurs both on the beam width (in azimuth and elevation) and on the radiated power profile with respect to the ideal profile (which is flat within the sector and null outside). A first aspect of antenna non ideality consequently would be to define performance criteria of an isolated antenna. However, this is not the purpose of the present work whose goal is to address the degradations brought by disturbers to the isolated antenna. Also, from the perspective of an antenna designer, the main beam of an antenna shall include all the interested directions that perfectly match an operator\u2019s sector requirements. For those reasons the sector is going to be defined from the characteristics of the isolated antenna as follows: it is the rectangle in angular coordinates given by the quarter power beam width of the isolated antenna (-6 dB from the maximum) in both E plane and H planes, completely defined by the elevation and azimuth angles min\u03b8 , max\u03b8 , min\u03c6 , max\u03c6 . In the case of an omnidirectional antenna, only the elevation beam width need be defined and the parameters below are adapted accordingly.\nIn practice, this definition is still not fully sufficient, since the radiation pattern is frequency dependent. Therefore, in the analysis conducted in sections III and IV, a unique frequency independent antenna sector is chosen, as the sector with the narrowest beam widths over the set of considered frequencies.\n978-1-4673-0292-0/12/$31.00 \u00a92012 IEEE", + "Given this sector definition, three performance indicators are proposed below in order to quantify the capability of an antenna to radiate optimally into the considered sector. Beforehand, some quantities need be defined:\nthe Sector Efficiency SEref is expressed in (1), where ),( \u03b8\u03c6refG is the power gain of the isolated (reference) antenna along \u03c6, \u03b8 azimuth and elevation angles respectively:\n\u222b \u222b\n\u222b \u222b\n\u03a9\n\u03a9 = \u03c0 \u03c0\n\u03c6\n\u03c6\n\u03b8\n\u03b8\n\u03c6\u03b8\n\u03c6\u03b8\n2\n0 0\n),(\n),( max\nmin\nmax\nmin\ndG\ndG SE\nref\nref\nref\n(1)\nThis quantity basically expresses that not all radiated power falls within the sector even for the isolated antenna, owing to the fact, from elementary principles, that a part of the radiation necessarily leaks out of the sector.\nIn the expressions below, ),( \u03b8\u03c6G is the power gain of the disturbed antenna. sectorS is the solid angle area of the sector.\nThe three performance indicators are the following:\n\u2022 The normalized in sector power gain (ISPG):\nrefref norm TESE\nGG \u22c5 = ),(),( \u03c6\u03b8\u03c6\u03b8 (2)\nwhere TEref stands for total efficiency of the isolated antenna.\n\u2022 The normalized in sector mean power gain (ISMPG) within the sector:\nsector\nmax\nmin\nmax\nmin\n),(\nS\ndGnorm\ngain\n\u222b \u222b \u03a9 =\n\u03c6\n\u03c6\n\u03b8\n\u03b8 \u03c6\u03b8 \u03bc (3)\nrefSE is a normalization factor. Owing to this factor and if we assume that the total, gain\u03bc can thus be understood as the mean power gain over the sector, normalized by the isolated antenna sector efficiency. The reason for incorporating this normalization factor is that the sector efficiency turns out to depend on the frequency, even for the isolated antenna. By normalizing, we retain in gain\u03bc only the effects due to the disturbances rather than the non idealities of the isolated antenna.\n\u2022 The in sector standard deviation of the power gains (ISSDPG) within the sector:\n( )( )\ntor\ngainnorm\nS\ndG\nsec\n2\ngain\nmax\nmin\nmax\nmin ,\u222b \u222b \u03a9\u2212 =\n\u03c6\n\u03c6\n\u03b8\n\u03b8 \u03bc\u03c6\u03b8 \u03c3 (4)\nBasically, the 3 parameters tell us if the radiated power is constant and sufficiently high within the sector, with respect to the isolated antenna. ISPG contains all values of the radiated power in all directions within the sector, aggregated for all disturbers, thus it provides us with very detailed information about the antenna radiation characteristic. ISMPG tells us what ISPG is in average for a given disturber. ISSDPG tells up to what extent the gain deviates from this mean value in all directions within the sector, in other words whether the radiation pattern is flat or non uniform within the sector. In case the statistical distribution of the power gain within the sector turns out to be normally distributed, these two parameters are sufficient to describe the distribution completely.\nIII. DIRECTIONAL ANTENNA-WALL INTERACTION In the first set of simulations, we designed an air patch antenna operating from 1.7 GHz to 2.1 GHz in proximity to a wall whose characteristics are shown in fig. 1-3. Given that the power radiation pattern is getting narrow at low frequency, the sector at 1.7 GHz is selected to cover the angles \u03c6 = [30-150], \u03b8 = [55-150] in degrees according to the definition in section II.\nTwo cases were considered, with either the antenna\u2019s ground plane perpendicular or parallel to the wall (as shown in fig.4). The electromagnetic simulations have all been carried out using CST [10] at three frequencies (f = {1.71, 1.905, 2.05} GHz). In order to simulate a generic scenario for the antennas", + "in proximity to a wall, three parameters (relative permittivity of the wall, distance between the patch antenna and the wall, wall thickness), as summarized in table.1, were taken into account as random variables for our statistical study. Overall, the number of realizations was thus 27, multiplied by 3 frequencies. In all cases, neither the antenna itself nor the materials constituting the absorbers are lossy in the simulations. Thus the considered effects are solely related to impedance mismatched or (mainly) to degradation of the radiation shape.\nFig. 5 and 6 reveal the strong impact of the nearby disturber on the patch antenna in both cases. It can be seen in fig. 5 that the main beam direction shifts away from the wall when the antenna gets closer to it. This phenomenon obviously occurs because the lower half of the original main lobe experiences a wall reflection and sums up with the other half, causing the divergence of the mean direction from the original one (90\u00b0). The shorter distance from the wall, the more power reflected by the wall and the larger reflection angle which deviates from the original main beam direction. In order to quantify this effect, we use the Mean Radiation Direction (MRD) proposed by Fleury [11], which is more robust than the maximum gain direction:\n\u222b \u03a9=\u03a9 dGe ),(),( \u03b8\u03c6\u03b8\u03c6\u03bc (5)\nwhere Te )]cos(),sin()sin(),sin()[cos(),( \u03b8\u03b8\u03c6\u03b8\u03c6\u03b8\u03c6 = . From this mean vector, we can extract the MRD as the direction of \u03a9\u03bc .\nFig. 7 highlights the distribution of the mean azimuth angle mean\u03c6 over the set of realizations. Notice that the difference of MRD between the disturbed antenna and the isolated antenna is" + ] + }, + { + "image_filename": "designv6_24_0001942_aim.2011.6027029-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001942_aim.2011.6027029-Figure4-1.png", + "caption": "Fig. 4. Rayleigh-Kapp method", + "texts": [ + " PM + PG = Ur \u00b7 Ir + V \u00b7 Ig (8) Equation 8, where Ur, V , Ir and Ig are voltages and currents from DC sources (figure 3), is used to find necessary data to determine thermal parameters, but this method is not suited because it is very sensitive to mechanical noise due to coupling. B. Identification methods with blocked rotor Methods that use blocked rotor are: Rayleigh-Kapp [12] and Potier [12] methods. For both methods, two motors are coupled. In the first method, one source is necessary, and for the second method two sources are used. Equation 9 is used for calculating necessary data for the Rayleigh-Kapp method (figure 4) where Em and Eg are back EMF1 of motor and generator, respectively. Im and Ig are currents which exist in motor and generator. Equation 10 is used for the Potier method (figure 5). 1ElectroMotive Force Ur = Em +R \u00b7 Im = Eg \u2212R \u00b7 Ig (9) PM + PG = V \u00b7 Im (10) In the above mentioned methods, it is not practically possible to have identical motors, and as a result, a very small rotation will appear. Oscillation of voltage will occur, what is not desirable for calculating the circuits. The best solution is to block the rotor to get thermal resistances more easily" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000532_s10778-019-00952-4-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000532_s10778-019-00952-4-Figure1-1.png", + "caption": "Fig. 1", + "texts": [ + "com. Translated fromPrikladnayaMekhanika,Vol. 55,No. 2, pp. 126\u2013132,March\u2013April, 2019. Original article submitted February 28, 2017. DOI 10.1007/s10778-019-00952-4 constraints must be ensured. Otherwise, a part of the kinematical chain will become (once, twice, etc.) statically indeterminate, and the mobility of its contours will increase by the same number of units, i.e., will not perform its functions. 1. Problem Statement. Basic Equations. Let us consider, as an example, the system shown in Fig. 1 [2], with input 1 and output 3 shafts, controlled friction clutches A, B and brakes C and D. Differential I is directly connected to II (their ring gears 2 and 2 are one link), and differential II is twice connected with differential III (their arms 5 and sun gear 4 and ring gear 4 are combined), i.e., S 3, and the number of degrees of freedom W 2 3 3 3. Therefore, for the stationary operation of the gearbox in Fig. 1, it is necessary to include any two of the four (A, B, C, D) friction elements. Let us derive the equations of motion of the system according to Fig. 1, considering the slippage in the operating clutches (i. e., with three degrees of freedom) and characterized by the generalized coordinates 1 3 5 , , of links 1, 3, and 5 from the Lagrange equations of the second kind [1]: A Q , [ ] 1 3 5 T , Q Q Q Q T [ ] 1 3 5 , (1) where A is a 3 3 matrix; is a column matrix of generalized accelerations;Q is a column matrixof generalized forces; the index T denotes transposition. The coefficients of the matrix A are given by A J J u J u u J u u u 11 1 2 12 2 4 42 2 12 2 6 42 2 64 2 12 2 ' ' ' , A J u u J u u u J u u 12 2 12 12 2 4 42 2 12 12 2 6 42 2 64 2 1 1 ( ) ( ) ' ' ' ( )1 12 12 2 u u , A J u u u J u u u u u 13 4 42 42 12 6 42 64 42 64 12 1 1 ' ' ' ' ' ' ( ) ( ) , A A 21 12 , A J u u J J u u u J u 22 2 12 2 12 2 3 4 12 2 42 2 12 2 6 12 1 1 1 ( ) ( ) ( ' ) ' ' 2 42 2 64 2 12 2 u u u , A J u u u u J u u u 23 4 12 42 42 12 6 12 42 64 1 1 1 1 ( )( ) ( )( ' ' ' ' ) ' ' u u u 42 64 12 , A A 31 13 , A A 32 23 , A J u J J u u 33 4 42 2 5 6 42 64 2 1 1 ( ) ( ) ' ' ' ", + " At t t 3 , the brakeC is clutched, its maximum load rises linearly until t t 4 and remains constant afterwards. The maximum transmitted friction moments [Nm] are: M A 0 100 , M t t t t t t t t t t t B 0 1 1 2 1 1 2 2 100 0 100 100 0 , , ( ) , , , , M t t t t t t t t t t t C 0 3 3 4 3 3 4 4 0 150 150 ! , , , , , , M D 0 0 . 2. Geometric, Kinematic, and Inertial Parameters.Themoments of inertia of the links [kgm 2 ]: J 1 1 , J 2 2 , J 3 2 , J 4 1.5, J 5 2 , J 6 1.5. The masses and moments of inertia of the satellite gears are neglected. The angular velocities of links 2, 4, 6 of the mechanical system in Fig. 1 can be expressed in terms of the generalized angular velocities 1 , 3 , 5 [5]: ( ) 2 1 12 3 12 12 1 u u u , generalized forces Q Q Q 1 3 5 , , is presented in Fig. 3b. The angular accelerations of links 1, 2, 3 are shown in Fig. 4a. The parameter values:M M k A B0 0 0 40 40 , , 1.4; the curves are shown for momentsM M k A z A0 0 0 ,M M k B z B0 0 0 . Figure 4b shows the curves of accelerations for the following parameters: M M k A B0 0 0 4 4 10 , , . From the figures it follows that the dynamic processes in the gearbox are fairly calm, despite a significant increase in the forcesM M A B 0 0 , , andM C 0 pressing the discs in the friction clutches A B, , andC" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002010_jae-2008-999-Figure9-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002010_jae-2008-999-Figure9-1.png", + "caption": "Fig. 9. Contacting three points in which the maximum displacement is generated to the +z-direction.", + "texts": [ + " Therefore, the deformation of the rotor or the stator was ignorable that the rotor and the stator could be assumed as rigid bodies. This assumption also induced another assumption that the contact only occurred only at the three peak points, although there were over three wave peaks. The reason for only considering three points was that the point number for sustaining a body is only three points when the mechanical deformation at the contacting surface is not occurred. Therefore, the elliptical displacement was calculated at the three peak teeth as shown in Fig. 9. The peak tooth was the one of which normal displacement to the +z-direction is maximum. The displacement to the x-, y-, and z-direction was simulated at the contacting three peak nodes using 3D-FEM. The elliptical displacement was drawn according to divisional steps at one contacting point by: xi = |xcom| \u00d7 sin [ 2\u03c0 n \u00d7 i + rad (xcom) ] (80) yi = |ycom| \u00d7 sin [ 2\u03c0 n \u00d7 i + rad (ycom) ] (81) zi = |zcom| \u00d7 sin [ 2\u03c0 n \u00d7 i + rad (zcom) ] (82) where, xi, yi, zi: the value of x-, y-, z-directional displacement at the i\u2212th divisional step of an elliptical displacement xcom: the complex of a displacement to x-direction ycom: the complex of a displacement to y-direction zcom: the complex of a displacement to z-direction n: the total divided number of an elliptical motion i: from 0 to n rad(k): the radian of a complex k" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000528_mikon.2012.6233542-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000528_mikon.2012.6233542-Figure2-1.png", + "caption": "Fig. 2 The front and back view of an inertial sensor node", + "texts": [ + " If the values of the fitness function does not satisfied the required threshold, then binary kinematic transformation, 3 and 4 , is applied onto the three particles and the updated positions of those particles are obtained by going through the R n velocity space. Then the Maxwell transformation is applied to find another set of solution and the corresponding fitness function is evaluated. The procedure continues if the required threshold is not reached. III. THE PRINTED MONOPOLE WITH METALLIC GRID The reserved area for antenna on the proposed WSN- enabled inertial sensor shown in Fig. 2 is 20mm 10mm. The meandering printed monopole or inverted-F antennas can be applied to implement the required antenna operated at 2.4GHz. However, more compact antenna is desired to reduce the size of the whole sensor node to be 20mm 20mm or 30mm 15mm, which forces the antenna area to be 20mm 5mm or 10mm 15mm, respectively. As shown in Fig. 3, a candidate antenna is found in [8] and comprised of a main radiator component, including the bended printed monopole with coplanar ground planes, and the metallic grid" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000957_1.1559581-Figure15-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000957_1.1559581-Figure15-1.png", + "caption": "Fig. 15 Braking at the boundary of two surfaces", + "texts": [ + "9,ui,1 and Du.0 0,v i,0.1 and Dv,0 0.9,v i,1 and Dv.0 Then, if the cursor is near the boundary and approaches it, the cursor is given force that pulls it back to its previous position where this force was calculated last time. The force is represented by the following equation: F5k Pi212Pi t i2t i21 (16) where k is a scalar constant, t i is current time, Pi is the current position of the cursor, t i21 is the time when this force was calculated last, and Pi21 is the position of cursor at that time ~see Fig. 15!. This force prevents the cursor from moving fast to the boundary and jumping across the boundary. Hence, the user feels the sharp edge precisely. Although we give the cursor braking force in order to reduce its speed near the boundary, we can move the cursor constrained on the surface at the appropriate speed. Hence, we can recognize shapes of surfaces and discriminate the continuity of a boundary of surfaces. 3.5 Restraint Force of Movement. The fourth type of haptic navigation is used to let users feel how much a surface is modified through force" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001830_icamimia47173.2019.9223397-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001830_icamimia47173.2019.9223397-Figure1-1.png", + "caption": "Fig. 1: Profile of Touristant ASV.", + "texts": [ + " In this paper, a linear model with 3-DOF, resulting from the linearization of the 3- DOF nonlinear model (surge, sway and yaw) and under a disturbance in the form of external factors, such as the wind velocity and wave height. II. Au t o n o m o u s Su r f a c e Ve h i c l e ASV is equipped with some devices, namely GPS (Global Positioning Systems), sensors, gas, pH sensors, bluetooth, and telemetry. The vehicle will move automatically in real time after the location is determined. Besides its utilization for a research water vehicle, ASV can also be used as a survey vehicle, inspection of river conditions, seismic surveys, rescue operations etc. The profile and specification of Touristant ASV are listed in Fig. 1 and Table 1. In general, the water-vehicle motions are divided into two types, namely translational and rotational motions. Transla tional motion consists of surge, sway and heave. Rotational 978-1-7281-3090-3/19/$31.00 \u00a92019 IEEE 329 Authorized licensed use limited to: University College London. Downloaded on November 02,2020 at 04:01:45 UTC from IEEE Xplore. Restrictions apply. motion comprises of roll, pitch, and yaw [9]. This study used the equation of motion for water vehicle with 3 degrees of freedom, that is surge, sway and, yaw" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000686_vlsidcs47293.2020.9179860-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000686_vlsidcs47293.2020.9179860-Figure3-1.png", + "caption": "Fig. 3. Microstrip Radial Stub replaced by Conventional Open Stub", + "texts": [ + "2, displays the input and matching network with the lumped and distributed path analyzed by the Keysight ADS 2017 Electro Magnetic simulation software. Wider bandwidth can be achieved through Open circuit stub with lower impedance lines than Z0, So use of a microstrip radial stub which produce low impedance, it does not experience from a large, distributed T-junction. Due to lengthy input and output matching networks, it is proposed that the use of long transmission line should be replaced by microstrip radial stub as depicts in . Fig.3. the effect of the radial stub in isolation with respect to the replaced transmission line. Gate and Drain biases are done in binary ways. For gate bias, a 1.5nH inductor was joined to the gate transistor in series. In the case of drain bias, these paths must carry an colossal amount of currents that are not advisable for low inductor value because of their lower metal width. The solution is to replace this using a quarter-wave length transmission line along with a compact size inductor. furthermore, no Authorized licensed use limited to: Cornell University Library" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003032_j.ifacol.2017.08.930-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003032_j.ifacol.2017.08.930-Figure1-1.png", + "caption": "Fig. 1. The geometric model of a pinhole camera.", + "texts": [ + " (5) The joint velocities q\u0307 relate to the end-effector velocity Ve = [ve \u03c9e] T expressed in the end-effector frame, where ve and \u03c9e correspond to the translational and angular part respectively, by: Ve=Jq(q)q\u0307, (6) where Jq(q)\u2208R6\u00d7n is the robot end-effector Jacobian. Consider now a 3D vision system which follows the pinhole camera model, mounted at the robot end-effector. Let [Xc Yc Zc] T be the axes of the vision system frame {C} attached at the centre of the vision systemOc. The coordinates of the image frame {I} are given by [\u03c7 \u03c8] T with OI denoting the centre of the image, as depicted in Fig. 1. Notice that the Zc axis of the camera frame is perpendicular to the image plane transversing OI. Given a set of m fixed 3D points Pi=[xi yi zi] T , i=1,...,m expressed in the camera frame, the corresponding 2D image features si=[\u03c7i \u03c8i] T , i=1,....,m are given as follows: si=[\u03c7i \u03c8i] T =(\u03bb/zi)[xi yi] T , (7) with \u03bb being the focal length of the vision system. Notice that the coordinates of Pi with respect to the camera frame are given by: [xi yi zi] T =RT c (q)[Pi0\u2212Pc(q)], (8) where Rc(q), Pc(q) and Pi0 are the orientation of the camera frame {C}, the position of its origin Oc and the position of Pi with respect to the robot coordinate frame respectively", + " The dynamic model of the robot is described by M(q)q\u0308+C(q,q\u0307)q\u0307+ g(q)= \u03c4 , where M(q)\u2208Rn\u00d7n is the positive definite inertia matrix, C(q,q\u0307)q\u0307 \u2208 Rn denotes the centripetal and Coriolis force vector, g(q)\u2208Rn is the gravity force vector and \u03c4\u2208Rn is the joint input torque vector. The system behaviour can be written in terms of the state vector [ qT q\u0307T ]T\u2208R2n as follows: d dt [ q q\u0307 ] = [ q\u0307 M\u22121(q)[\u2212C(q,q\u0307)q\u0307\u2212g(q)+\u03c4 ] ] . (5) The joint velocities q\u0307 relate to the end-effector velocity Ve = [ve \u03c9e] T expressed in the end-effector frame, where ve and \u03c9e correspond to the translational and angular part respectively, by: Ve=Jq(q)q\u0307, (6) where Jq(q)\u2208R6\u00d7n is the robot end-effector Jacobian. Fig. 1. The geometric model of a pinhole camera. Consider now a 3D vision system which follows the pinhole camera model, mounted at the robot end-effector. Let [Xc Yc Zc] T be the axes of the vision system frame {C} attached at the centre of the vision systemOc. The coordinates of the image frame {I} are given by [\u03c7 \u03c8] T with OI denoting the centre of the image, as depicted in Fig. 1. Notice that the Zc axis of the camera frame is perpendicular to the image plane transversing OI. Given a set of m fixed 3D points Pi=[xi yi zi] T , i=1,...,m expressed in the camera frame, the corresponding 2D image features si=[\u03c7i \u03c8i] T , i=1,....,m are given as follows: si=[\u03c7i \u03c8i] T =(\u03bb/zi)[xi yi] T , (7) with \u03bb being the focal length of the vision system. Notice that the coordinates of Pi with respect to the camera frame are given by: [xi yi zi] T =RT c (q)[Pi0\u2212Pc(q)], (8) where Rc(q), Pc(q) and Pi0 are the orientation of the camera frame {C}, the position of its origin Oc and the position of Pi with respect to the robot coordinate frame respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002593_9781119971009.ch10-Figure10.13-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002593_9781119971009.ch10-Figure10.13-1.png", + "caption": "Figure 10.13 The Galileo spacecraft configuration, showing the position of the RTG sources (Courtesy of NASA/JPL/Caltech)", + "texts": [ + " They provide independence of distance from the Sun (deep-space missions are possible). 3. They can provide low power levels for long periods of time. 4. They are not susceptible to radiation damage in the Van Allen belts. 5. They are suitable for missions with long eclipse periods, for example, lunar landers. The disadvantages of RTG systems need also be considered, and include 1. They adversely affect the radiation environment of the satellite whilst in orbit. This will influence the spacecraft configuration significantly as may be seen from Figure 10.13, which shows the Galileo spacecraft. In this instance, the RTG needs to be deployed on a lengthy boom away from the main satellite bus. 2. Careful handling procedures are required during satellite integration owing to the radiation hazard posed by the radioactive source. 3. High temperature operation is required for efficient energy conversion. This impacts upon the thermal environment of the vehicle, and again on vehicle configuration. 4. RTGs are a source of interference for plasma diagnostic equipment that may be carried as part of the scientific objectives of the mission" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000234_icelmach.2016.7732762-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000234_icelmach.2016.7732762-Figure4-1.png", + "caption": "Fig. 4. Lamination of mcB: for 2 and 4 poles according to the PM magnetization direction", + "texts": [ + "5 kW vs 4 kW for the 4- and 6-pole motor, respectively) exhibits higher losses. Consequently, it is expected that the 2-pole motor exhibits a higher temperature. LSSM motor exhibits higher power density than the corresponding IM and it could be reasonable to increase the rated power of the LSSM. The measured temperature at steady state of the prototypes are reported for reference in Section III. B. mcB: for 2 and 4 poles mcB is suitable to be used in 2- and 4-pole machines. Its lamination is shown in Fig. 4 and its geometrical data are reported in Table III. The size of mcB is similar to the size of mcA, with the exception that the external diameter has been slightly increased (from 200 mm to 210 mm). Due to the higher PM flux of this rotor structure, the stator back iron flux density in the 2-pole machine would be higher than 1.8T . The mcB PM volume is the same of the mcA one. Fig. 5 shows the distribution of flux lines and the correspondent flux density at no load for the 2- and 4-pole machines" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001886_iet-map.2010.0020-Figure15-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001886_iet-map.2010.0020-Figure15-1.png", + "caption": "Fig. 15 Hybrid six-pole triple-mode filter consisting of a triplemode conductor-loaded cavity resonator coupled to a triple mode dielectric-loaded resonator", + "texts": [ + " As mentioned earlier, one of the advantages of the conductorloaded cavity resonator to dielectric-loaded cavity resonator is that in addition to lower cost and ease of manufacturing, it offers a significantly wider spurious-free window. However, the dielectric resonator has a considerably higher Q-factor per unit volume. A potential application of the proposed resonator is to realise a hybrid triple-mode filter where a triple-mode conductor-loaded cavity resonator is coupled to a triple-mode dielectric-loaded cavity resonator to realise a six-pole triple-mode filter. The schematic of the proposed hybrid triple-mode filter is shown in Fig. 15, and IET Microw. Antennas Propag., 2011, Vol. 5, Iss. 10, pp. 1136\u20131142 doi: 10.1049/iet-map.2010.0020 its operating and wideband frequency responses are shown in Figs. 16 and 17, respectively. Combining the conductorloaded and dielectric-loaded resonator results in a filter structure with a reasonable Q-factor that meets the requirement of many practical applications, whereas the presence of the conductor-loaded cavity resonator suppresses the undesired spurious modes of the dielectric resonator close to the operating band of the filter" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002584_1999-01-0782-Figure15-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002584_1999-01-0782-Figure15-1.png", + "caption": "Figure 15. Stamped Kuhl Wheel Center", + "texts": [ + " Spoke pairs as well as most of the hub area details are then formed in stages with the spoke made by wiping the blank edges toward what will ultimately be the outside of the wheel. Finally, the material between the spokes of each pair is removed and discarded. MATERIAL CONSIDERATIONS \u2013 The complexity of this design made further traditional analysis problematic at best, so our design iterations were carried out using intuition and finite element analysis. Typical of the wheels developed is the 4-spoke, 4 lug 5X14 design shown in Fig. 15 above. This wheel is intended to sustain a 1,588 Nm cornering load for a minimum of 200,000 cycles. Wheel stresses corresponding to this load are shown in Fig. 16. Peak stresses at the outer fiber are less than 315Mpa, a reasonable level for formed work-hardened steel. Some fairly high stresses are also evident at the base of the curved web between spoke pairs. Addition of shape features to the hub area in the vicinity of this web appears to reduce these stresses. Careful optimization of material thickness and details of spoke shape have resulted in designs with weight reduction of 10 to 15% percent based on the whole wheel weight, compared with stamped center designs for the same application" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001573_j.apm.2014.11.058-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001573_j.apm.2014.11.058-Figure2-1.png", + "caption": "Fig. 2. Division of a coaxial type magnetic gear mechanism.", + "texts": [ + " In this section, the 2-D equivalent magnetic circuit network method is further modified and employed to solve the magnetic flux density distribution within the inner and outer air gaps of the coaxial magnetic gear mechanism, shown in Fig. 1. For the 2-D magnetic circuit model, the coaxial magnetic gear mechanism is divided into a certain number of nodes and reconstructed by connecting each node with its adjacent nodes of magnetic permeance to form an equivalent magnetic circuit network. Without loss of generality, the whole magnetic field of the coaxial magnetic gear mechanism is analyzed. As can be seen in Fig. 2, the mechanism can be divided into seven layers in the radial direction due to different materials, including: the yoke of the low-speed outer rotor (layer 1), permanent magnets on the low-speed outer rotor (layer 2), the outer air gap (layer 3), the stationary frame and steel pole-pieces (layer 4), the inner air gap (layer 5), the permanent magnets on the high-speed inner rotor (layer 6) and the yoke of the high-speed inner rotor (layer 7), respectively. Then, the coaxial magnetic gear mechanism is divided into N parts in the circumferential direction, i.e. there are 7N nodes in this coaxial magnetic gear mechanism, as shown in Fig. 2. The accuracy of the analytical results is directly proportional to the number of nodes. Analogous to Ohm\u2019s Law in electricity, the 2-D equivalent magnetic circuit network of the whole structure of the coaxial magnetic gear mechanism is shown in Fig. 3. In the analysis, the geometric dimensions and characteristics of the magnetic materials of the magnetic gear mechanism should be specified in order to get the permeance and magnetomotive force between each node. Basically, the inverse of the reluctance R is the permeance P" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000393_mawe.200900548-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000393_mawe.200900548-Figure2-1.png", + "caption": "Fig. 2. The MZGS-100 fatigue stand setup, where: 1 \u2013 specimen, 2 \u2013 rotational head with a holder, 3 \u2013 base, 4 \u2013 holder, 5 \u2013 lever (effective length = 0.2 m), 6 \u2013 motor, 7 \u2013 hydraulic connector, 8 \u2013 rotating disk, 9 \u2013 flat springs, 10 \u2013 unbalanced mass, 11 \u2013 driving belt, 12 \u2013 spring actuator, 13 \u2013 spring.", + "texts": [ + " Coefficients of the Ramberg-Osgood equation describing the cyclic strain curve under tension-compression with R = \u20131 for 10HNAP steel are the following: the cyclic strength coefficient K9 = 832 MPa, the cyclic strain hardening exponent n9 = 0.133. The 10HNAP cyclically stable material. The critical value of the integral for 10HNAP steel is JIc = 0.178 MPa N m [4]. In the fatigue tests, a MZGS-100 device was used at the Department of Mechanics and Machine Design in Opole (Opole University of Technology) [8], Fig. 2. All fatigue tests were performed under loading control and the load ratio R = Mmin / Mmax = \u20131, \u2013 0.5 and 0. Unilaterally restrained specimens were subjected to cyclic bending with the amplitude of moment Ma = (12.17, 15.64, 18.91, Opole University of Technology, Faculty of Mechanical Engineering, ul. Mikolajczyka 5, 45-271 Opole, Poland Correspondence author: D. Rozumek, Opole University of Technology, Faculty of Mechanical Engineering, ul. Mikolajczyka 5, 45-271 Opole, Poland E-mail: d.rozumek@po" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000011_095765005x31108-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000011_095765005x31108-Figure1-1.png", + "caption": "Fig. 1 Centrifugal blood pump: (a) Cordier diagram; (b) impeller; and (c) volute casing", + "texts": [ + " In this study, a centrifugal blood pump is designed to produce the volume flow rate of 11 l/min against the static pressure rise 71 mmHg of blood (typical blood density 1045 kg/m3 and viscosity 0.00345 Pa s) with the impeller rotational speed of 3610 r/min, i.e. the blood pump system satisfies the performance of the head coefficient (c) 0.070 at the flow coefficient (f) 0.018 as a centrifugal pump with the specific-speed (Ns) 0.97, and the specificdiameter (Ds) 3.85 lying on the collected field data [4] of efficient industrial turbomachines (Fig. 1(a)). Some representative design factors to be considered first in carrying out detailed design optimization can be briefly stated as follows. The blade angles along the streamwise direction should be properly distributed such that the static pressure blade loadings along the blade surface are optimized to minimize the interior stagnant or recirculation areas, and the inlet flow angle has to be designed to improve the pump efficiency and the pump performance at the off-design conditions. Regarding the volute casing, the gradually increasing circular cross-section from the cutwater to the throat is adopted to reduce the losses due to friction and the impact of the fluid exiting from the impeller, and the volute tongue (or cutwater) should be aligned into the general direction of the flow leaving the impeller", + " 3(f) that the volute tongue has been properly aligned into the general direction of the flow (absolute velocity) leaving the impeller. Table 1 Specifications of a centrifugal blood pump impeller Inlet tip diameter (mm) 17.61 Inlet hub diameter (mm) 11.90 Exit diameter (mm) 30.00 Exit width (mm) 2.70 Number of blades 5 Tip clearance (mm) 0.30 Length in axial direction (mm) 5.70 Blade angle at inlet tip (8) 75.80 Blade angle at inlet hub (8) 70.55 Blade angle at discharge (8) 67.92 Unshrouded-impeller of which blade angles are measured from meridional plane. Fig. 1 Continued JPE126 # IMechE 2005 Proc. IMechE Vol. 219 Part A: J. Power and Energy the impeller; (b) static pressure distribution (blade loading) around the impeller blade surface; (c) relative velocity vector distributions in the impeller; (d) relative velocity vectors at the impeller leading edge; (e) relative velocity vectors at the impeller trailing edge; (f) absolute velocity vector distributions in the volute casing; and (g) overall static pressure distribution in a centrifugal blood pump Proc" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003391_ijvnv.2012.046175-Figure10-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003391_ijvnv.2012.046175-Figure10-1.png", + "caption": "Figure 10 Sole reaction forces on the feet and ZMP (see online version for colours)", + "texts": [ + " To maintain a stable posture during walking, the ZMP position is controlled using the force sensors located under the legs as follows; ( ) ( )( ) ( ) 4 1 1 3 2 4 2 2 4 1 3 1 r r l l r r l l r l m i i i x x F F F F x F F F F F F = = + \u2212 \u2212 + + \u2212 \u2212 +\u2211 , (31) ( ) ( )( ) ( ) 4 1 3 4 3 4 2 1 2 1 2 1 r r l l r r l l r l m i i i y y F F F F y F F F F F F = = + + + + + + + +\u2211 (32) where xm and ym are the coordinates of the ZMP position at which the reaction torques are null, r iF and l iF are the left and right leg ground reaction forces measured by the force sensor, respectively. To make the ZMP position closer to a desired value yd (Figure 10) we use a PD controller that gets input ym of equation (32), and feeds back signal \u03b8y zmp(t) through an integrator to the pitching joint of the ankle. It is important to notice here that the time constant of the integrator is set relatively large so that we have much slower response of the ZMP controller than that of the gyro feedback. Also, in the proposed approach, the control of the ZMP requires not the robot model to be accurate. As a result, equation (29) can be extended to; ( ) ( ) ( ) ( ) ( ) ( ) 2 ( ) ( ) ( ) ( ) ( ) ( ) p l y am l fb s zmp km l p l y l fb s zmphm t t t t t t t t t t t t \u03b8 \u03b8 \u03b8 \u03b8 \u03b8 \u03b8 \u03b8 \u03b8 \u03b8 \u03b8 \u03b8 \u03b8 \u23a7 = + + + \u23aa\u23aa = \u2212\u23a8 \u23aa = + \u2212 \u2212\u23aa\u23a9 , (33) The input of the PD controller compensating the ZMP deviation is given by; ( )1 0 ( ) ( ') ', t zmp m de t k y t y dt= \u2212\u222b (34) where yd is the desired ZMP frontal position, k1 is the gain of the integrator that has to be set small enough such that there is no oscillation of the upper body" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002744_1.5122082-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002744_1.5122082-Figure3-1.png", + "caption": "FIGURE 3. The meridional shape and shape of the scapula apparatus 2D low-flow low-pressure stage", + "texts": [], + "surrounding_texts": [ + "The design of the centrifugal stage in the Universal Modeling Method is based on the choice of dimensions,\nbased on design constraints, similarity criteria for compressibility u k,M , flow coefficients\u03a6des and loading parameter:\n Tdes u2 2 des c / u (2)\nwhere \u0422 is loading factor.\nOne of the features of the primary design of impellers in the Universal Modeling Method is that the dimensions\nof the input to the impeller are determined on the basis of minimizing the relative input velocity:\n2 1 3\n2 3 0\n1 1\n2 '\n \n \ndes F\nD shaft\nD\n\u0424 K D A D\nK (3)\nwhere 1\n1 1\n1 sin \n \n \nb\nb\nz K\nD ( blade blockage factor), 1\n0\nD\nD K\nD ,\n2 2\n00 1 1 14\n shaft\nF\nD DF K\nF b D (ratio of area at blade\ninlet and an impeller inlet), 1 0 DD D K (blade inlet diameter), DA is ratio of an impeller inlet diameter and diameter that corresponds to minimum of inlet velocity.\nAt 1DA relative speed is minimal. The loss of efficiency in the IMP is minimal, if the loss factor is minimal:\n1 2\n \n\nimp\nimp\nT\nw (4)\nwhere imp is loss coefficient of an impeller, 1w is relative velocity at an impeller inlet.\nDue to the growth of friction losses in narrow channels of the flowing part, as well as loss of friction of the outer\nsurfaces of discs and leaks in labyrinth seals, the efficiency of steps with des 0.040 cannot be extremely high.\nRecommendations for calculating the dimensions of the input in the 2D IMP [7]: 1DA , 0,9FK , 1.02DK \nUnder these conditions, a low-flow low-pressure stage with design parameters des = 0.015, T des = 0.40,\nstructural constraints shaftD = 0.40, bimp = 0.012 (width of impeller blade), similarity criteria, 1.4k , uM = 0.65\n(impeller Mach number) has such a meridional shape and shape of the scapula.\n030032-2", + "When the speed 1w is minimized, the maximum possible blade height is obtained, but the scapular angles are small, the interscapular channels are long. Since friction losses are controlled by the ratio of the length of the channel to the hydraulic diameter [8], it is unclear whether such primary design is optimal.\nAn alternative approach is demonstrated by the flowing part of the low-cost impellers of the model stages of Clark (USA). The figure shows the meridional shape and shape of the RDC blading units of the low-flow highpressure stage of the model stage XXX3-Q from Clark.\nLow-flow high-pressure stage XXX3-Q has the parameters des = 0.015, T des , = 0.685, design constraints\nshaftD = 0.331 (shaft diameter of an impeller), b imp = 0.012. The characteristic dimensions of the 2D IMP XXX3-Q:\n1D = 0.481, 1b = 0.025, 2b = 0.025, impz = 15 (number blades), 1b l = 390, 2bl = 450. At the stage XXX3-Q, the\nblade height at the inlet is sharply reduced and the entrance angle of the blades is increased. Figure 4 shows the characteristics of the XXX3-Q stage according to the Clark firm data and the result of their simulation using the 8th version of the Universal Modeling Method model.\nOf the Universal Modeling Model (IDENT) program. uM = 0.463. The stage is tested at uM = 0.463 and 0.785. The characteristics of stage XXX3-Q at uM = 0.785 are simulated in a manner similar to those shown in Fig. 4. A comparison of the principles of the primary design of the Universal Modeling Method and Clark's firm is done with\n030032-3", + "the example of the stage XXX3-Q and the stage 2DI-0015-070-035 with the parameters des = 0.015, T des = 0.70,\nCharacteristic sizes of 2D IMP-0015-070-03: 1D = 0.451, 1b = 0.0447, 2b = 0.028 (width of channel), impz = 17, 1b = 20.50, 2b =470. Optimization of the shape of the blade grate on the analysis of velocity diagrams of an\ninviscid quasi-three-dimensional flow. Fig. 6 compares the speed diagrams of the impeller XXX3-Q and 2D IMP0015-070-035.\nFigure 7 shows the speed diagram of 2D IMP XXX3-Q at a flow rate ni =0.021. The ratio ni / des = 1.4 is a very large discrepancy. 2D IMP XXX3-Q has a much higher speed level and a very large blade load. In [9], the ratio of the average load to the average speed is recommended to be limited /mid midw w 0.45. In 2D IMP XXX3-Q, the load parameter is approximately /mid midw w 0.75. At the same time, the shape of the interlace channels of 2D IMP XXX3-Q is preferable to that of 2D IMP-0015-070-035 - both visually and in relation to the length to the hydraulic diameter. Fig. 8 compares the characteristics of the stage 2D IMP XXX3-Q and 2D IMP-0015-070-035 with stator elements of the stage XXX3-Q.\n030032-4" + ] + }, + { + "image_filename": "designv6_24_0002736_s11044-017-9578-3-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002736_s11044-017-9578-3-Figure3-1.png", + "caption": "Fig. 3 The TezGoz spherical parallel manipulator", + "texts": [ + " This property enables us to identify the base inertial parameters via the least square method. In the following section, this method will be used to determine and identify the base inertial parameters of an orthogonal 2-DOF SPM. In this section, an orthogonal 2-DOF SPM is considered. After a brief description of the manipulator, first, the proposed dynamic model is validated by using a multibody simulation software. Then, based on the method given in the previous section, the base inertial parameters are determined and identified. Figure 3 shows a 2-DOF SPM manipulator, named TezGoz, which is developed in the Human and Robot Interaction Laboratory at the University of Tehran. This mechanism is, in fact, an orthogonal structure of the general mechanisms introduced in Sect. 2 in which the adjoining revolute joints are perpendicular to each other. In addition, link-1 is connected to the end-effector via two revolute joints. The additional revolute joint, is intended to improve the rigidity and stiffness of the manipulator. It goes without saying that, since the aforementioned revolute joints have the same axis of rotation, the extra revolute joint does not affect the kinematic and consequently the dynamic behavior of the manipulator" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003985_s10846-013-9874-y-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003985_s10846-013-9874-y-Figure3-1.png", + "caption": "Fig. 3 Schematic representation of micro manipulator coordinate system", + "texts": [ + " The value of travel distance of the prismatic joint is d3 from coordinate (x2, y2, z2) to (x3, y3, z3). The coordinate (x3, y3, z3) is the position of end effector. The assignment of coordinate system to each link is similar to robot arm configuration which is represented by homogenous transformation matrix. This is calculated through combination of homogenous rotation and translation matrices. The schematic diagram for forward kinematics of micro manipulator with coordinate transformation is shown in Fig. 3. If the rotations take place about a fixed frame, the equivalent homogenous transformation matrix (1T3) is straight forward because all rotations occur about axes of reference frame. The transformation matrix of each link with respect to joint is denoted by 0T1, 1T2 and 2T3. The total transformation matrix 0T3 can be written as 0T3 = 0T1 1T2 2T3 (1) or in term of joint variables 0T3 = \u23a1 \u23a2\u23a2\u23a3 \u2212C\u03b81S\u03b82 S\u03b81 C\u03b81C\u03b82 (d3C\u03b81C\u03b82 + a1C\u03b81) \u2212S\u03b81S\u03b82 \u2212C\u03b81 S\u03b81C\u03b82 (d3S\u03b81C\u03b82 + a1S\u03b81) C\u03b82 0 S\u03b82 d3S\u03b82 0 0 0 1 \u23a4 \u23a5\u23a5\u23a6 (2) The global end effector position can be written as \u23a1 \u23a2\u23a2\u23a3 x y z 1 \u23a4 \u23a5\u23a5\u23a6 = 0T3 \u23a1 \u23a2\u23a2\u23a3 0 0 0 1 \u23a4 \u23a5\u23a5\u23a6 (3) where x, y, z are the end effector position with respect to base frame" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003984_aim.2009.5229822-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003984_aim.2009.5229822-Figure5-1.png", + "caption": "Fig. 5. Utilization of cylindrical PMs for rotor poles", + "texts": [ + " The prototype can be employed for experimental measurement of magnetic field of the rotor and actuator torque output. The current stator size is much larger than that of the rotor, which can facilitate the experimental works inside the stator. In the theoretical study, dihedral-shaped PMs are utilized as the rotor poles. This type of PM pole can benefit the torque modeling as well as actuator design optimization. However, it is relatively complicated to be fabricated and the cost is high. Therefore, rotor poles with stacked cylindrical PMs are proposed as an alternative. As shown in Fig. 5(a), several cylindrical PMs of different sizes can be stacked together to form a rotor pole. Because the cylindrical PMs are widely available on the market, the cost of PM poles can be reduced by 80%. Furthermore, by using cylindrical-shaped magnet poles, the inertial moment of the rotor can be reduced by more than 60%, which is favorable for improving dynamic performance of the rotor. The back iron block is mounted at the rotor center to stick the all PM poles together and reduce the magnetic energy loss", + "1 times compare with the previous torque model. Hence, \u03b8 \u2032 p = 2.1\u03b8p. The rotor coordinates and the measurement coordinates have the relationship: \u03b8 = \u03c0/2\u2212\u03b8p. Therefore, \u03b8 \u2032 = \u03c0/2\u22122.1\u03b8p. For the ith coil, it is \u03b8 \u2032 i = \u03c0/2 \u2212 2.1\u03b8pi. Second, ratios among three torque components can be observed from experimental result. It is found that torque components in x- and ydirections need to be multiplied approximately by 2.4. Third, the magnitude of the actuator torque can also be modified according to experimental result. As shown in Fig. 5(b), the cylindrical PM pole is enclosed by a dihedral-shaped PM pole. By varying dimensions of the PM pole, the magnitude of torque output can be adjusted. Therefore, dimensions of PM pole can be chosen to make the magnitude of theoretical result coincident with that of experimental result. In this prototype construction, the PM-pole dimensions are determined as \u03b1 = 30\u25e6, \u03b2 = 30\u25e6, Rb = 15mm and Rr = 46.5mm. The actuator torque for single coil is thus Ti = [ Txi Tyi Tzi ] = Tc [ 2.4gx(\u03c0/2 \u2212 2.1\u03b8pi, \u03c6i) 2" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002267_iros.1996.569005-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002267_iros.1996.569005-Figure2-1.png", + "caption": "Figure 2: The result of recognition and poseestimation on both left (a) and right ( b ) images. The estimated 3-0 pose of the manipulation landmarks are reprojected as the doited ellrpses t o verafy the accuracy of pose-estimation.", + "texts": [ + " 1 Introduction The objective of this research is to develop a vision algorithm that provides sufficient information for a bin-picking robot to manipulate complex industrial objects. Typical stereo views of a workspace are shown in Fig. 1, where alternator covers are the target objects to be manipulated. In this paper, we present an algorithm that is capable of identifying such complex objects as well as estimating their 3-D pose by stereo vision technique. A sample result of the algorithm is shown in Fig. 2, where the pose-estimated manipulation landmarks (i.e. the large bearing holes and their adjacent four screw holes) are reprojected onto the original images. For many years now, researchers have proposed various techniques for vision-based bin-picking. While there are many possible rationales to explain this circumstance, particularly in industrial applications, the facts attributed to the difficulty of implementing such systems are: a) the complexity of industrial objects; b) the lighting reflections generated by the common metallic colors on industrial objects; and c) the cluttered nature of object placements that generate mutual occlusions", + " If the supporting features are optimized with the objective function case of FA < F,, over both image frames, then the 3-D position of the supporting features are estimated. The 3-D object center p and the normal vector n with respect to the world frame are also estimated. If no supported features are matched (i.e. FA > F,,,), substitutes for supporting features are generated by appropriately choosing four evenlyspaced points along the seed feature's ellipse boundary. This substitute is then used to compute the pose of the object (i.e. vectors p and n). The result of this module is shown in Fig. 2, where the accuracy of the pose estimation is shown by reprojecting the seed and supporting features onto the original images. The result of the 3-D pose estimation is then passed to the Manipulator Interface Module ( M I ) as in Fig. 3 where motion-path-plans are generated for grasping. 7 Experimental Results In our experiments, several types of alternator covers that share the common model shown in Fig. 4 were used as target objects. The objects were randomly cluttered, with possibility of being upside down" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002797_dscc2011-6167-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002797_dscc2011-6167-Figure3-1.png", + "caption": "Figure 3. MOTORS IN ACTUATION SYSTEM", + "texts": [ + ", left ventricle, via apex and eventually places the end effector with the prosthetic aortic valve in the target point. As stated earlier, this feeding mechanism itself is positioned on top of another robotic system which is basically responsible for holding feeding mechanism and compensating for the heart motion. More details about the current mechanism can be found in [9]. There are four brushless DC motors in actuation system corresponding to the previously mentioned catheter\u2019s degrees of freedom. Figure 3 depicts how these motors are arranged in the design. DC motor 1 is in charge of pulling the cables and controlling the bending angle in the cable-driven joint. Since the tension in cables has to be independent of the catheter\u2019s other degrees of freedom, this motor along with relevant force transmission systems such as pulleys are directly attached to the catheter and the whole box moves with it. Both rigid inner tube and flexible outer one are driven inside left ventricle by the thrust force produced by DC motor 2" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003713_freq.2011.044-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003713_freq.2011.044-Figure3-1.png", + "caption": "Figure 3. The geometry of a linear array with a point of rotation that does not coincide with the array origin.", + "texts": [ + "mV/, is defined, rt will stay constant for all measurements, while ut D sin t varies. Because of mechanical tolerances and uncertainties concerning the phase center of the array elements, this is almost impossible to achieve without additional efforts. However the problem can be dealt with in a way that the offset of the center of rotation from the array origin (\u0131y ,\u0131z), itself becomes an unknown parameter to be estimated [8]. In that way the distance rt becomes a function rt D rt. ; \u0131y ; \u0131z/. From Figure 3 and geometrical considerations it follows that rtD p .\u0131y cos C\u0131z sin /2C .r0C\u0131y sin \u0131z cos /2: (22) We performed a grid-search for the center of rotation by solving equation (20) for a given set of positions (\u0131y; \u0131z). The one with the smallest residuum is the estimated center position of rotation. When simplifications like the far-field or the narrowband assumption do not hold, the individual entries of S can be calculated for each channel and for each measurement respectively. One simply has to evaluate the distance rt for every channel individually from geometrical considerations and use equation (1) for the individual channels" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000219_estc.2006.280002-Figure18-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000219_estc.2006.280002-Figure18-1.png", + "caption": "Fig. 18 The layout of the wideband LNA and the multi-bandpass filter designed off-chip", + "texts": [ + " The schematic of the wide band LNA and the filter are shown in Fig. 15 and Fig. 16, respectively. This traveling wave LNA is also designed with 0.15ptm GaAs PHEMT. It works from DC to 13GHz. The layout of the onchip solution is shown in Fig. 17. The area of chip is 0.89mm2, respectively. The noise figure and S21 of this design are also shown in Fig. 19. The gains on the 2.4 GHz and 5.2 GHz channels are respectively 10.1 dB and 9.1 dB. Noise figures are respectively 6.3 dB and 4.4 dB. The layout of the off-chip solution is shown in Fig. 18. The areas of chip and packaging board are lmm2 and 23.lmm2, respectively. Noise figure and S21 of this design are shown in Fig. 19. The gains on the 2.4 GHz and 5.2 GHz channels are respectively 8.3 dB and 8.9 dB. The noise figures are respectively 5.6 dB and 4.8 dB. Due to the similar quality factor of on-chip and off-chip passives, FoM of the two solutions is very close to each other. The results of cost-performance analysis of both solutions are summarized in Tab. 9. With this design example, it is found that the cost of off-chip solution is even higher than that of on-chip solution, while the performances of both solutions are similar" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000203_032131-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000203_032131-Figure2-1.png", + "caption": "Figure 2. Structure design diagram.", + "texts": [ + " In order to provide firm support and handle the power during stair climbing created by the collision of the floor and the crank, the motors were assembled on a large fortified profile structure on the device body. It is important to ensure the balance of the design, so the motors were attached to the side of the body, putting the majority of the weights on the base motors and wheels. For creating the body angle self-adjusting system, I added four rotating shafts to enable the body to change its angle relative to the legs. As shown in Fig.2, the wheel is designed as compound structure, which have a normal wheel and two stair-climbing wheel. The controller hardware design consists of an Arduino Mega, two motor drivers, a Bluetooth module, and a JY-901 gyroscope module. The built system initially receive Bluetooth signal by a Bluetooth module and the Arduino Mega is able to send the corresponding command to the motor drivers which are able to control both the direction and the speed of the motors. Due to the combination of large voltage motors, the power was distributed parallelly by a large voltage battery using multiple motor drivers" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000954_mmce.22111-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000954_mmce.22111-Figure2-1.png", + "caption": "FIGURE 2 Illustration of dipole antenna on top of metasurface with bias source", + "texts": [ + " Section II presents the design of PDEBG metasurface and design approach for the dipole antenna. The simulated results of tunable PDEBG and radiation pattern obtained are presented with some measured results in Section III and conclusions for wired dipole antenna ferrite-based EBG are presented in Section IV. Among constitutive properties of ferrite material, permeability is dependent on applied dc Hbias. Design parameters are mentioned in Table 1. Applied dc Hbias ranges from 80 k to 110 k A/m. Dimensions of metallic patch unit cell of metasurface is 4 mm\u00d7 4 mm (px \u00d7 py) as shown in Figure 2. While ground has dimension of 5 mm \u00d75 mm (bx \u00d7 by). Thickness of substrate is 2.54 mm, its permittivity value is 12.5 and loss tangent is 0.135. The metasurface is made of 7 \u00d77 array of this unit cell. Teflon substrate of thickness 1 mm is placed above ferrite substrate. Figure 2 shows the dipole antenna on top of the metasurface, connected to the dc bias source. For simplicity, 4 \u00d74 array is shown in the figure. Overall dimension of teflon and ferrite substrate is 40 mm \u00d7 40 mm (L \u00d7 W). Unit cell for this design consists of primitive structure of metal patch on teflon substrate. Below teflon layer, ferrite slab is incorporated. This structure is simulated using periodic boundary conditions. The desired magnetic response is obtained separately for both the polarized wave" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000083_ir-03-2018-0042-Figure37-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000083_ir-03-2018-0042-Figure37-1.png", + "caption": "Figure 37 Structure for wearable vehicle joints and links", + "texts": [ + " It is clear that this posture could be uncomfortable for the wearer in the full-scale model. However, due to the mass of the rear wheels mechanism in the full-scale model (around 16 Kg), the back and forth oscillations generated from the ZMP controller will implicitly compensate for this COM shift with no need to tilt the human torso. The different links are designed to accommodate any humanoid robot from 70mm up to 110mm. 3mm aluminum alloy are used for different links which are connected to one another by using spacers with different heights as shown in Figure 37. The system\u2019s joints have the same structure of the main system which is supported in both sides by ball bearings. Couplings are used to connect adjacent joints. The walking pattern consists of seven phases. The walking pattern repeats itself after the seventh phase. The KTX-PC starts in phase 1, \u201cTwo Leg Stand\u201d where the right leg is in front and the left leg is behind. Both legs are on the ground and the Center of Mass (COM) is between the two legs. Phase 2, \u201cOne Leg Stand\u201d. In this phase, the ankle roll, rear wheels mechanism and torso generate a torque whichmoves the COM to the inside edge of the right leg" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000825_kem.334-335.477-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000825_kem.334-335.477-Figure2-1.png", + "caption": "Fig. 2 Flat and curved wing configurations.", + "texts": [], + "surrounding_texts": [ + "The use of advanced composite materials can give an attractive potential for reducing the structural weight of a modern aircraft. In the initial design phase of a laminated composite wing, it is necessary to know its deformation pattern and variation of aerodynamic load distribution due to its structural flexibility. The composite materials have been used widely especially for aerospace vehicle structures because of its weight saving advantages. It is well-known that the optimum design of wing can be achieved by the aeroelastic tailoring of composite wing structures [1-4]. This paper has focus on the weight optimization of composite flat and curved wings by designing the proper lamination lay-ups using the generic algorithm. The computational analysis system for dynamic-flutter optimization has been developed based on the coupled numerical method using finite element method, efficient flutter analysis method and micro genetic algorithm. In order to show weight saving efficiency, optimization results for composite flat and curved wing models are also compared with the base weight of isotropic wing models." + ] + }, + { + "image_filename": "designv6_24_0001044_j.proeng.2017.01.201-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001044_j.proeng.2017.01.201-Figure2-1.png", + "caption": "Fig. 2. Simplified three-dimensional physical mode of the frame. Fig. 3. Finite element model of the frame.", + "texts": [ + "7 mm in density, 1500mm in length, 500mm in width, with 5# steel channel used as standard parts to support at 520mm and 980mm..As shown in Fig. 1. At first in the functional module SKETCH in ABAQUS, a two-dimensional section was drawn for the frame and then in the functional module PART, it was stretched in different faces to form a three-dimensional physical model, only a half of which would be extracted for analysis and calculation based on symmetric frame structure and load. The simplified three-dimensional physical model of the frame is shown in Fig. 2. After incorporating with proper materials, section attributes and assembly parts, divide the built threedimensional physical model into grids. To meet calculation accuracy and reduce calculation amount, use the threedimensional physical unit C3D4 for sweeping grid division, and then with reduced integral calculation, divide the modelinto 19591 cells and 41440 nodes. Grid divisions are shown in Fig.3. The frame bears not only the engine, chassis and cargos, but also various forces and torques created during travel of the vehicle" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002755_tvlsi.2012.2227848-Figure11-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002755_tvlsi.2012.2227848-Figure11-1.png", + "caption": "Fig. 11. SOI FinFET 6T SRAM (111) configuration. (a) (FEOL+BEOL) and (b) FEOL only. Dielectric regions are not shown.", + "texts": [], + "surrounding_texts": [ + "Owing to the width quantization property, multigate FETs with large electrical widths need to have multiple fins. We synthesized multifin FinFETs using the bulk and SOI FinFETs generated earlier at the 22-nm/14-nm/10-nm nodes. They consisted of four fins each, with shared raised source/drain epiregions that are via-contacted and connected using metal-1, as shown in Fig. 8(a) and (b). We varied the fin pitch, FP, which is the distance between the centers of consecutive fins, and computed the parasitic (FEOL+BEOL) capacitances for each layout using the setup described in Fig. 7(b). From Fig. 9(a), we can see that the trends in CDRAIN,TOT are in stark contrast to the single-fin results in Section III. While moving from SOI to bulk FETs, there is a 11.5%, 10.8%, and 8.8% increase in CDRAIN,TOT for the 22-, 14-, and 10-nm nodes, respectively, which can be attributed to the shared drain-to-bulk fin capacitances in bulk FETs. However, in the case of CGATE,TOT [Fig. 9(b)], there is only a 2%\u20134% increase from SOI to bulk FETs. An increase in FP from 40 to 70-nm results in a 20%, 31%, and 36% increase in CDRAIN,TOT for the 22-, 14-, and 10-nm nodes, respectively, while CGATE,TOT increases by 16%, 26%, and 28%, respectively. These results suggest that gate-toepi-source/drain/metal-1 capacitances begin to dominate as FP increases or the technology node decreases, and they highlight the need to model the entire (FEOL+BEOL) structure." + ] + }, + { + "image_filename": "designv6_24_0003952_978-3-642-33457-3_7-Figure7.22-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003952_978-3-642-33457-3_7-Figure7.22-1.png", + "caption": "Fig. 7.22 Stress in the magnet as result of the sleeve fitting", + "texts": [ + " At the same time, carbon fibres are wound around a very thin glass fibre ring whose inner radius is for 95\u00b5m smaller then the outer radius of the magnet. Finally, the carbon/glass fibre ring is pressed onto the rotor. The rotor final structure was modeled in 3D using Ansys Workbench software. Compression at the magnet inner surface resulting only from the press-fit is much smaller than calculated from the 2D modelling. However, what concerns is a large stress concentration at the line close to boundary between the iron shaft and glass fibre ring (pointed by arrows at Fig. 7.22). Further, according to this model, if the rotor remained at room temperature, the contact between the magnet and iron would be lost beyond 180.000 rpm (Fig. 7.23). Maximum possible speed is increased, though, if the operating temperature rises. Maximum equivalent stress in the magnet at the maximum speed and temperature is at the outer magnet surface and amounts to 110 MPa (Fig. 7.24) which is still below the compression limit. Maximum tensile stress in carbon fibres\u20141147 MPa\u2014is in a very good agreement with results from 2D modelling" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001472_9781118354179.auto089-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001472_9781118354179.auto089-Figure8-1.png", + "caption": "Figure 8. Twin clutch designs for road vehicles. (a) Passenger car. (Reproduced with permission from ZF Friedrichshafen AG.) (b) Commercial vehicle. (Janic\u0301ijevic\u0301, Jankovic\u0301, and Todorovic\u0301, 1987. Reproduced by permission of Faculty of Mechanical Engineering, University of Belgrade.)", + "texts": [ + " This article is \u00a9 2014 John Wiley & Sons, Ltd. DOI: 10.1002/9781118354179.auto089 Also published in the Encyclopedia of Automotive Engineering (print edition) ISBN: 978-0-470-97402-5 not so common, but have still found application in a range of vehicles. The basic difference with these clutches is related to the number of output shafts\u2014some have only one output shaft and that is practically the only difference in comparison with single plate clutches. The others have two output shafts, allowing for additional functions. Figure 8a shows a modern twin plate clutch (by ZF Sachs) with diaphragm spring for a high performance car. Figure 8b (Janic\u0301ijevic\u0301, Jankovic\u0301, and Todorovic\u0301, 1987) shows an old design of a twin plate clutch with central coil spring for a large commercial vehicle (arrows indicate torque input and output). For the same maximum clutch torque capacity, such designs require lower normal (clamp) force, making the spring system lighter and cheaper. These designs also offer thermal advantages, in terms of lower heat fluxes (more friction plates with larger total friction surface) and higher thermal capacity (more pressure plates), but have a disadvantage in poorer heat dissipation characteristics (because of smaller free contact areas with surrounding air)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002153_amm.813-814.915-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002153_amm.813-814.915-Figure3-1.png", + "caption": "Fig 3: Von mises stress contour of curved) Fig 4: Total deformation contour 3- spoke design (Steel ) of curved 3- spoke design (Steel)", + "texts": [], + "surrounding_texts": [ + "Static Structural Analysis. A pressure of 6 MPa is applied on the face of rim and all DOF are fixed for the hub inner face as shown in Fig 2b. The response of the wheels in terms of Total deformation, von mises stress, structural stiffness and specific structural stiffness are calculated and compared the results for all 6 design with three materials discussed in previous section. Tabular Values Spoke type/Result type Inclined spokes Curved spokes Y-shape Spokes Deformation [mm] 0.4257 0.2956 0.2327 Von mises stress [Mpa] 574.76 384.85 791.4 Stiffness [N/M] 813.26 1171.3 1487.9 Mass[Kg] 9.3295 9.8311 10.831 Specific structural stiffness[N/M-kg] 87.171 119.14 137.38 stiffness of 3 spoke curved design made up of steel is better than others. \u2022 All 5 spokes designs produced the higher von mises stress than the 3 spokes designs. It can be observed that the von mises stress does not vary much based on the type of the material chosen for manufacturing. All three spoke designs lower von mises stress than the other designs. Modal Analysis Results. It is done to find out the natural frequencies of the structure. All degrees of freedom are fixed at the hub as shown in fig 2a and performed the modal analysis for all designs. The typical contours are shown in below section. Contours Tabular Values Type/Mode Number Inclined spokes Curved spokes Y- shape 3 5 3 5 3 5 1 126.7 178.9 134.6 196.5 118.2 166.1 2 132.7 179.2 141.9 195.6 125.2 166.4 3 168.5 245.9 172.2 257.1 173.0 269.9 4 222.9 248.5 230.7 262.8 207.5 283.9 5 225.0 249.3 234.7 267.9 270.1 327.0 6 233.1 342.2 276.9 380.5 270.4 328.2 Mass (Kg) 1.914 2.139 1.985 2.258 2.123 2.483 Observations. All 5-spoke designs have higher natural frequency than the 3-spoke designs. For three shapes of the spokes, curved spoke designs produced the higher natural frequencies. Designs produced the nearly same natural frequencies irrespective of the material. Even though aluminum and magnesium have less mass than the steel, they produced the nearly same frequencies." + ] + }, + { + "image_filename": "designv6_24_0000231_s00170-017-0346-6-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000231_s00170-017-0346-6-Figure4-1.png", + "caption": "Fig. 4 Circular approximation. a Profile description. b Reaction forces and their levers. c Die opening", + "texts": [ + " The deformed plate has a complex shape, which can be described by a piecewise function [10]. Nevertheless, in the first instance, the shape of a bent plate loaded by means of a large radius punch can be approximated by a circle and straight lines. Although the approximation does not allow representing the multi-breakage, since the minimum radius is equal to the punch radius, it provides a valuable insight into the influencing parameters of large radius bending. The geometrical elements of this approximation are presented in Fig. 4a. For the circular approximation, the plate is modeled to wrap closely around the tool; thus, the radius RA of curvature at the mid-plane of the plate is equal to the radius of the punch Rp plus half the thickness of the plate. From the geometrical description, the length of the straight part of the plate midplane is as follows: ld \u00bc w 2cos\u03b8 \u2212 Rd \u00fe t 2 \u00fe RA tan\u03b8 \u00f02\u00de la \u00bc \u03b8RA \u00f03\u00de where w (see Fig. 4c) can be evaluated in function of the die opening w0: w \u00bc w0 \u00fe 2Rd tan 45 \u2212 \u03b1d 4 \u00f04\u00de Figure 4b shows the normal and tangential reaction forces on the die shoulders and their levers from point C. Since the plate wraps closely around the punch, the radius of curvature and the moment are the same in points A and C. The flat top of the trapezoidal profile (see Fig. 3d) can represent this. From this assumption, the levers can be determined by the following: blev \u2261 ld \u00f05\u00de blev2 \u2261 t . 2 \u00f06\u00de The bending moment is specified as follows: MA \u00bc Fnblev \u00fe Ftblev2 \u00f07\u00de For a Coulomb friction approximation and a friction coefficient \u03bc, the relation between Ft and Fn is given by the following: Ft \u00bc \u03bcFn \u00f08\u00de Taking into consideration Eq", + " Additional tests were conducted to cover the variability between different samples and between test repetitions. The complete experimental data are presented in the Dataverse research data repository [24]. This repository provides an open access to the complete experimental data set used in the current contribution. In case of several repetitions, the average values of the measurements are shown. The following geometrical parameters were selected: & Dies with four different openings have been used: 80, 60, 50, and 40 mm. Table 1 presents the die parameters for the circular approximation (see Fig. 4). & Punches with radii of 10 and 20mm have been used for all the dies, a punch with a radius of 30 mm for the dies of 60 and 80 mm, and a punch with a radius of 40 mm with the die of 80 mm. & For every combination of the tooling, plate thickness, and material, plates have been bent up to four different punch strokes, which resulted in part angles of approximately 90\u00b0, 110\u00b0, 130\u00b0, and 150\u00b0 before springback takes place. However, for Weldox 1300 plates with thickness of 6 mm and the punch radius of 10 mm, bends were made only up to 130\u00b0 and 150\u00b0 due to cracking at the smallest angles", + " Solid lines represent the experimental data and dashed lines the circular approximation model. For bending force graphs, a dashed line represents the value of 0.3 for the friction coefficient and a dotdashed line the value of 0.1. Figures 33 and 34 show that the circular approximation model overestimates the values of the bending force. Therefore, this approximation can be used for the safe estimation of the bending force, since the prediction is conservative. Due to the circular approximation of the bent profile (see Fig. 4), for some cases, the bending force even decreases with an increasing bending angle (see the black dashed line in Fig. 34). However, even in these cases, the force value is overestimated. Figures 35 and 36 show the comparison for the bend allowance. The circular approximation model overestimates the bend allowance when compared to experiments. In contrast to the bending force and the bend allowance, the calculated values for the springback lack a consequent trend with respect to the experimental data (see Figs" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001044_j.proeng.2017.01.201-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001044_j.proeng.2017.01.201-Figure5-1.png", + "caption": "Fig. 5. Loads of the model and its boundary conditions. Fig. 6. Operating stress nephogram.", + "texts": [ + " Due to symmetric force application, only a half of the frame was analyzed in its static condition, with its loads distributed in the whole surface. Along the negative direction of Y axis, the force application was 5000N. In this paper the frame contact was considered as rigid contact. Thus during force application to parts, the contact points between the frame bottom supporting parts were set stationary and their constraint types were set as ENCASTRE(U1=U2=U3=UR1=UR2=UR3=0), and the constraint types of parts' sections were set as ZSYMM(U3=UR1=UR2=0). Loads and boundary conditions are shown in Fig.5. 3.1.2. Results shown as follows As shown in Fig. 6, most of stress concentrates at the supporting points of the frame, in which the frame gets its maximum stress value in the whole force application area. Unit bending rigidity of the frame is equivalent to EJ of the frame girder. If on the surface of the frame girder there is evenly distributed load F, then the maximum deflection of the girder is [8]: X 3 max 48J Fl y (2) Wherein: xJ \u2014\u2014 bending rigidity coefficient of the girder; E elastic modulus, in 2/ mN ; most of stress concentrates at the contract surface between the frame bottom and supporting parts, with the frame\u2019s maximum stress of 86" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000083_ir-03-2018-0042-Figure22-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000083_ir-03-2018-0042-Figure22-1.png", + "caption": "Figure 22 FEA of the wearable vehicle while operating walking after optimization technique", + "texts": [], + "surrounding_texts": [ + "The human walking gait data freely provided by the OpenSim model Gait2392 is used. OpenSim software is built for the musculoskeletal simulation of the human during various motion activities (Delp et al., 2007). The OpenSim gait is shown in Figure 24. This gait data describes a full step of a normally walking human starting, and ending with the toe-off phase of the left foot. This obtained motion data is exported into MSC ADAMS model of the human. Revolute joints are added between the human pilot body parts and the normal motion splines are applied to those joints. The CAD model of the wearable vehicle is exported also to MSC ADAMS. The next step is to merge the twomodels (the humanmodel and the wearable vehicle model). Multiple variations have been carried out on the model of the wearable vehicle with the human. After the two models are merged, the human model is interfaced to thewearable vehicle throughMSCADAMS contacts. The load carried by the wearable vehicle in the simulation is 55 Kg. To measure the zero moment point (ZMP) in walking mode, force sensors were attached to the wearable vehicle feet. All-terrains wearable vehicle B.M. Sayed et al. Industrial Robot: the international journal of robotics research and application Volume 46 \u00b7 Number 6 \u00b7 2019 \u00b7 740\u2013762" + ] + }, + { + "image_filename": "designv6_24_0003125_1.42775-Figure9-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003125_1.42775-Figure9-1.png", + "caption": "Fig. 9 Optimized design at Mach 0.85 and 35 .", + "texts": [], + "surrounding_texts": [ + "Sevenwing-body configurations were studied with varying sweep angles (35, 30, 25, 20, 15, 10, and 5 ). The Mach number was systematically reduced from 0.85 for the highest-sweep wing to 0.79 for the lowest-sweep wing, and the lift coefficient was simultaneously increased to maintain MCL roughly constant. Initial aerodynamic shape optimizations were performed on these configurations, subject to the constraint that thickness could not be reduced anywhere on the wing. The results of this study are summarized in Table 2. These optimizations show that theML=D of the different configurations are all within 1.8% of the maximum value, suggesting a relatively flat design space for aerodynamic performance. Moreover, taking M p L=D as a better approximation to range factor, this trend yields up to a 4% improvement, favoring the lower-swept wings. At the very least, these results show promise that reduced-wing-sweep designs are possible for efficient transonic cruise. To illustrate the effectiveness of the aerodynamic shape optimizations performed herein, Fig. 5 provides the evolution of the pressure distributions for the 10 -sweep wing. Pressure distributions of the seed wing are depicted by the design-00 dashedline curves. Also included are intermediate states for design cycles 10, 20, and 35 (the total number of design cycles was preset to 50). Pressure distributions of the final design are given by the solid-line curves of design-50. The seed wing exhibits very strong shocks over most of its span. During the evolution of this optimization, the shock system is monotonically reduced in strength, with the final design comprising only very weak shock waves. Note that the drag of the seedwing is 207.3 counts and that the drag of the final design is 157.3 counts; hence, the optimization reduced the wing drag by 50 counts: about 25% in this case. This represents a substantial enhancement in aerodynamic performance that would most likely never be realized by a cut-and-try approach. The wings obtained from the aerodynamic shape optimizations were then used in an aerodynamic\u2013structural optimization procedure; this coupledmethod also allows formodest planformvariations within limits set by the feasibility of morphing the mesh. The procedure is outlined in Fig. 3 and is a two-stage process. Following an aeroelastic simulation and an aerodynamic adjoint calculation, the structural elements are redesigned while holding the aerodynamic loads fixed. Once a structure with minimum weight is obtained that satisfies the stress constraints, the gradients for the airfoil points and the planform variables are determined to find the design that leads to an aerodynamic and structural performance improvement. With this new design, the sequence of aerodynamic\u2013structural simulations and aerodynamic and structural adjoint calculations are repeated until a local minimum is determined. The results of the aerodynamic\u2013structural optimizations are summarized in Table 3. A similar behavior is observed for the different optimized designs. Figures 6\u20139 show the pressure distribution on a few of the wings with varying sweep used in this study. The aerodynamic performance is approximately the same and the structural weight is also roughly constant, suggesting that the aerodynamic\u2013 structural design space is also relatively flat. The planform variables of sweep and span, in particular, reveal interesting trends. The optimizer tended to slightly increase the sweep across the various configurations, whereas the span was slightly increased for the higher-sweep wings and slightly decreased for the lower-sweep cases. The aerodynamic\u2013structural optimizations permit changes in the wing thickness; in fact, the thickness of the outboard wing was slightly reduced at the lower sweep angles, yielding a small reduction in shock drag that contributes to the trend of an increase in the range factor M p L=D with reduced sweep. Similar to the pure aerodynamic optimizations, there is about a 4% improvement in range factor for the lowest-swept wing. The actual wing sections of the optimized 35 and 10 swept wings are displayed in Figs. 10 and 11. Figure 10 shows that the primary difference between the aerodynamic\u2013structural and pure aerodynamic optimizations for 35 is in the thickness of the outboard airfoil sections near the critical station. Figure 11 illustrates that this trend for the 10 case is more exaggerated. Just as interesting, varying the sweep for the pure aerodynamic optimizations has a dramatic effect on the airfoil camber distribution, as shown in Fig. 12. In the studies summarized by Tables 2 and 3 the lift coefficient was varied to maintain a roughly constant MCL, with the consequence that variations ofML=D are dictated by the value of CD. In fact, the optimum cruising lift coefficient may vary with Mach number in a different manner. To address this question, we selected the 10 and 35 wings for further analysis. As shown in Tables 4 and 5, the lift-to-drag L=D ratios of both wings show an increasing trend as the lift coefficient is further increased. However, it is doubtful whether a practical design could operate at such high lift coefficients, for a variety of reasons. First, these results are for single-point optimizations, andwhen the design point is too extreme, it typically leads to a rapid degradation away from the design point. Second, the wing must be able to support a 1:3 g turn without experiencing buffet, and this sets a limit on the usable design lift coefficient. Third, very high lift coefficients may require an excessively high cruising altitude beyond the capability of the engines or may require a decrease in wing area with a consequent decrease in fuel volume." + ] + }, + { + "image_filename": "designv6_24_0003207_apmc.2012.6421498-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003207_apmc.2012.6421498-Figure3-1.png", + "caption": "Fig. 3. Geometry of the orthogonal dual-polarized feeds.", + "texts": [ + " They are placed uniformly with separation of S=6.5mm. The dualpolarized antenna with two orthogonal -shaped probes is mounted on a box-shaped reflector. The -shaped probes are placed orthogonally in the separations between the metal posts. The gap distance between the probe edge and the metal post is Fgap= 1mm. As the complementary concept, one probe can excite a pair of electric dipole and a pair of magnetic dipole simultaneously, which can reinforce the broadside radiation and cancel out the back radiation. Fig. 3 depicts the details of the orthogonal -shaped probes. The -shaped probe consists of Portion 1, Portion 2 and Portion 3 as shown in the Figure. Each probe is connected to a coaxial launcher. For a high isolation, the probes are required to have different heights, with a difference of F = Fh2-Fh2-tc = 1.5mm. Port 1 of the shorter probe is for exciting the +45o polarization and Port 2 of the taller probe is for exciting the -45o polarization. Detailed dimensions of the antenna structure are summarized in Table I" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001190_bf00771891-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001190_bf00771891-Figure2-1.png", + "caption": "Fig. 2. Distr ibut ion of d i sp lacements w through the thickness of the root sect ion of a r ec tangu la r bar : 1) b / a = 1; 2) b / a = 2; 3 )b / a \" * ~", + "texts": [ + " -~-~ -t---g-l, ~/1 +Ta-~ ~ ; (3) u(A)-~k2~maxa - - - - ~ - l n l / l + 4 a 2 / b ~ + ~ - ~ - y - - - a r c t g +k~Ta T l n ~ l+b2/4a~-I + 2 ( 1 - - v ) ln Vf+4a~/y+__l ]. I /1 -t- 4a2/b e 1 ' (0) = O; (4) [ ~ ( )l [ 2b 1 l+bVa~+ 1] 2b 2 a a 2 (1 - -v ) In t:l+a~/b 2 + 1 +__d_ ln- u (0) ~ k2em,xa 2 arctg + ~ ~- .-arcig b- . \"+ k3*a i t, 1 + a2/b ~ -- 1 I/l + b~/aa - - i Similar equations may be obtained for other points of the c ross section. w l ( ~ ) a n d - - w ~ ( ~ ) through the bar thickness at The var iat ion of relat ive displacements Wrel= klOmax a [r I k xa y = 0 for different rat ios of bar width 2b to its thickness 2a is shown in Fig. 2. As can be seen, with a l inear rule for normal s t r e s s a var ia t ion the rule for displacement wl variat ion differs markedly f rom a l inear one, in view of which, it is only poss ib le to ta lk about a mean \"equivalent\" value of the root sect ion rotat ion angle, Accord ing to the data in [5] we de te rmine the equivalent rotat ion angle ~01equ f rom the equality condition of the opera t ion of no rma l s t r e s s e s a(x, y) fo r actual d i sp lacements wl(x, y) and for d i sp lacements o c c u r - r ing by a l inear ru le , ~01 equx: x d x " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002388_2015-01-0090-Figure11-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002388_2015-01-0090-Figure11-1.png", + "caption": "Figure 11. Strain probes on a cross member in lateral direction. An inset displays an actual corresponding cross member attached with strain gauge.", + "texts": [ + " The distances from wheel centre to level road at start position (yi) and stop position (yf) were measured. In the simulation, vertical displacement of yf\u2212 yi (Figure 9) was assigned to the left end of the front axle. Figure 10 is a contour plot of equivalent strains on the truck frame predicted by the simulation. For convenience, lateral and longitudinal strains were defined to comprehend our discussion later on. The lateral strain was normal strain measured in a direction along the length of cross members near their ends as showed in Figure 11. The longitudinal strain is normal strain located on the parallel flanges of side rails (U-shaped channel), Figure 12. With difficulty in mounting the strain gauges to a truck due to limit access, it was necessary to attach a strain gauge inside the channel for the top flange and outside the channel for the bottom flange. of ramp Two cases of a one-wheel ramp condition were investigated, i.e. ramping of a front left wheel and a rear left wheel. Practically one left wheel was on the top position of the ramp while another right spring experienced its maximum bump travel" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002563_physreve.100.013003-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002563_physreve.100.013003-Figure3-1.png", + "caption": "FIG. 3. A schematic physical model representation of the helical structure of the linkages. The white links represent the hinge sides of thin plates, whereas the yellow links represent the free sides. They define the helical skeletons with opposite handedness to each other.", + "texts": [ + " An individual plate has four sides: two are the crease lines (a valley and a mountain crease) and the other two are free. However, in the interfolded configuration, one of the two free sides is always occupied by the crease line of the pairing strip. Thus, in the assembled geometry, only one side of any plate is really force- and moment-free. We denote the other three as the hinge sides. These hinges are all one-dimensionally connected, and each constitutes a single helical frame with a period equal to that of the unit cell; see the white links in Fig. 3 that schematically show our structure. In contrast, the free sides also form a helical structure, but it makes one full helical turn every four unit cells with the handedness opposite to that of the hinge helix (see the yellow links in Fig. 3). Thus, by definition, there are three such identical helices of the free sides, each of which is separated by the distance of the unit cell along the long axis of the spring. (Note that only one of the three is displayed in Fig. 3.) Therefore, the skeleton of our paper spring may appear as the quadruple-stranded helix: one strand is a short pitch helix of the hinge lines with one handedness, and the other three are long-pitch helices with the opposite handedness. Note, however, that the exact fourfold rotational symmetry with a regular spiral progression along the long axis is geometrically impossible; the design satisfies the discrete fourfold rotational symmetry only approximately. Our spring actually cannot be represented as a mechanism of rigid straight lines of constant length linked by the hinges", + " 5, which compare well with those obtained from our numerical simulations shown in Fig. 6. The scaling plot in Fig. 5 shows that all the experimental and numerical data 013003-3 collapse onto a single curve, suggesting that the stretch-twist coupling is insensitive to the plate thickness t . Although an actual deformation involves complex bending and stretching of the plates, the observed twisting behavior of the spring may be predominantly geometric; it can be understood with a purely kinematic consideration. We focus on the helical linkage illustrated in Fig. 3 and assign a set of vectors of constant length a to those association lines. The unit cell, i.e., a full helical turn, is thus composed of a series of four vectors q1\u2013q4 (see Fig. 3). To proceed, we make the following simplifying assumptions. First, upon deformation, each association line remains a straight line of fixed length a, making a constant angle \u03c0/2 \u2212 \u03b8 with respect to the stretching axis (z axis): qi \u00b7 z\u0302 = sin \u03b8. (1) Second, any pair of consecutive vectors make a right angle, that is, the plates keep their square vertices upon deformations, which requires qi \u00b7 qi+1 = 0. (2) As discussed in the previous section, these two simplifying assumptions are actually not geometrically compatible but are still useful to derive an effective kinematical theory as long as the sheet deflections remain small enough" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001741_s00170-016-8553-0-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001741_s00170-016-8553-0-Figure3-1.png", + "caption": "Fig. 3 Three-dimensional schematic diagram illustrating preparation of the tensile test specimens", + "texts": [ + " The diamond-polished samples were subsequently polished with 50-nm colloidal silica slurry for 6 h using VibroMet 2 Vibratory polisher (Buehler). To obtain orientation maps, an accelerating voltage of 20 kV, working distance of 15 mm, and a step size of 50 nm were used. The HKL CHANNEL5 software was used to perform EBSD data visualization and post processing. In order to gain information about tensile properties of the welds, tensile specimens with the axis oriented perpendicular to the rolling direction were cut using WEDM process according to ASTM E8M-04 (see Fig. 3). Tensile test trials were carried out by keeping a cross head velocity of 0.5 mm/min. In order to ensure repeatability of fracture stress data, three samples for each condition were tested. The hardness of the weld joints was determined by microhardness measurements using a Vickers diamond indenter at a load of 100 g and dwell time of 10 s. For each zone, microhardness measurements were taken from three points at an interval of four times the indenter size to avoid the effects of localized strain hardening in the vicinity of the indentation" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003985_s10846-013-9874-y-Figure19-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003985_s10846-013-9874-y-Figure19-1.png", + "caption": "Fig. 19 Force analysis during peg-in-hole assembly in each hole", + "texts": [ + " The experimental data of force with voltage are collected as given in Table 2. The average force data are plotted in Fig. 18. It shows the almost linear relationship. The maximum value of force attains upto 12 mN. Using these force data during peg-in-hole assembly operation, the force of first IPMC finger (F1) and force of second IPMC finger (F2) are applied through voltages. During holding the peg, frictional force components of these IPMCs are dominated along with the weight of peg (W) as shown in Fig. 19. The peg touches the hole at the chamfer position. The sum of frictional force components along with peg weight must be equal or greater than vertical reaction force (N) component balances at chamfer angle (\u03b8) through force balance equation which is given below; \u03bcF1 + \u03bcF2 + W \u2265 N cos \u03b8 (9) where, \u03bc is frictional coefficient (0.3); \u03b8 is chamfer angle (45\u25e6); W is weight of steel peg (0.6 mN); F1 and F2 are calculated by relationship of bending moment with voltage at maximum condition (Fig. 10b). After satisfying this condition, the output force 10 mN from IPMC fingers is enough to hold the steel peg (diameter 1 mm and length 30 mm) when peg is inserted in the hole" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000567_tmag.2017.2703844-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000567_tmag.2017.2703844-Figure8-1.png", + "caption": "Fig. 8. Formation of magnetic gear. (a) Rectangle current sheet. (b) Structure of magnetic gear (p=1).", + "texts": [ + " Especially, the error of 2-D axis-symmetric FE model rises to 5.2% when reaches 15.2 deg. Thus, the applicability of 2-D axis-symmetric FE model is limited by radius of airgap. Table \u2161 Comparison of the Maximal Thrust Force by 3-D and 2-D Axis-symmetric FE Models Ra (mm) (deg) Max of Ft (kN) Error of 2-D FE models (%) 3-D 2-D 10.5 31.2 0.375 0.468 24.9 15.5 22.3 0.643 0.704 9.6 23.5 15.2 1.024 1.077 5.2 35.5 10.1 1.600 1.632 2.0 Similar to the derivation of 2-D axis-symmetric model, the rectangle current sheet shown in Fig. 8(a) is the projection of helical-shape current sheet on zr plane, and the structure shown in Fig. 8(b) is developed. The same current density is assigned in Fig. 8(a), so the 0018-9464 (c) 2016 IEEE. Personal use is permitted, but republication/redistribution requires IEEE permission. See http://www.ieee.org/publications_standards/publications/rights/index.html for more information. GT-10 4 torque Ta is given by '2 ( , ) R h a c rR T p J B r z rdr (8) Thus, the magnetic gear in Fig. 8(b) is expected to predict the torque of PMLS when similar magnetic fields distribution is obtained. And thrust force could be gained with the gear ratio. Fig. 9 shows the unfold-view of helical-shape PMs, similar to the basic system of PMLS in [7]. The line ac is the actual width of a pair of PMs, and the lines ab and ad are equal to twice the pole pitch measured along axial and circumferential directions, respectively. The degree of closeness to magnetic gear or ring is dependent on . Therefore, the angle determines the model error of magnetic ring and magnetic gear. According to Section \u2162, the error of 2-D axis-symmetric FE model is less than 5% when lead angle is lower than 15 deg. The question here is how much the range of lead angle should be so that the error of magnetic gear is tolerant. In order to confirm the appropriate lead angle for magnetic gear in Fig. 8(b), 2-D torque FE model is developed in Fig. 10. Due to the axial symmetry of magnetic gear, it\u2019s shown in Cartesian coordinate system. Four 6-pole-pair PMLSs are studied. The magnet axial widths of models are 46, 26, 21 and 18 mm in sequence. The screw rotates 30 mechanical degrees (half an electrical period) while the nut keeps stationary. Characteristics of torque are shown in Fig. 11. The curve of 3-D model approaches that of the equivalent 2-D model with the increasing width of PMs. The comparison of results by 2- D torque FE model and 2-D axis-symmetric FE model is shown in Table \u2162" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001744_lisat.2009.5031571-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001744_lisat.2009.5031571-Figure1-1.png", + "caption": "Figure 1. Complex electromagnetic environmental effects environments[1].", + "texts": [], + "surrounding_texts": [ + "978-1-4244-2348-4/09/$25.00 \u00a92009 IEEE\nK. D\u2019Ambrosio is a Intern with Northrop Grumman Aerospace Systems, Bethpage, NY USA. 516 704 3015; kristie.d\u2019ambrosio@ngc.com R. Pirich is with Northrop Grumman Aerospace Systems, Betghpage, NY USA. A. Kaufman is with the Computer Science Department at Stony Brook University, Stony Brook, NY USA. D. Mesecher is with Northrop Grumman Aerospace Systems, Bethpage, NY USA. P. Anumolu is with Northrop Grumman Aerospace Systems, Bethpage, NY USA.\nour systems to handle computations that are more difficult because the complexity of our world will only continue to increase.\nIndex Terms\u2014 Method of Moments, Multilevel Fast Multipole\nMethod, Graphics Processing Unit, ISR Systems\nI. INTRODUCTION\nhe study of electromagnetic interference is a familiar problem for most applications of electronic component integration which poses a significant challenge for efficient computation. Cosite interference is caused by radiated and conducted interactions of equipment and may include receiver desensitization, distortion of antenna patterns\nand extensive channel blockage. This problem has becoming increasingly more important to the defense industry because tactical command and control of military forces is increasingly reliant upon effective communications and efficient computation. Tactical communications are severely impacted when many apertures in close proximity experience such cosite interference. Effective mitigation of cosite interference requires a systematic treatment of multiple degradation mechanisms\nT", + "including basic mitigation techniques such as multicoupling, adaptive processing, filtering, and antenna modification, which can all be applied independently. Selecting synergistic techniques for a systematic, effective cosite interference treatment is a challenge that researchers have been working to overcome for years. With the help of new technology, such as Graphics Processing Units, our computation and communication systems are able to operate more efficiently and more effectively.\nThe efficiency of a technology is often heavily related to the tools it utilizes to complete its job. With respect to electromagnetic interference, these tools are first principle electromagnetic codes that compute the solution to various problems. One such technique is the Method of Moments or MoM calculation, which is a full wave solution of Maxwell's integral equations in the frequency domain. An advantage of the MoM is that it is a \"source method\" meaning that only the structure in question is discretised. Boundary conditions do not have to be set and memory requirements scale proportional to the size of the geometry in question and the required solution frequency. Another solving technique is the Multilevel Fast Multipole Method or MLFMM. This differs from the MoM in that it groups basis functions and computes the interaction between groups of basis functions, rather than between individual basis functions and hence provides a faster and higher frequency solution path. The software that we have been utilizing to execute these various solving techniques is translated as Field Computation for Objects of Arbitrary Shape or FEKO. The FEKO program is a full wave, method of moments (MoM) based, computer code for the analysis of electromagnetic problems such as: EMC, shielding, coupling, dielectric media and scattering analysis. FEKO includes the Multilevel Fast Multipole Method (MLFMM) which is an efficient method for solving electrically large problems. The MoM has been hybridized with asymptotic high frequency techniques, physical optics and the uniform theory of diffraction. The MoM/FEM (Finite Element Method) hybrid is a very efficient formulation for the analysis of inhomogeneous dielectric structures.\nNorthrop Grumman Corporation has spent a significant amount of research investigating optimizations of this FEKO program and its solution solving methods. One such optimization is a computation technique called Compressive Sensing or CS that has been developed by researchers at Duke University. This technique has the ability to effectively use randomization to minimize the redundancy inherently found in electromagnetic computation. With the application of some smart extrapolation, an incredibly accurate solution can be formulated using a fraction of the information originally necessary for a MoM or MLFMM computation.\nNorthrop Grumman Corporation has put a great deal of effort into the minimization of the computational peak memory & calculation time. The FEKO software has two built-in default settings that ensure the accuracy of the system as the company has specified for the user. We have predicted that these settings are overestimations and can be optimizated to greatly reduce computation runtime. These settings include the iteration number, which is directly related to the amount of time a simulation takes, and the convergence level which represents how accurate a given solution is. Through experimentation, we have been able to find the optimal numbers for both these settings that allow the experimental solution to not deviate too drastically from the actual solution and therefore preserve the accuracy we required of the simulation. By finding these numbers we were able to get an acceptably accurate solution in the shortest amount of time." + ] + }, + { + "image_filename": "designv6_24_0000022_robot.1999.770392-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000022_robot.1999.770392-Figure2-1.png", + "caption": "Fig. 2: Searching motion of Multiple Active Antenna", + "texts": [ + " For simplifying our discussions, we assume that the compliance of the object is sufficiently small compared with that of the beams. It is also assumed that the curvature of the object\u2019s surface is negligible. Now, suppose that the motor is rotating to search an object. Two points in the same distance from the center of rotation passes on the same trajectory, since two beams are implemented so that they may rotate in the same searching plane. Now, assume that one beam makes contact with an object, as shown in Fig.2(a). For a further rotating motion, the other beam makes contact with the same point of the object, as shown in Fig.2(b). When the contact distance S > d, rotation angle 6 < 7~12. In such case, by utilizing the geometrical relationship between the beams in Fig.S(a), we can obtain the following equation. S = dsin8+ Scos8, (1) where d is a half of the gap between the parallel beams and 8 is the rotation angle as shown in Fig.3. When the contact length 0 < S 5 d, the equation obtained from Fig.3(b) is S = dsin(7r - 8) - S c o s ( ~ - 8) (2) where 6 2 ~ / 2 . From eq.(l) and (2), we can easily compute the contact distance S as follows" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001297_s10766-009-0125-6-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001297_s10766-009-0125-6-Figure5-1.png", + "caption": "Fig. 5 The task graph T \u2032", + "texts": [ + " We consider a small example to explain the algorithm. Let T be a task graph with 3 vertices, v1, v2, v3, such that an edge from vi to v j exists if and only if j = i + 1. Let e(vi ) = 1 for all 1 \u2264 i \u2264 3 in T . The deadlines of the vertices are d(v1) = 2, d(v2) = 3, and d(v3) = 2. The minimum intertriggering separation times associated with the edges are p(v1, v2) = 3, and p(v2, v3) = 3 (see Fig. 4). Let T \u2032 be the graph that is formed by joining two copies of this task graph T in the fashion described in Sect. 4.1.2. T \u2032 is shown explicitly in Fig. 5. The table built by Algorithm 1 for T \u2032 is shown Table 1. For any 1 \u2264 i \u2264 6 and 1 \u2264 e \u2264 6, the (i, e)th cell of this table contains the values of ti,e, and t i i,e respectively. Row 1 in the table is filled according to the initialization statements in line 2 of the algorithm. The condition (e(v1) = e) is never satisfied in row 1, because the e(v1) = 0. Hence, As specified in line 2 all the values t1,e in this row Fig. 4 The task graph T are set to\u221e. The rest of the rows are constructed according to the recursive procedure (lines 5 to 11 in Algorithm 1). Consider the computation of t2 2,1 at row 2 and e = 1. This values corresponds to the vertex v2 of the task graph T \u2032. Note from Fig. 5 that for this vertex e(v2) = 1 = e. Thus, according to the conditions of line 8 t2 2,1 = d(v2) = 3. It maybe also be worked out from line 9 that t2,1 = t2 2,1 = 3 because t2 2,1 < t1,1. The values of other cells in the table maybe worked out similarly. In the following section, we reformulate this computationally expensive algorithm as a streaming algorithm for implementing it on a GPU. 4.2 Schedulability Analysis on GPUs Before discussing the details of our GPU-based engine\u2014GPUSched, we outline the overall scheme to reformulate any algorithm as a stream processing application to run it on a GPU (Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002426_tvcg.2007.1033-Figure28-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002426_tvcg.2007.1033-Figure28-1.png", + "caption": "Fig. 28. (a) Finger model. (b) Thumb model.", + "texts": [ + " [60]) and, for a fully automatic system, aspects of natural grasps other than the ability to apply forces must be taken into account. Much research remains to be done to better understand human grasping (for example, see [61]) and to develop algorithms that quantify the \u201cnaturalness\u201d of grasps. For example, comfort and reachability of the hand and arm pose must certainly be considered. APPENDIX Based on the biomechanics work by Valero-Cuevas et al. [62], [63], we use a hand model that has the following configuration (see Fig. 28): The fingers have four DOF. The metacarpophalangeal (MCP) joint has two DOF of flexion/ extension and adduction/abduction. The proximal inter- phalangeal (PIP) and the distal interphalangeal (DIP) joints both have one DOF of flexion/extension. The thumb has five DOF in total. The carpometacarpal (CMC), MCP, and thumb interphalangeal (IP) joints each have one DOF of flexion/extension. The CMC and MCP joints both have another DOF of adduction/abduction. In the actual human hand, the ring and little fingers have two additional DOF of flexion/extension and adduction/abduction at the CMC joint" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001764_tia.2010.2057398-Figure17-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001764_tia.2010.2057398-Figure17-1.png", + "caption": "Fig. 17. Stall test setup.", + "texts": [ + " In the same way, thermocouples 2 and 4 were positioned on a tooth at the top and the bottom, respectively. As the model considers the heat transferred from the rotor bars to the rotor lamination, the temperature rise of the tooth is critical for the model validation process. Thermocouples 5 and 6 were placed on one of the end rings. The radial end ring expansion was measured to check the calculated value. A picture of the test setup is presented in Fig. 16. With the shaft mechanically locked, as presented in Fig. 17, the motor was energized for several seconds. Two motors were tested under this condition: 2000-hp twopole copper squirrel cage and 1500-hp four-pole aluminum squirrel cage, both rated at 4.16 kV and 60 Hz. A summary of the test results is presented in Table I. The plots of the two-pole machine, tested under 2000 V, are presented in Figs. 18 and 19. All the results for the temperature rise were consistent with the calculations, indicating that the skin effect was properly considered. The small deviations observed in the peak temperature are probably related with the inaccuracy of the test measurements" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001994_j.rcim.2007.04.003-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001994_j.rcim.2007.04.003-Figure5-1.png", + "caption": "Fig. 5. Block diagram for robot with deburring tool (hybrid control).", + "texts": [ + " The controller is designed by considering these coordinates, whose control input t can be transformed into torque in the joint coordinates. The transformation is t \u00bc JT\u00f0x\u00deuh. (38) In the constraint coordinates, the force and position errors are defined. f di 2 R1 is the desired force in the xi direction (i \u00bc 1,2) and xid 2 Rn 1 is the desired position in other directions. From the assumption that the stiffness ke is known, ep is defined as ep \u00bc K 1 ef ep2 ! \u00bc K 1 f d1 f 1 x2d x2 ! , (39) where k \u00bc ke 0 0 I . xi is the displacement of the robot, xid the desired position , and f i 2 R1 the force component of the xi direction. Fig. 5 depicts hybrid control of a robot with a rigid tool without a pneumatic cylinder. Simulation study was performed to investigate the performance of the controllers developed for the robotic deburring systems with different tools: (1) the hybrid controller for the rigid tool based system; (2) the coordination controller for the single active pneumatic tool. ARTICLE IN PRESS C. Kim et al. / Robotics and Computer-Integrated Manufacturing 24 (2008) 462\u2013471468 Fig. 6 shows the simulation results for the hybrid control system" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002661_iros.2011.6094731-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002661_iros.2011.6094731-Figure4-1.png", + "caption": "Fig. 4. Detailed CAD model of the vehicle\u2019s central part. The slots in the vertical carriers are designed as flexible mounts for different payloads. In the picture an additional 1.6 GHz lightweight Atom computer board is mounted below the FCU.", + "texts": [ + "2 kg including the solar cell array and a backup battery due to a total average power limit of 180 W. The weight of the solar cell arrangement, mechanical mounts, charging and monitoring electronics accumulated to 350 g. A 1350 mAh 3 s LiPo battery with 100 g weight was chosen. Thus the final configuration had a weight of 1050 g, leaving enough margin to charge the battery. The mechanical design is motivated by fast changing requirements for the payload units. Fig. 1 shows a CAD model of the complete system and Fig. 4 the design of the central part, which is inspired by computer racks. It consists of vertical carriers with slots for horizontal payload inserts. The complete frame is built with 2-D milled carbon fiber and carbon fiber-balsa-wood sandwich material. It is assembled only by screws without the need of gluing components, enabling very fast component changes or repairs. The FCU features a combination of a high speed IMU and two 32-bit 60 MHz microcontrollers for the flight control algorithms. It is specially designed for the needs of small UAVs" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001066_detc2006-99207-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001066_detc2006-99207-Figure2-1.png", + "caption": "Figure 2. Symmetric position and principal axes of the coupler curve.", + "texts": [ + " A rotation of 90\u00ba about the symmetric coupler point gives a candidate line for locating the \u201chip\u201d H or origination point of the applied force. This line corresponds to the locus of hip locations giving a transmission ratio of 1 when the coupler point (pedal) is on the axis of symmetry of the coupler curve. While this is a good initial guess for relative location of the hip, it may not result in optimal force transmission over the whole range of pedal motion, so a parallel shift or offset of this line can be introduced as an additional free variable (shown as d in Figure 2). For any instantaneous position of the mechanism, the Kennedy-Aronhold theorem [13] dictates that the instant center I13 which determines the center of curvature of the coupler curve can be located by finding the intersection of the (extension) line of link 2 and the (extension) line of link 4, as shown in Figure 3. When the crank and rocker links are parallel, the instant center lies at infinity, and the coupler curve has zero curvature at this point. 2 Copyright \u00a9 2006 by ASME ms of Use: http://www" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000284_s10846-005-0932-y-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000284_s10846-005-0932-y-Figure6-1.png", + "caption": "Figure 6. In this figure the established ortho-normal coordinate system is illustrated.", + "texts": [ + " The length of the baseline b is calculated in cm. To calculate the modified equations, giving the coordinates of a random point in space, we solve the stereo problem using as cameras the two virtual cameras created by PSVS. Each of the above virtual cameras can be separately calibrated and the matrix A with the intrinsic parameters can be found. A right-hand ortho-normal coordinate system with origin the optical center O of the real camera and Z-axis to coincide with the optical axis of the real camera is established (Figure 6). The coordinates of a point P in space, with respect to this coordinate system, are expressed by the vector X = [x, y, z, 1]T while the coordinates of optical centers O1 and O2 are provided from vectors XoR = [xoR, yoR, zoR]T and XoL = [xoL, yoL, zoL]T, respectively. Then, using matrix relations, the calculation of vectors mL = [uL, vL, 1]T, mR = [uR, vR, 1]T in the image plane for each virtual camera of a point P is possible. Here, the assumption made, is that the two virtual cameras are parallel to the real camera as a result of the above alignment and checking procedure" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002164_taes.1986.310776-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002164_taes.1986.310776-Figure4-1.png", + "caption": "Fig. 4. Disc displacements.", + "texts": [ + " ASSUMPTIONS AND CONSTITUTIVE RELATIONSHIPS In order to simplify the field equations which describe the mechanical and electrical behavior of the gyroscope, it is assumed that the disc is thin. This permits the mechanical problem to be treated as one of plane stress [7]. Here the stresses T, Tzr, and T.o are zero on the upper and undersurfaces of the disc, and the thinness assumption implies that they can be neglected in the body of the disc. The remaining stresses are taken to be constant in the thickness direction OZ. The displacement u of a typical point P in the disc is shown in Fig. 4 and is given by the radial and tangential components ur and uo. On the electroded surfaces of the disc there can only be an electric field component normal to the surface Ez. This condition is true for the whole disc undersurface and most of the adjacent upper sulface. If fringing is neglected in the region of the electrode edges, it can be assumed that Ez is the only electric field component within the disc. The constitutive relationships for axially polarized PZT and lithium niobate for this planar situation can be derived from the general expressions given in the IRE Standards [8] and are given by IEEE TRANSACTIONS ON AEROSPACE AND ELECTRONIC SYSTEMS VOL" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001692_kem.622-623.819-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001692_kem.622-623.819-Figure2-1.png", + "caption": "Figure 2. (a)Schematic view of diffusion bonding apparatus and (b) cross sectional view of the formed article with inner copper channel", + "texts": [ + " The maximum operating temperature is 1,200\u2103. The heating chamber is insulated with refractory blocks and water cooling is forced in the upper bolster plate to protect press frame structure from heating. The personal computer based automatic control system can accurately monitor gas pressure by a flow control valve with diaphragm type integral positioner. Figure 1(b) shows the duplex stainless steel outer jacket, blow formed with gas pressure of 7MPa at 980\u2103[3]. The diffusion bonding was conducted in the special tool[4] as shown in Figure 2(a). Fixture made with a corrosion resistant high alloy steel was designed to allow hydrostatic gas pressurization from inside so that solid state bonding of copper and steel is possible. The fixture assembly is placed in the furnace with ceramic heaters and bonding is conducted in inert environment after uniform temperature has been achieved throughout. Since this tool needs a gas pressure for manufacturing a diffusion bonding product, it is required to sustain the sealing condition at high temperature and high pressure. Figure 2(b) shows the successful bonding of formed article with inner copper channel bonded to outer steel jacket [4]. IN718 is one of the most popular metals for high temperature application in aerospace and propulsion industry. Since it is difficult to machine due to its high strength, abrasiveness, low thermal conductivity, and high cost of machining, the superplastic forming of IN718 has been of a major interest to the aerospace industries, and a considerable amount of research is carried out in this field" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003865_tia.2019.2923717-Figure11-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003865_tia.2019.2923717-Figure11-1.png", + "caption": "Fig. 11. Flux line and flux density without load currents", + "texts": [ + " Finally, the shape of the determined outer rotor PM vernier generator is shown in Fig. 10. Fig. 10. Designed structure of outer rotor PM vernier generator V. SIMULATION AND EXPERIMENTAL RESULTS To prove the validity of the proposed design procedure and the derived equations, the performance characteristics of the designed DD vernier generator of Fig. 10 are analyzed by FEM. First, the flux distribution and the phase back EMF characteristics without armature currents are calculated at the base speed 214rpm and illustrated in Fig. 11. It shows that the 0 0.2 0.4 0.6 0.8 1 1.2 1.4 1.6 1.8 2 -0.1 0 0.1 0.2 0.3 0.4 0.5 0.6 p w g m /D g 0 0.2 0.4 0.6 0.8 1 1.2 1.4 1.6 1.8 2 -0.1 0 0.1 0.2 0.3 0.4 0.5 0.6 g m /D g 0 0.2 0.4 0.6 0.8 1 1.2 1.4 1.6 1.8 2 0 0.1 0 0.2 0.4 0.6 0.8 1 1.2 1.4 1.6 1.8 2 3 4 5 6 7 x 10 -3 D g2 l s tk [ m 3 ] 0 0.2 0.4 0.6 0.8 1 1.2 1.4 1.6 1.8 2 0 0.1 0.2 0.3 0.4 g m /D g pw=3 pw=2 pw=1 0.5 0.52 0.54 0.56 0.58 0.6 0.62 0.64 0.66 0.68 0.7 0.01 0.02 0.03 0.04 0.05 0.06 0.07 g m /D g pw=2 pw=3 pw=1 260mm rotor stator NdFeB magnets shaft 2 auxiliary teeth 0093-9994 (c) 2019 IEEE" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001801_j.triboint.2016.11.027-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001801_j.triboint.2016.11.027-Figure2-1.png", + "caption": "Fig. 2. Schematic view of structure of nested compression spring foil bearing.", + "texts": [ + " The loaddisplacement curves for the bearing with different spring numbers and operating speeds were measured. The theoretical and experimental results presented in this paper provide a guideline for designing the NSFB with high load capacity and stability. Building a theoretical model that can accurately predict NSFB performance is necessary to promote its application and improve its design and manufacturing efficiency. A prediction model of NSFB is built in this section by coupling the Reynolds equation, the finite element top foil model, and the elastic foundation model. Fig. 2 details the structure of the proposed NSFB. A smooth top foil is supported by a series of compression springs. Each spring is screwed into a piece of sheet metal and mounted to the grooves on the inner side of the bearing sleeve. Two contact points are placed between adjacent spring coils because of the small distance between adjacent springs. The dry frictions between two adjacent springs are generated considering the axial preload of the springs when the gas film pressure is applied on the support structure" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000882_tmtt.2004.823575-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000882_tmtt.2004.823575-Figure1-1.png", + "caption": "Fig. 1. Three-dimensional view of the designed triangular type CMRC. h = 1:524 mm, \" = 2:94, W = 3:92 mm, W = 0:5 mm, W = 5:22 mm, L = 0:5 mm, L = 9:8 mm, L = 10:7 mm, and L = 14:3 mm.", + "texts": [ + " More importantly, with proper control of the slowly varying phase angle in the stopband of our CMRC, this method shows a better delay mismatch error for the reflected signals. The proposed method not only shows a significant improvement in the IMD products without affecting the fundamental signal levels, but also utilizes extremely simple, low-cost, compact, and easy-to-construct circuitry, resulting in a new approach in designing power amplifiers with high linearity and efficiency. 0018-9480/04$20.00 \u00a9 2004 IEEE Fig. 1 depicts the CMRC structure with a triangular pattern proposed in [19] and [20]. This structure features a very simple yet compact one-dimensional (1-D) design that offers broad stopband ( for less than 10.0 dB) and moderate slow-wave factor (1\u20132 times more than the conventional microstrip). The width of the CMRC has been made wider than that of the 50- transmission lines in order to reduce the passband insertion loss [21]. In this study, as the CMRC is used as a reactive termination for both the second and third harmonics, it is important to discuss some physical parameters that can affect the covering bandwidth as well as the multiresonant points inside the stopband" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001388_s1644-9665(12)60163-0-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001388_s1644-9665(12)60163-0-Figure3-1.png", + "caption": "Fig. 3. Part of the subframe (a), photographs of forming of the extruded profile: bending (b), pressing (c), hydroforming (d) [21]", + "texts": [ + " In the last few years, the demand for weight reduction in modern vehicle construction has led to an increase in the application of hydroforming processes for the manufacture of automotive lightweight components made from steel or aluminium. It results from improvements in stiffness and crash behaviour due to the reduction of welding seams, and with reduced assembly costs. Currently, use of aluminium alloy which is light weight material is increasing. Extruded aluminium profiles have been used for automobile frame parts [10] to get higher stiffness and light weight, Figure 3. Many progresses have been made in forming of square sectioned profile by stretch bending [12, 15] and hydroforming [16\u201317]. The use of lightweight materials such as aluminium and magnesium can reduce the weight of passenger vehicles up to 40\u201375% by replacing ferrous auto body structures and body panels [18]. It was reported that a 10% weight reduction in an average automotive body could improve the fuel efficiency by 6\u20138% [19]. However, the cost for the lightweight materials is estimated to be higher than that of the mild steel structures because of the raw material price and the production costs with the existing manufacturing technologies" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001971_apcap.2014.6992489-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001971_apcap.2014.6992489-Figure1-1.png", + "caption": "Fig. 1. Geometry of the proposed wideband CP patch antenna excited by 3D M-probe.", + "texts": [ + " The 3D M-probe is implemented in four pieces of printed circuit blocks. The proposed 3D M-probe feeding mechanism is equivalent to a dual-feed sources, which consists of one coupling portion and one directed probe-fed portion. The geometry of the antenna is simple and easy to fabricate. It finds an advantage of ease of is implementation and integration with RF circuitries by PCB fabrication technology. II. ANTENNA GEOMETRY The geometry of the proposed 3D meandering probe fed CP patch antenna is illustrated in Fig. 1. It comprises a circular patch, 3D M-probe portion and a square ground. The radiating patch is in the shape of circle with radius R of 15 mm. The meandering probe is formed by printed strips and the metalized via holes on the substrates as illustrated in Fig.2. The substrate is from Rogers 4003 with a dielectric constant of 3.38, and a layered thickness of 1.6 mm. The 3D M-probe is shown in detail in Fig. 2. The circular patch is excited at the two points, A and B. A is excited by the coupling feeding method, while B is excited directly by the probe" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003310_978-981-10-5544-7_23-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003310_978-981-10-5544-7_23-Figure1-1.png", + "caption": "Fig. 1 Geometry of proposed antenna a front view, b back view, c inner dumbbell dimensions", + "texts": [ + " The microstrip patch antenna with Defected ground will provide little bit higher bandwidth and nominally less return loss. Whereas without DGS, it will provide narrow bandwidth and return loss will be high. For improving the radiation DGS is integrated on the ground plane. Previously the circular ring and y-shape-strip with defected ground plane for WiMAX and WLAN application. The paper achieved three resonant frequencies at 2.61, 3.5, 5.4 GHz [11, 12]. The proposed antenna is having the UWB nature. The simulated results are presented with key structure parameters are analyzed (Fig. 1). Table 1 explains the antenna parameters of UWB proposed antenna. It consists of the radiating patch at the upper part of the substrate, which consists of a circular shaped ring and semicircle strip. The substrate has a length and width of Ls = 38 mm, Ws = 25 mm. The circular ring patch consists of outer radius R1 and inner radius R2, respectively. The semicircle consists of the outer radius of R3 and inner radius of R4. The circular ring patch antenna is fed by a 50 X microstrip feed line which is having a width of Wf = 2" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002590_0022-2569(71)90034-6-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002590_0022-2569(71)90034-6-Figure3-1.png", + "caption": "Figure 3. Intersection of cylinder with Torus, having four points of third order tangency.", + "texts": [ + " If Aa and At are made to coincide at the centre A of a spherical joint 34, a closed-loop linkage is formed, and A is then constrained to follow one branch of this curve of intersection. By suitably choosing the proport ions of the torus and the size and position of the cylinder, one branch of the curve of intersection can be made to have third order At the points E and G the spatial curve of intersection has stationary zero curvature; at points F and H it has stationary curvature of radius r. By considering the section y = \u00b1 (r/p) q through the torus, it can readily be calculated that the condition that must be satisfied in order to obtain a curve such as that shown in Fig. 3 is r X h,_q r = 4 - ( p - - r ) ~ p., ( p - - r ) \" (I) If the revolute pair 12 is the driving axis, with link 2 the input crank, this can have full rotation, the output link 4 combining both rotatory and translatory motion. Its rotatory motion will be from + 0 t o - ~ where Also sin 0 = -q (2) r p = q cot tk+ b. (3) Equations (2) and (3) can be used to replace r and p by ~b in equation ( I ) to obtain sin 3 0 -- sin\" 0 ( 1 - 2 cos 0) +s 'n0 l--cos0\u00f7co, 0, 0 which may be solved for b/q when any required value of ~ (half of the output angular range) is given" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001666_b978-0-12-809880-6.00026-6-Figure26.4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001666_b978-0-12-809880-6.00026-6-Figure26.4-1.png", + "caption": "FIGURE 26.4 (A) Hydraulic bone chamber implant used to apply cyclic compressive loading to tissue engineered constructs in vivo. Implanted constructs (B) receiving the mechanical stimulus (right) had 9-fold more new bone formation than no load controls (left). (C) Stiff (top) and axially compliant (below) fixation plates. (D) Compliant fixation (right) enhanced bone formation and altered local strain disributions over stiff fixation (left). (E) Delayed loading increased vascularization compared to stiff fixation.", + "texts": [ + "5 Hz were found to have ninefold more new bone formation compared with contralateral control constructs that did not receive loading. Using preclinical models of large bone defect regeneration, several groups have begun to evaluate the influence of in vivo mechanical loading on tissue-engineered bone repair [105e110]. For example, Boerckel et al. developed a rat segmental bone defect model that enables the control of ambulatory load transfer through modulation of axial fixation plate stiffness [106] (Fig. 26.4A). These studies demonstrated that bone regeneration depends on both the magnitude and timing of in vivo loading, with early loading inhibiting vascular ingrowth leading to pseudarthrosis but delayed loading enhancing bone regeneration and remodeling leading to restored functional bone properties [105e107] (Fig. 26.4B,C). Notably, this delayed loading protocol failed to induce regeneration in the presence of structural scaffolds that mimicked the properties of native bone [110], which suggests that stiff scaffolds can impede the beneficial effect of mechanical loading by stress shielding or inhibition of tissue ingrowth. Collectively, these observations suggest that the mechanical environment should be considered together with the physical and biochemical properties of engineered extracellular matrices for tissue regeneration" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000548_6.2003-1868-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000548_6.2003-1868-Figure2-1.png", + "caption": "Figure 2: Gantry-crane model.", + "texts": [], + "surrounding_texts": [ + "We use the Lagrangian approach to derive the equations of motion of the gantry cran shown in Fig- ure (m + M) x\u0308 + mL\u03c6\u0308 cos (\u03c6) + mL\u0308 sin (\u03c6) + (1) 2mL\u0307\u03c6\u0307 cos (\u03c6) \u2212 mL\u03c6\u03072 sin (\u03c6) = Fx L\u03c6\u0308 + g sin (\u03c6) + 2L\u0307\u03c6\u0307 + x\u0308 cos (\u03c6) = 0 (2) For safe operation, the swing angle should be kept small. In this study, we assume that changing the cable length is needed only to avoid obstacles in the path of the load. This change can be considered small also. Using these two assumptions and dividing equation (1) by M , we reduce the equations of motion to x\u0308 \u2212 mtg\u03c6 = F\u0304x (3) L\u03c6\u0308 + g\u03c6 + x\u0308 = 0 (4) where mt = m M , F\u0304x = Fx M (5) Because the motor has a small time constant relative to the mechanical system, the force exerted by it can be considered as a constant gain and expressed as F\u0304x = KmxVx (6) where Vx is the input voltage to the motor." + ] + }, + { + "image_filename": "designv6_24_0003937_978-1-4471-5104-3_5-Figure5.8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003937_978-1-4471-5104-3_5-Figure5.8-1.png", + "caption": "Fig. 5.8 Structure of a typical multi-pole permanent magnet generator", + "texts": [ + " High power electronic switches change up from one electromagnet to the next as the shaft turns. The PMSG model with a single coil shown in Fig. 5.7 is not usual because its torque is very unsteady and a higher number of coils would be advisable. Once the rotor began its movement by one of the poles, the next one will be just near enough to receive the next magnetic push and so on. With such configuration, the rotor movement will be smoothed out as the number of poles increases. More typically, the number of poles could be three or a multiple of three as shown in Fig. 5.8. This generator can be modeled much like the ordinary salient pole synchronous generator with the stator connected in the star and the two rotor poles seen as an electromagnet. The stator windings are sinusoidally distributed and shifted from each other by 120 as the phases represented in Fig. 5.9 by resistors and inductances. In the rotor magnetically coupled to the stator there is an equivalent winding to the magnet that is represented also by a resistance and an inductance. The current going through the stator coils is assumed as the positive sense to describe the generator excitation" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001452_j.isprsjprs.2008.10.001-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001452_j.isprsjprs.2008.10.001-Figure1-1.png", + "caption": "Fig. 1. Circular image block imaging constellation. Between the first and second block, the camera will be turned into the opposite direction.", + "texts": [ + " Therefore, there should be two such image sequences taken from the same spot and, in these image blocks, the alignment of the camera with respect to the tangent of the trajectory should be opposite. In the first block, the camera is set in the direction of +90\u25e6 with respect to the supporting bar, and in the second block\u221290\u25e6. This resolves the problem of parallel image rays. Now there are images in both image sequences where the same object can be seen on images which are taken, atmost, two times the length of the bar apart from each other, see Fig. 1. In principle, orientation information of a camera in each epoch of exposure could be acquired mechanically. The alignment of the camera at the end of the bar could be measured beforehand and the angle of rotation during imaging could be observed with an appropriate sensor device. Nonetheless, angular values of image rays should be resolved with high precision, and measurements based on mechanical devices hardly fulfill that requirement. Therefore, the exterior orientation of images will be based only on image observations. The need for auxiliary devices cannot be entirely neglected though. These instruments can provide good initial values for the exterior orientation values at the final computing stage. Although, in Fig. 1 the imaging is depicted to be acquired by a single camera, there is no reason not to use two cameras simultaneously. In that case, cameras would be attached to both ends of the bar and the navel point would be in the middle. However, the imaging should not be confusedwith stereo imaging. In stereo imaging a stereo pair is treated as a computing unit but, in this computing model, image observations are treated as image ray bundles in common adjustment. There have been investigations using stereo pairs in block adjustment, where the bundle adjustment has been computed with constrained stereo pairs (King, 1994)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002584_1999-01-0782-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002584_1999-01-0782-Figure3-1.png", + "caption": "Figure 3. Simplified Wheel used for Preliminary Modeling Calculations", + "texts": [ + " These two factors made the Kuhl Wheel seem like an excellent starting point to reduce the weight of the conventional stamped steel wheel. To evaluate the feasibly of the Kuhl Wheel design, preliminary calculations were made to understand the peak stress levels on the Kuhl Wheel during cornering and braking. A spoke geometry was selected for the wheel that used ten rectangular shaped spokes, each spoke being offset and parallel from a radial line drawn outward from the hub towards the rim. This arrangement is shown in Fig. 3. We believed that the arrangement of the spokes in offset pairs would provide minimum deflection of the hub with respect to the rim during cornering while ensuring hub torsional loads resulted in primarily spoke tension. CORNERING FORCE ANALYSIS \u2013 Initial analysis was concerned with cornering loads and the first estimates were done by considering the spoke to be a simple beam, fixed at the hub end and guided at the rim end as shown in Fig. 4. The cantilever treatment at one end was believed to be a reasonable representation of the clamp retention of the spoke at the wheel hub" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003181_s10916-016-0462-0-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003181_s10916-016-0462-0-Figure4-1.png", + "caption": "Fig. 4 Active surgical tool. a Structural diagram of surgical tool; b Simplified model of surgical tool; c Schematic diagram of rotatingQ around the tip", + "texts": [ + " This method requires iterative operations similar to other methods, and the frequency of iteration is determined by initial conditions. This paper adopts an easy method for the correct and rapid calculation of the coordinates of the tips in the registration process. Instead of using the iterative operation, the method adopted in the present paper directly uses mathematic analysis to solve equations, thus making the calculation simple and fast. In this method, three near-infrared light-emitting diode chips (a, b, and c) are installed on the top of the surgical tool (Fig. 4a). In an energized state, points a, b, and c emit near-infrared light, which are easily captured by the tracking system when laid in the effective working scope of the system. To quickly find the relationship between the three light-emitting points and the tips in the registration process, the distances between these three points of the surgical tool are unequal. Figure 4b shows the establishment of the coordinate system after calibrating the internal and external parameters of the system acquired by near-infrared cameras. In the registration, the specific location from the tip of the surgical tool to the working area of the system is fixed, and the surgical tool rotates around the tip. This process must ensure that both cameras can shoot three light-emitting points of the surgical tool (Fig. 4c). The coordinates of the tip Q should meet the following relation: Pq \u00bc Rq\u22c5Pq \u00fe Tq; \u00f01\u00de where Rq and Tq are the rotation matrix and translational matrix of the transformation from position U to position V in the surgical tool, respectively. The coordinates of the tip Q can be obtained by transforming Eq. (1): Pq \u00bc I\u2212Rq \u22121\u22c5Tq; \u00f02\u00de where I is the 3\u00d73 unit matrix. Suppose that Rq=(I\u2212Rq), the angle of revolving to a specific coordinate axis may be zero in the coordinate transformation. The coordinates of Pq cannot be acquired by Eq" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002597_j.ijthermalsci.2007.06.020-Figure12-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002597_j.ijthermalsci.2007.06.020-Figure12-1.png", + "caption": "Fig. 12. Electric submersible pump well system basic model (dimensions are in m).", + "texts": [ + " Several 2D finite-element models of the electric submersible pump system are needed for the analysis of the electric submersible pump system design and operation under different possible conditions. The VectorField\u2122 Opera-2D software package [22] is used for modeling and simulating the 2D electric submersible pump system. Figs. 4 and 8 show that the profiles of the motor surface temperature and the total magnetic-flux density decrease near to the motor ends. Hence, it is not worth to calculate the eddy currents in axial direction using 3D FEM. The basic model of the electric submersible pump well system, including the motor, oil and well casing, is shown in Fig. 12. From the central part and moving out radially, there are the: rotor shaft, rotor core, rotor cage, airgap, stator windings (coils), stator core, motor casing, oil (fluid) mixture, well casing, and cement. For the present 2D finite-element modeling, the numbers of elements, nodes and regions were set as 31 291, 16 244 and 22, respectively. Firstly, the meshes were automatically refined. Then, in order to increase the computation accuracy, the meshes were manually refined for the motor and the well casings" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001388_s1644-9665(12)60163-0-Figure11-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001388_s1644-9665(12)60163-0-Figure11-1.png", + "caption": "Fig. 11. Audi TT Coupe with hydroformed rear axle components [42]", + "texts": [ + " Today, steel suppliers offer new steel grades, e.g. high strength steels [39\u201340], competing with aluminium as new materials for lightweight constructions combined with hydroforming. One of the leading automotive companies using hydroforming is BMW [41]. In the newest models of BMW (e.g. BMW M3) hydroformed exhaust components are used (Figure 10). New technology of tube hydroforming has become so profitable from many points of view, that other car producers have introduced new components into cars, as for example Audi (Figure 11). There exists a considerable interest to reduce vehicle weight through the adoption of lightweight materials, such as aluminium alloys, while maintaining energy absorption and component integrity under crash conditions. The interaction between tube hydroforming and behaviour during crash events was studied using lightweight automotive structural members [43]. There was used a high-pressure hydroforming process in which tubes with various corner radii in the tube cross-section were produced, Figure 12" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000427_0921-5107(91)90125-f-Figure13-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000427_0921-5107(91)90125-f-Figure13-1.png", + "caption": "Fig. 13. Schematic diagram of multipolar ECR plasma-etching system.", + "texts": [ + " Since the motion of the electrons is constrained by the external magnetic field, fewer are lost by collisions with the reactor walls than in a conventional r.f. plasma and therefore the plasma potential relative to ground is much lower. Th e resultant energies of ions reaching the sample to be etched are typically 15 eV or lower. Since this is less than the displacement threshold for damage in most semiconductors, ECR etching should lead to much lower levels of damage than conventional RIE processes. A schematic of the ECR reactor that we use is shown in Fig. 13 (Plasma-Therm SL720). The sample is manually loaded into the load lock and then transferred into the etch chamber on a robotic arm. The system utilizes a 195 TABLE 3 Depth of various components of the near-surface residue found after reactive ion etching AlxGa j _.~As with a CC12F2 discharge for 4 min at 4 mTorr Component Depth (A) Control A B C D Fluorocarbons 0 < 10 < 10 < 10 < 10 As203 < 10 < 10 < 10 < 10 < 10 Ga203 10-20 30-40 30-40 40-50 20-30 GaF 3 0 0 10-20 20-30 20-30 AI203 60-70 60-70 110-120 100-110 90-100 AIF 0 40-50 80-90 60-70 50-60 A, 10CC12F2-1002, 0", + " bias superimposed at the wafer position ( 13.56 MHz). Pumping is accomplished through a very high conductance pump manifold linking the process chamber to a 1000 1 s- ~ turbomolecular pump. ECR is provided by a plasma source of the multipolar tuned-cavity design. The microwave cavity plasma source has been described extensively elsewhere [12] and only a brief description is given here. The resonant cavity is a brass cylinder of 17.8 cm inside diameter terminated at the top 196 by an adjustable short (Fig. 13). A variable-length launching probe enters the resonant cavity at the side, impressing microwave energy from the magnetron-waveguide assembly to an evacuated quartz \"cup\" of 100 mm diameter where the plasma resides. The brass resonant cavity is at atmospheric pressure. Within the baseplate are eight high strength rare earth magnets which produce the B field level of 0.0875 T necessary for resonance. Cyclotron resonance occurs on an ECR surface within the quartz cup. In as much as large solenoidal magnets are not used, there is no degrading B field to accelerate ions to the wafer position coulombically" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000173_eumc.2003.177571-FigureI-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000173_eumc.2003.177571-FigureI-1.png", + "caption": "Fig. I . Dimension of the proposed antenna", + "texts": [ + " In principle, the bandwidth of a microstrip patch antenna can be enhanced by increasing the thickness of substrate, by using lower substrate permittivity constants or by introducing parasitic resonators around the vicinity of the original antenna (multiple resonant techniques) [6] . In this paper, a coupled-parasitic patch was placed above the power combining patch antenna, to provide multiple resonant frequency for both the differential and common mode. 10% operating bandwidth, 1.8GHz-1.99GHz (return loss<-lOdB) was obtained for this configuration. 11. ANTENNA STRUCTURE Fig. I depicts the configuration of the proposed broadband power combining patch antenna. Substrate FR4 with ~ ~ 4 . 6 and height=1.6 mm is used for both the upper and bottom layer. At the top layer, a coupled-parasitic patch with the dimension of Wl=24.2 mu, L1=54.1 m m Gl=1.5 mm is used. Between the gap G1, a 100 ohm chip resistor is placed at distance LRI= 21.1 mm from the antenna edge. Such geometry is chosen because the symmetrical properties of this coupled-parasitic patch can provide a second resonant frequency for both commou and differential mode excitation" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001913_6.2017-1745-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001913_6.2017-1745-Figure2-1.png", + "caption": "Figure 2. Notation and reference frames used for the quadrotor dynamics.", + "texts": [ + " The quadrotor\u2019s translational dynamics in V and the rotational dynamics in C can be written as v\u0307v = \u2212 [ \u03c8\u0307E3 ] \u00d7 vv + gE3 + F v m (3a) R\u0307 = R [\u03c9c]\u00d7 (3b) \u03c9\u0307c = \u2212J\u22121 [\u03c9c]\u00d7 J\u03c9 c + J\u22121\u03c4 c (3c) where g is the gravity constant, m is the mass of the vehicle, J is the inertia of the vehicle, E3 = [0, 0, 1] T , [x]\u00d7 = 0 \u2212x3 x2 x3 0 \u2212x1 \u2212x2 x1 0 where x = [x1, x2, x3] T , and F v and \u03c4 c are the applied force and torque expressed in V and C, respectively. The applied force F v is approximated as F v = \u2212R\u03b8\u03c6E3TM (4) where TM = \u22114 i=1 fi, R\u03b8\u03c6 = Rn2 (\u03b8)Rn1 (\u03c6), and fi is the thrust created by the ith propeller as shown in Figure 2. We take fi = KT W\u0303 2 i (5) where W\u0303i = Wi \u2212Wmin, Wi is the PWM signal fed to the ESC driving ith motor, Wmin is the minimum pulse width to start the motor, and KT is the thrust constant. The torque \u03c4 c is approximated as \u03c4 c = l\u221a 2 (f2 + f3 \u2212 f1 \u2212 f4) l\u221a 2 (f1 + f3 \u2212 f2 \u2212 f4) K\u03c4 ( \u2212W\u0303 2 1 \u2212 W\u0303 2 2 + W\u0303 2 3 + W\u0303 2 4 ) 4 of 11 American Institute of Aeronautics and Astronautics D ow nl oa de d by C A R L E T O N U N IV E R SI T Y o n Ja nu ar y 12 , 2 01 7 | h ttp :// ar c. ai aa .o rg | D O I: 1 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000283_mop.29797-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000283_mop.29797-Figure6-1.png", + "caption": "Figure 6 Simulated surface current distribution of FSA and Koch FSA at (a) 5.2 GHz, (b) 5.38 GHz, (c) 6.38 GHz, and (d) 6.32 GHz. [Color figure can be viewed in the online issue, which is available at wileyonlinelibrary.com]", + "texts": [ + " The standing wave resulting from reflections at the farther end of the spiral arm increases the effective current concentration over the arm segment responsible for radiation. This could be seen in Figure 5. According to ring theory of spiral, at frequency 5.2 GHz the effective circumference of 57.69 mm (5c/5.2 GHz) will be responsible for the radiation [6]. At this length, current magnitude for Koch-FSA is much higher than that of simple FSA as shown in Figure 5. The simulated snapshots of current distribution for the two structures are shown in Figure 6 at frequencies at which |S11| dips are more pronounced. The polar radiation plots in xz plane for LSA, FSA, and Koch-FSA are depicted in Figure 7 at the same frequencies for which surface current snapshots are given. Evidently patterns for all the prototypes are nearly similar. A stable radiation pattern is observed for all resonant frequencies. The angular width or beam width of the proposed antenna is nearly constant with an average value of 44.68 in E-plane. A novel Koch inspired Fibonacci spiral antenna and simple Fibonacci spiral antenna have been proposed and investigated" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000894_piers.2017.8262402-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000894_piers.2017.8262402-Figure2-1.png", + "caption": "Figure 2. E-field distributions at 3.4 GHz in the DRA (a) due to port 1 (b) due to port 2.", + "texts": [ + " The plus shaped CPW transmission line at port 1 behaves like a magnetic dipole loop and provides strong coupling to the dielectric resonator [6]. As a result, HEy 11\u03b4 mode is generated in the DR due to port 1. During port 2 excitation, the plus shaped CPW line acts as an aperture and microstrip line behaves as a main feedline. As a result, HEx 11\u03b4 mode is generated in the dielectric resonator. The proposed structure achieved high isolation due to generation of two orthogonal modes in the DRA. Fig. 2 displays the E-field distribution in cylindrical DR at 3.4GHz due to port 1 and port 2. From Fig. 3 it is clearly shown that HEy 11\u03b4 and HEx 11\u03b4 mode is generated at 3.4GHz with port 1 and port 2 respectively. Figure 3 shows the different scattering parameter of the proposed MIMO antenna. From Fig. 3, it can be observed that the operating band for both port are almost same. The impedance bandwidth for port 1 and port 2 is 11.7% (3.2\u20133.6 GHz) and 11.4% (3.3\u20133.7 GHz) respectively and the isolation between the two ports is within the required band exceeds \u221228 dB" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003315_detc2005-84562-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003315_detc2005-84562-Figure2-1.png", + "caption": "Figure 2. Symbolic representation of a spur gear system", + "texts": [ + " The rigid driving shaft, loaded by an input torque, TM, and supported by bearings, is connected to the driving gear in the gear system. The driving gear(pinion) transfers motion to the driven gear. At the output side of this gear system, the driven gear(gear) is connected to an output load, TL, via a rigid shaft (driven shaft) supported by rolling element bearings. The rolling elements (bearings and supports) are assumed to be flexible, and the effect of their linear compliance in X ,Y and Z directions are included in the gear system dynamics. X \u2212Y \u2212Z coordinate system is an inertial frame of reference. Figure 2 depicts the symbolic representation of meshing gear model. Each gear mass is treated as a rigid body which can freely translate and rotate in space. Therefore, each mass processes six degree of freedoms which are defined as the (X ,Y, and Z) coordinates of center of mass and the three angles that define the orientation of the rigid body with respect to the inertial frame. In this paper, the Z\u2212Y \u2212X Eulerian angles [14], \u03c8, \u03c6 and \u03b8, are used to define orientation of a rigid body. On the other hand, linear positions of mass center is the position vector measured with respect to the Copyright c\u00a9 2005 by ASME rms of Use: http://www.asme.org/about-asme/terms-of-use Dow origin of inertial frames. To develop the dynamic model for our geared system, we define the body-fixed axes for a pinion and a gear, namely, x1 \u2212 y1 \u2212 z1 and x2 \u2212 y2 \u2212 z2 axes, respectively. The origin of body-fixed axes is located at the center of mass. The orientations of these axes are defined as \u03c81, \u03c61 and \u03b81 for the pinion and \u03c82, \u03c62 and \u03b82 for the gear, as shown in figure 2. The basic vector equations that describe the general motion of a rigid body are the followings. For the translational motion of the center of mass: Fext = m aCM (6) For the rotational motion of the body, written in Euler\u2019s angle form: {\u039b\u0308} = [B]\u22121 [ [ICM]\u22121[ Mext CM \u2212 \u03c9B\u00d7 [ICM] \u00b7 \u03c9B]\u2212 [B\u0307]{\u039b\u0307} ] (7) where m is mass of the body. aCM is the linear acceleration at the center of mass of the body. Fext is the sum of external forces that act to the body. \u03c9B is the absolute angular velocity of the body" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002203_0010-4655(91)90210-c-Figure17-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002203_0010-4655(91)90210-c-Figure17-1.png", + "caption": "Fig. 17. Effects of substrate thickness on the active impedance locus of a series-loaded array (a = 0.52 m, W = 0.26 m, b = 0.39 m, = 2.5, XL = 0.46 m, h = 0.01 m). d = 0.03 m, f (MHz): [200, 220, 240, 260, 280, 300, 320, 340i; d = 0.05 m, f (MHz): [200,220, 240, 260, 280, 300, 320, 340]; d = 0.20 m, f (MHz): [200, 220, 240, 260, 280, 300, 330, 360, 400].", + "texts": [ + " Schaubert / Infinite arrays of 2-D microstrip structures of the constant shunt current assumed in ref. [3]. and d on 0.05 A curves. For a printed dipole with Such assumption becomes less accurate for thick- thin substrate, a series RLC model is usually asness of the order of 0.20 A or larger (as shown in sumed. This model predicts that the locus will go table 3), particularly when operated at a frequency from being capacitive to inductive as observed in far from resonance. Figures 17 and 18 show the fig. 17. However, these models are no longer accuimpedance loci for the series-loaded array and for rate for antennas with a substrate as thick as the shunt-loaded array as a function of thickness, don 0.20 A. It was suggested earlier that this To understand these results, it is helpful to review shunt-loaded array behaves as an array of topsome common models assumed when dealing with loaded monopoles. Thus, one would expect (as printed dipoles and microstrip antennas. Usually, seen in fig. 18) the impedance locus to follow the a microstrip antenna with a thin substrate is mod- trend of a dipole rather than that of a typical elled using a parallel RLC circuit" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001804_8.3540-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001804_8.3540-Figure7-1.png", + "caption": "FIG . 7. Shear stresses.", + "texts": [], + "surrounding_texts": [ + "234 J O U R N A L O F T H E A E R O N A U T I C A L S C I E N C E S \u2014 M A R C H , 1 9 5 6\nIf the ith element is not an axially loaded member the above equation can still be applied, provided it gives the correct relationship between force and distortion. If the Ramberg-Osgood equation does accurately repre sent this relationship for all members, then the D matrix is diagonal, and\nDa = eu/Fi\nTherefore,\nDu = DLti[l + (3/7)(Ft/FBl)*-1] (15)\nwhere DLii = ( \u2014 ) \\AEJt\nI t is clear by comparing Eqs. (5) and (15) that\nu = (3/7) I V - 1\nIn matric form\n0 = (3/7) F ^ \" 1 (16)\nwhere YD = [7,1. Substituting Eq. (16) in Eq. (14) gives\nF ( m ) = Yw + [nCRY(i)Dn-i + j ] - i X\n[YL ~ CRY(i)\u00bb\"-iYw - Y(i)] (17)\nwhere CR = (3/7) FR^CFR (18)\nIn order for Eq. (17) to apply, the members of the structure must satisfy the following conditions:\n(1) Each member of the structure must have a loaddistortion diagram that can be approximated satisfac torily by the Ramberg-Osgood equation over the range \u00a3>f loading to which the member is subjected.\n(2) The load-distortion diagram must be symmetric with respect to the origin.\n(3) The Ramberg-Osgood exponent n must be the same for all the members.\n(4) The flexibility matrix must be diagonal. In other words, any given element distortion must depend only on the corresponding element force.\n(5) During the loading process there must be no un loading of any member after it has yielded.\nIf the above conditions are met one can utilize Eq. (17) in the analysis of structures that have members carrying other than axial load. These conditions are somewhat restrictive; however, it is possible to begin again with Eqs. (7) or (14), choose a different form for the matrix <\u00a3, and derive equations similar to Eq. (17) which apply to other cases.\nInitial Approximation\nIt is natural to choose the linear distribution YL as the initial approximation to the true distribution F. But if solutions are being obtained for successively in creasing values of external load, the best approximation to Y for a given loading could be the true distribution obtained for the preceding case.\nConvergence\nNo conditions for the convergence of Eq. (17) are given; however, convergence is very rapid in the ex ample that follows.\nThe nonlinear equations can have more than one set of real roots. The significance of the possible existence of these multiple roots needs further study to be under stood; however, the roots converged upon in the ex ample are intuitively reasonable.\nApplication\nThe simplified swept-wing structure shown in Fig. 1 was analyzed accounting for plastic behavior. Fig. 2 shows the numbering of element forces and distortions. Element forces in spar caps and rib caps were taken as average axial loads; element forces in shear panels were taken as shear flows. Spar web stiffeners were assumed rigid, while the outboard bulkhead and bulkheads a t the side of the fuselage and at the plane of symmetry were assumed rigid in their own planes, but free to warp. Structure and loading were symmetric about the air plane plane of symmetry. Bending loads PE were ap plied to the structure at the outboard end.\nFig. 3 shows the loads assumed to have been acting on the triangular shear panels when unit shear flows were acting on the mutually perpendicuar sides. I t was assumed that a uniformly distributed tensile load acting along the hypotenuse could be replaced by con centrated loads at the adjacent corners as shown.\nThe structure had five redundants numbered as in\nFig. 4. The figure also shows the unit vertical dummy load applied at the center of the outboard bulkhead for the purpose of computing the vertical deflection at that point.\nTable 1 gives sizes of members, elastic moduli, and the forces FR causing stresses equal to the 0.7.E secant yield stress SL. The table also gives the linear flexibility matrix DL.\nFig. 5 shows the stress-strain curves used in the analy sis. For these curves E = 107 lbs. persq. in., G = 4 X 106 lbs. per sq. in., S\u00b1 = 45,000 lbs. per sq. in. for ten sion, 27,000 lbs. per sq. in. for shear. The RambergOsgood exponent n was equal to 20.\nTable 2 gives the matrices / , F0> and fA. These matrices were computed by the method of reference 4, but they can also be derived directly from the equations of statics. Omitted elements are zero.\nA program was designed for the solution of Eqs. (8), (9), (18), (17), and (10) on the high-speed digital com puter, and the required calculations were made. Table 3 presents the resulting load distributions and deflec tions for two values of external load. Loads 1 to 12 are in lbs., loads 13 to 24 are in lbs. per in., deflections are in inches. The method converged to four significant figures in two iterations for the 160,000-lb. load. For the 200,000-lb. load six iterations were required to give convergence to four significant figures. In both cases the initial approximation was the linear distribution. For the higher loading condition the reduction in load\nD ow\nnl oa\nde d\nby U\nN IV\nE R\nSI T\nY O\nF O\nK L\nA H\nO M\nA o\nn Ja\nnu ar\ny 27\n, 2 01\n5 | h\nttp ://\nar c.\nai aa\n.o rg\n| D\nO I:\n1 0.\n25 14\n/8 .3\n54 0", + "M A T R I C S O L U T I O N O F C E R T A I N N O N L I N E A R P R O B L E M S 235\nTABLE 1 TABLE 3\ni\n1.1\n2\n3 4\n5\n6\n7\n8\n9\n10 n 12\ni\n13\n14\n15\n16\n17\n18\n19 20\n21\n22\n23\n24\nA i ( i r > . 2 )\n4 J\n4\n1\n1\nt i ( i n . )\n.080\n\u2022 1\n. .060\n\u2022c )k0\nFR ( I t ).)\n180,000\n' 180,000\n^5 ,000\n45 ,000\nFR ( l b . / i n . )\n2160\n1\n2160\n101 30\n1 0 6 DL\ni\n1,00\n.50\n1.00\n.75\n.50\n1.00\n.50\n1.00\n.75\n.50\n4 . 0 0\n4 . 0 0\n1 0 6 DL\ni\n5000\n1875\n3125\n5000\n1875\n3125\n1250\n625\n1250\n937.5\n625\n2500\nE = 107 l b . / i n . 2 G = 4 x 106 l b . / i n . 2\nTABLE 2\ni\n2 5\n7 10\n15 1 8 2 0\n2 3\n2 4 5 7\n9 10 1 1 12\n1 5 1 8 20 22\n23 24\n3\n1\n1 2\n5\nf. .\n.500 - .500 - .500\n.500 .025 .025 - .025 - .025\n.500 -.625 -.500 - .500\n.625 .500\n.833.. . - .833. . . - .025 -.025\n.025\n- .04166. .\n.025 .o4 l66 . .\ni\n1 3 4 6 8\n9 1 1 12\n1 3 16\n19 2 1 22 2k\n9 12\n17\nk 1 1 Ik\nJ\n3\n1 \u2022 3 k k k\n5 5\nhi .500\n- .500 - .500 - .500\n.500 .500\n- .666. . . .666...\n.0125 .0125 -.0125 - .0125\n.033.. . -.045833\n15.00 20.00\n1.00 -15.00 -20.00\n1.00\ni\n1 2 3 6 7 8\n13 1 6\n1 9 2 1 2 4\ni\n1 2 6\n7 1 1 12\n1 3 16\n19 2 1 2k\nJ\n1\n1\nJ\n1\nL\nF 0. .\n1J -240,000 -256,000 - 80,000\n240,000 256,000\n80,000 - 2,000 - 2,000\n2,000 2,000 2,000\n\\ i\n-2.00 -4.40\n2.00 4.40 \u2022\u2022\n- l . o o\n1.00\n- .025 - .025\n.075 - .025 - .025\nD ow\nnl oa\nde d\nby U\nN IV\nE R\nSI T\nY O\nF O\nK L\nA H\nO M\nA o\nn Ja\nnu ar\ny 27\n, 2 01\n5 | h\nttp ://\nar c.\nai aa\n.o rg\n| D\nO I:\n1 0.\n25 14\n/8 .3\n54 0", + "236 J O U R N A L O F T H E A E R O N A U T I C A L S C I E N C E S \u2014 M A R C H , 1 9 5 6\n\u2014 TEST o ANALYSIS X SUCCESSIVE^ APPROXIMATIONS\n2 3 4 5 LOAD IN MEMBER KIPS\nFIG . 8. Comparison with experimental results.\nat the center of the rear spar cap resulting from plas ticity was ten per cent. Note that all axial loads were given at element centers; axial loads at other points can be obtained from statics. Figs. 6 and 7 compare the linear and nonlinear distributions of axial and shear stresses for the 200,000-lb. load.\nFig. 8 presents the load distribution in a simple struc ture computed by the present method compared with the experimental results of reference 6. The figure also\nshows the convergence. In this case the Ramberg-Osgood exponent n was equal to 40.\nCONCLUSIONS\nGeneral equations have been presented for the stress and deflection analysis of statically indeterminate structures that are characterized by the dependence of member flexibility upon member load. For cases where the flexibility matrix is diagonal, the Newton-Raphson method has been applied to give an iterative form for the solution of the nonlinear equations. For plastic structures this form was particularized through the in troduction of the Ramberg-Osgood relationship, and in an actual problem the iterative application of the form was shown to be rapidly convergent.\nREFERENCES\n1 Wilder, Thomas W., An Analysis of Statically Indeterminate Trusses Having Members Stressed Beyond the Proportional Limit, NACA T N 2886, 1953.\n2 Langefors, B., Analysis of Elastic Structures by Matrix Trans formation with Special Regard to Semimonocoque Structures, Jour nal of the Aeronautical Sciences, Vol. 19, No. 7, pp. 451-458, July, 1952.\n3 Wehle, L. B., and Lansing, W., A Method for Reducing the Analysis of Complex Redundant Structures to a Routine Procedure, Journal of the Aeronautical Sciences, Vol. 19, No. 10, pp. 677-684, October, 1952.\n4 Denke, P. H., A Matric Method of Structural Analysis, The Proceedings of the Second U.S. National Congress of Applied Mechanics, 1955.\n5 Ramberg, Walter, and Osgood, William R., Description of Stress-Strain Curves by Three Parameters, NACA T N 902, 1943.\n6 Steinbacher, F. R., Gaylord, C. N., and Rey, W. K., Method for Analyzing Indeterminate Structures Stressed Above Proportional Limit, NACA T N 2376, 1951.\nThe Transverse Curvature Effect {Continued from page 224)\nFor the case of Pr = 1, and zero pressure gradient Crocco's integral [Eq. (10)] gives\nF0=^ lim (v - / 0 ) + (y - \\)M*tf(0)\nFi = ~ (J^ M\u201e* + p'Weo) + (y- DM/ X\nTherefore, for incompressible flow the coefficient of the first-order correction to the displacement thickness is simply \u2014 /i(o\u00b0). But for incompressible flow\nhit) = fo'in) f [A/\u201e(\u201e) + /,(,) ]d* Jo\nf fo\"fidn so that / i ( \u00bb ) = P [Aln(V) + Jr(V)]dV\nJo Jo\nD ow\nnl oa\nde d\nby U\nN IV\nE R\nSI T\nY O\nF O\nK L\nA H\nO M\nA o\nn Ja\nnu ar\ny 27\n, 2 01\n5 | h\nttp ://\nar c.\nai aa\n.o rg\n| D\nO I:\n1 0.\n25 14\n/8 .3\n54 0" + ] + }, + { + "image_filename": "designv6_24_0003392_bf01531397-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003392_bf01531397-Figure4-1.png", + "caption": "Fig. 4.", + "texts": [ + " This gives the basis to consider that the fatigue strength of the nut thread is at least not less than that of the bolt. The analysis of the shear stress distribution at the base of the turns shows that the static strength safety factor for the first and last turns is 2.93 and 1.31, respectively. Attainment of an average stress level of T = 0.5o t along the section was taken as the limiting stressed state leading to shearing of the turn. In our case, this is 410 MPa. The figured nut has zones with raised level of stresses. These are the corner transition and the cross-section near the contact surface (Fig. 4). The results of calculations of the stresses in the maximally loaded cross-section located approximately over the turn 20 were used for determining the static strength safety factor, i.e., K = o0.2/o~I TABLE 2. Main Calculated Stresses in the Most Stressed Thread Root of the Tie Bolt Point i n Stresses, MPa l Fig. 2 o, [ o, o~ i 1 811,7 \" ' ( 2 1154,0 2~0.:3 31 ! 3 I ] 71,1 :~19,3 65.4 4 108(1,0 2673) 39,9 5 799. I 190,.I ~,~,') Fig. 5. i! J ~ MPa Stresses at the bolt head periphery: i) calculation; 2) experiment" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002816_icmmt49418.2020.9386952-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002816_icmmt49418.2020.9386952-Figure3-1.png", + "caption": "Fig. 3 the simulation model in HFSS.", + "texts": [ + " As discussed earlier, this MNG structure can be used to enhance the isolation between fuze antennas. 978-1-7281-5733-7/20/$31.00 \u00a92020 IEEE 20 20 In te rn at io na l C on fe re nc e on M ic ro w av e an d M ill im et er W av e Te ch no lo gy (I C M M T) | 97 8- 1- 72 81 -5 73 3- 7/ 20 /$ 31 .0 0 \u00a9 20 20 IE EE | D O I: 10 .1 10 9/ IC M M T4 94 Authorized licensed use limited to: University of Canberra. Downloaded on May 20,2021 at 03:29:12 UTC from IEEE Xplore. Restrictions apply. III. ISOLATION WITH/WITHOUTMETAMATERIAL The isolation simulation model in HFSS is shown in Fig. 3 where the waveguide slot array operates in X band and the Ku band fuze antennas are set in the center of the X band array in order to reduce the influence of the fuze antenna's far filed pattern after installing the radome. During the simulation, the metamaterial array contains 8*13 unit cells which satisfy the periodic boundary condition in the metamaterial unit cell simulation. To prove the isolation enhancement capability of the mu-negative metamaterial, the simulated S21 is presented in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003727_j.2042-3306.1989.tb02089.x-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003727_j.2042-3306.1989.tb02089.x-Figure3-1.png", + "caption": "Fig 3. The characteristic angle-angle diagram inregrated with evenrs of the stride cycle", + "texts": [ + " The minimum value for tarsal flexion had a slightly higher coefficient of variation suggesting this value shows some variability. Measurement of the areas of the angle-angle diagrams for different strides indicated moderate variability with coefficients of variation below 10 per cent except for horses A, B, and F. Minimum and maximum values for tarsal height recorded from the caudal view exhibited low variability. The measurements of limb abduction and adduction had high variability as shown by the large standard deviations. Figure 3 represents the characteristic angle-angle diagram for the stifle angle versus the tarsal angle, integrated with outlines of 52 EQUINE VETERINARY JOURNAL the stride cycle. In the first part of the stance phase the hindlimb is acting as an inclined strut that is opposing progression. Retraction of the limb results in forward progression. Flexion of both the stifle and the tarsus occurs until the mid-stance phase. Past the mid-stance phase the stifle still flexes but tarsal extension occurs. In the horse the stifle flexes throughout the stance phase, whereas the tarsus flexes initially and then extends" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000568_12.859243-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000568_12.859243-Figure1-1.png", + "caption": "Figure 1. Typical CubeSat Kit frame (from Ref.6).", + "texts": [ + " A major problem with stellar imaging through the atmosphere is atmospheric turbulence: for large terrestrial telescopes resolution is limited by the atmosphere rather than the size of the telescopes aperture2. The obvious solution is to put large telescopes in orbit, out of the atmosphere. However, large space telescopes must have large and heavy primary mirrors3. As a result, high resolution space telescopes are very expensive systems and delivering them to orbit incurs large further expenses. Very small satellites, known as micro-satellites and nano-satellites have become increasingly popular in recent years. One example of such small satellites is the CubeSat4, , 5 6 format (Fig. 1): satellites based on this format are limited to a cube-shaped enclosure of 10 10 10 cm or multiples of such cubes (for example, 10 10 30 cm); a satellite encased in such an enclosure may unfold or open after delivery to the orbit to form a larger structure. This compact size and low weight allows these satellites to \"hitch-hike\" together with a larger satellite, significantly reducing the cost of their launching and delivery to orbit. While adopting a micro-satellite format can alleviate costs, one cannot fit a large mirror (i" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001212_6.2008-5955-Figure10-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001212_6.2008-5955-Figure10-1.png", + "caption": "Figure 10. Example of Implemented Vortex Lattice Method", + "texts": [ + " This section provides an overview of the current analysis capabilities if the framework. 12 of 23 As shown before, the aerodynamic module provides different levels of analysis complexity to choose. The simplest level provides parametric aerodynamic analyses such as those present in other state-of-practice programs. Aircraft lift, drag and stability derivatives are calculated based on parametric and semi-empirical formulations augmented with a potential aerodynamics flow non-planar multiple lifting surfaces vortex method developed in-house as shown on Figure 10 which uses the simplified geometric model shown in Figure 6(c). Drag calculations include Lift-induced, parasite, and transonic wave drag effects. The induced drag is calculated based on the vortex method Trefftz plane downwash. The Oswald efficiency factor and downwash effects are obtained from the vortex method and corrected with parametric technology models.32 Parasite drag is calculated using a detailed aircraft components build-up,33 taking into consideration viscous separation and mutual interference effects, with a skin friction formulas modeled by Sommer & Short formulation34)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000083_ir-03-2018-0042-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000083_ir-03-2018-0042-Figure6-1.png", + "caption": "Figure 6 Ankle joint description", + "texts": [ + " Elastic material is also used to augment the wearer\u2019s leg during walking mode. Rotation movement is applied by using elastic material also, which is attached to the ankle shoes plate. The wearer is attached to the wearable vehicle through a flexible tie. Force sensors locations were selected by taking into consideration the static pressure distribution of the human feet along the foot direction when a person is standing. The wearable vehicle has only two-force sensors that are installed in the feet module instead of using four sensors. The detailed design is shown in Figure 6. The components of the fast motion mode which allow the system to be used as a vehicle, including the front wheel mechanism, rear wheels mechanism and seatingmechanism. Front wheels are fixed onto the shank of the wearable vehicle. These links can be adjusted by telescopic columns that are fixed in different positions to adapt to the two modes of motion by using a securing mechanism as shown in Figure 7. The castor wheel is fixed by axial shaft and two ball bearings on both sides to decrease friction and wearing" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003635_s12206-009-0321-8-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003635_s12206-009-0321-8-Figure1-1.png", + "caption": "Fig. 1. Finite element model of sub-frame.", + "texts": [ + " And then the effects of some parameter variation on the modal characteristics of the vehicle sub-frame are investigated. Most medium class passenger cars usually have adopted #-type front sub-frame. The principal purpose of using a front sub-frame is to isolate vibration and harshness from the rest of the body. \u2020This paper was presented at the 4th Asian Conference on Multibody Dynamics(ACMD2008), Jeju, Korea, August 20-23, 2008. *Corresponding author. Tel.: +82 2 2220 0446, Fax.: +82 2 2293 5070 E-mail address: hhyoo@hanyang.ac.kr \u00a9 KSME & Springer 2009 Fig. 1 shows the finite element model of #-type front sub-frame. The #-type front sub-frame consists of total eight subparts, that is, upper and lower cross member, left and right side members, A and G-point brackets. An equivalent model of the vehicle sub-frame that only consists of a beam element is constructed based on the FOA technique, as shown in Fig. 2. The #-type front sub-frame is a symmetric structure except for front and rear engine mounts. So, by dividing the finite element model into seven pairs of symmetric sides, a total of fourteen areas, the section properties, such as cross-sectional area, area moment of inertia and shear coefficient are extracted for constructing an FOA equivalent model" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002792_s10817-013-9280-y-Figure13-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002792_s10817-013-9280-y-Figure13-1.png", + "caption": "Fig. 13 Collapsed routing graph derived from the routing graph in Fig. 11 when the starting and ending domain intervals are A and G respectively", + "texts": [ + " Thus Algorithm 4 first computes the collapsed routing graph by collapsing starting states D+ s , D\u2212 s into one single node, Ds, and ending states, D+ e , D\u2212 e into one Algorithm 4 Path finding Input: RG, the routing graph Is, starting instance Ie, ending instance Output: P, the path that leads from Is to Ie, if one exists Identify the domain interval Ds of Is Identify the domain interval De of Ie if Ds == De then P = (Ds, \u03c3s, \u03bbs, \u03bbe) else Compute the collapsed routing graph CRG P = A\u2217(CRG, Ds, \u03bbs, De, \u03bbe) end if single node, De in the routing graph. Figure 13 shows the collapsed graph derived from the routing graph in Fig. 11. Then we feed the A\u2217 algorithm with the collapsed routing graph. In the Appendix we define the path-cost function used in A\u2217 to estimate the distance to the goal and prove that it is not greater than the exact distance as required by the A\u2217 algorithm. Figure 14 shows two paths with minimum variant parameter arc length computed by Algorithm 4 that solves the reachability problem for the problem in Fig. 4. The starting instance Is \u2208 A is defined by \u03bbs = 5 and \u03c3 = {+1,+1}, and the ending instance Ie \u2208 G defined by \u03bbe = 5 and \u03c3 = {\u22121,\u22121}" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000865_iwem.2015.7365050-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000865_iwem.2015.7365050-Figure4-1.png", + "caption": "Fig. 4. The designed two-dimensional beam scanning antenna array (a) 3D view (b) side view", + "texts": [ + "3(a) shows that, over the range of 93 GHz~95 GHz, return losses of all the input ports are under -20 dB, while the amplitude differences of the output ports are within 0.3dB, and the phase differences are less than 5\u00b0. \u2162. THE MULTI-BEAM ANTENNA ARRAY According to the grating lobe restrain condition, the array spacing dx in the x direction is 1.6mm, and dy is 2.4mm. However, the width of a SIW structure is 1.086mm, so a transition between the beam-scanning network and the antenna array is required, as shown in Fig.4 (a). LTCC technology As shown in Fig.4, the multi-beam antenna array is composed of 18 LTCC layers. 3 LTCC layers are used to form a SIW structure. For clear description, the LTCC substrate is divided into six SIW layers, as shown in Fig. 4 (b). The 2D beam-scanning network is at the following five layers. The 2 \u00d7 4 SIW slot antenna array is on the top layer. Signals transfers between adjacent layers by coupling slots. There are two types of coupling slots: one is used for reverse coupling, and another is used for synthetic coupling, as shown in Fig.4 (a). The simulated |S11|, 3D beams of the designed multi-beam antenna array are shown in Fig.5 and Fig.6. Fig.5 shows that the simulated return loss of all the ports is under -20dB over the range of 93GHz ~ 95GHz. The designed multi-beam antenna can realize eight beams on two dimensions and coverage from -60\u00b0 to 60\u00b0 in elevation plane, as shown in Fig.6. The gains of all beams are over 11.4 dBi. \u2163. CONCLUSIONS This work proposes a SIW 2D beam-scanning network working at 93 GHz ~ 95 GHz. Compared with other works, multiple phase shifters can increase the flexibility of the beams" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003028_asemd49065.2020.9276314-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003028_asemd49065.2020.9276314-Figure6-1.png", + "caption": "Figure 6. Diagram of experimental device.", + "texts": [ + " The reason is that the exciting current is maximized, and the rotor claws are locally saturated, so air gap magnetic density has a large distortion, and the distorted air gap magnetic density brings a large distortion of EMF. In Fig. 5 (b), output torque waveform under exciting current 4A is shown, and the average torque goes down to 20N m. However, output torque ripple is obviously improved after the exciting current is reduced to 4A. IV. EXPERIMENTALL VERIFICATION In order to verify the effectiveness of new topology structure, test platform of hybrid exciting claw-pole machine, as shown in Fig. 6. No-load and load experiments under state of prototype power generation are tested, and the EMF and output current waveform could be given through oscilloscope. Fig. 7 shows the relationship of exciting current and amplitude of EMF at different rotating speeds. At condition of exciting current 1A and rotational speed 4000rpm, results are verified by experimental measurement as shown in Fig. 8. By Fourier decomposition, the amplitude of fundamental wave component is 23.1V and the fifth harmonic component is relatively large" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003998_s00202-006-0054-y-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003998_s00202-006-0054-y-Figure5-1.png", + "caption": "Fig. 5 SSDR machine portion (see Fig. 3a): particular of the side view (bottom) and section A-A, front view (top); some teeth, their tooth heads, slots (with parallel sides) and rotor PMs are shown. The line is the trace of the cylindrical surface considered for the 2D FEM analyses", + "texts": [ + " A hollow shaft presents two sets of skewed grooves, with opposite slope, on the outside surface, required for the mechanical coupling with the half rotors. On its internal surface, the hollow shaft is equipped with one set of grooves (parallel to the shaft axis), needed for the mechanical coupling with the main shaft. The hollow shaft is driven to axially slide along the main shaft. The skewed grooves transform the rectilinear motion of the hollow shaft into an angular displacement between the two half-rotors, provided that the half-rotors are maintained in their axial position by thrust bearings. A portion of the machine is shown in Fig. 5. Each stator slot has parallel sides; each tooth has a tooth head. Now an equivalent two-dimensional analysis is carried out, referred to a cylindrical surface, considered at half radial size l; the trace of this surface is line in Fig. 5 top. The expressions of the generic harmonic of the flux linkage and of the phase e.m.f. will be derived, starting from easily valuable quantities. These expressions include both the winding factor kw and a factor taking into account the field regulation, called regulation factor, and dependent on the relative rotational displacement between the two rotors. As will be shown in the following, for a SSDR tooth coil machine, the basic no-load quantity cannot be the tooth flux. In fact, at first let consider the disposition in which the two rotors are aligned each other, in the generic angular position \u03b8r with respect to the stator (Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003937_978-1-4471-5104-3_5-Figure5.24-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003937_978-1-4471-5104-3_5-Figure5.24-1.png", + "caption": "Fig. 5.24 Thyristor phase controlled ballast load", + "texts": [ + " The control technique used to maintain the generated voltage and frequency at their rated values is used to keep the total load connected to the machine at a near constant value using a ballast load as shown in Fig. 5.23. Since the terminal voltage under this condition is a constant value, voltage sensing is used to control the ballast load. The ballast load can be implemented in several ways [28\u201332]. One way of obtaining a variable load is to use a resistor with two anti-parallel thyristors operating in phase control mode as shown in Fig. 5.24. By changing the firing angle b, the fundamental value of the current going through the resistor-thyristor circuit can be controlled. When b \u00bc 0 there is a full current impressed through the resistor-thyristor circuit, i.e., maximum load. When b \u00bc 180 , the current through the resistor-thyristor circuit is zero and a programmable ballast load is possible for values of b in between 0 and 180 . Of course, as b is increased, the displacement factor of the resistor-thyristor circuit increases, thus absorbing reactive power" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002888_s11665-018-3265-2-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002888_s11665-018-3265-2-Figure2-1.png", + "caption": "Fig. 2 Shear deformation of a unit cell in ECSEE Channel L2: (a) the spiral path curve, (b) the amplifier unit cell, (c) graphic plane unit deformation, (d) the angle relation in torsion, and (e) the ECSEE-produced sample cross section", + "texts": [ + " With the exception of the parallel shear plane direction lines, all directions and lengths of direction lines could change with the progressive deformation. Simple shear deformation consists of a series of parallel sliding layers formed by shear slip. In view of the significance of the ECSEE Channel L2, the elliptical cross section is approximately simplified into a round cross section. Deformation flow of the ECSEE-induced material appeared along the space spiral curve during the plastic deformation (indicated in Fig. 2). The spiral track of ECSEE deformation is designated in theCartesian coordinates x\u2013y\u2013z. The marked point on the x\u2013y plane and a unit cell (blue block in the figure) are selected as the research object. The amplifier unit cell as the study focus of the micro-torsion of torsional angle du is depicted in Fig. 2(b). The initial unit cell ABCD\u2013EFGH transforms into the oblique prism A\u00a2B\u00a2CD\u2013E\u00a2F\u00a2GH under the driving of simple shear. Shear angles a and b are initiated by the constructive change of the unit cell. As far aswe can see, the angle a is greater than b due to the torsion characteristic (in Fig. 2(b)). The different stress and strain distributions in the central and peripheral positions on the cross section of ECSEE-induced sample can be unambiguously observed. Previous studies have verified the non-uniformity of deformation in ECSEE (Ref 4, 5). To investigate the influence factors of ECSEE strain, geometric analysis was adopted with the simplified model of a round cross section instead of the elliptic cross section. As it is indicated in Fig. 2(b) and (c), the rectangle ABCD transforms into the parallelogram A\u00a2B\u00a2CD under the application of shear deformation. Assuming the coordinates of point A (x1, z1), and point A\u00a2 (x2, z2), Z2 \u00bc Z1; x2 \u00bc x1 \u00fe z1c tan a \u00f0Eq 1\u00de As shown in Fig. 2(d), the geometric relationship is given by c tan a \u00bc tan 90 a\u00f0 \u00de \u00bc AA0 AD AA00 AD \u00bc duD1 2z1 ; \u00f0Eq 2\u00de where D1 is the diameter of the as-received sample or the die channel. Therefore, the expression for the strain can be obtained as (see Fig. 2(c)) e \u00bc ln DB0 DB \u00bc ln ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi x21 \u00fe z21ctan 2a\u00fe z21 p ffiffiffiffiffiffiffiffiffiffiffiffiffiffi x21 \u00fe z21 p \u00f0Eq 3\u00de Using formulas (1) and (2), we can obtain e \u00bc ln DB0 DB \u00bc 1 2 ln 1\u00fe duD1\u00f0 \u00de2 4 x21 \u00fe z21 ! \u00bc 1 2 ln 1\u00fe d2u sin2 w 4 ; \u00f0Eq 4\u00de where w is the angle between the line of an arbitrary point and the center point and the central axis. As it can be seen from the above formulas, the strain value of an arbitrary point on the assumed round cross section of ECSEE-induced sample is associated with the angles w and du", + " When the diameter of the cylindrical sample is determined, the length parameters of Channel L2 act a decisive role in the torsion shear deformation (Ref 2, 4, 5). Additionally, the ratio of the major-axis and minor-axis lengths m is also associated with the elliptical section. In previous investigations, the strain distribution of the elliptical section differed from that of the conventional round section of as-formed billet (Ref 4, 5). As it is shown in Journal of Materials Engineering and Performance Fig. 2(e), the cross-sectional area can be divided into Ireinforced region (red-filled), II-disappearing region (blue slash-filled), and III-normal region (green-filled). Compared to the round cross section, Region I represents a strong torsional torque. As far as the strain value of points on the cross-section boundary is concerned, the maximum value is obtained at the point N and the minimum at the point M with ignoring the point in the central area. The strain value of an arbitrary point, as shown in formula (4), is also multiplied by the coefficient related to the variable m greater than 1" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000298_1.1807415-Figure11-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000298_1.1807415-Figure11-1.png", + "caption": "Fig. 11 Pressure decrease within pocket area in hydrostaticoperating bearing", + "texts": [ + "asmedigitalcollection.asme.org/ on 08/07/20 higher by 13.2% and 8.3%, respectively, than that by VTEXPRESS. Generally speaking, under the hydrostatic-operating condition, the load capacity of the bearing is proportional to the difference of pressures between the bottom and top pockets. By checking the pressure distribution in each pocket, it is found that in the top pocket area ~pocket 1! there is an obvious decrease in the pressure field from the pocket inlet pressure ~post-orifice pressure!. Figure 11 shows this phenomenon in hydrostatic-operating condition. It can be seen that the pressure values on most areas in the top pocket are less than the pocket inlet pressure. This probably is because the clearance of the bearing in this area increases more than the other areas after applying the eccentricity. Although there is also a similar mechanism in the bottom pocket as well as the other two pockets, the simulation results show this influence is more evident right in the pocket opposite to the eccentricity direction" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002538_ol.22.000582-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002538_ol.22.000582-Figure1-1.png", + "caption": "Fig. 1. (a) Conventional imaging mode where the laser illumination (LI) is incident upon a lens (L) and focused onto an observation plane (OP). A superresolution filter (SF) is used to increase the resolution of the system. The confocal mode (b) is composed of an imaging lens (IL) and a collector lens (CL) to image onto a point detector (PD). PS, point source.", + "texts": [ + " 1997 Optical Society of America The ability to improve the resolving power of optical systems beyond the limits imposed by diffraction has long attracted considerable interest1 for theoretical reasons and also because of the enormous practical benefits it would bring to such diverse applications as astronomy,2 optical data storage,3 and confocal scanning microscopy.4 In optical-superresolution techniques one typically places a filter at the exit pupil of the system to modify the incident light properly (see Fig. 1). The superresolved point-spread function can be characterized by the spot size, Strehl ratio, and field of view. The normalized spot size G, or simply the spot size, gives a measure of the resolution and is defined as the ratio h1yh1 A, where h1 and h1 A \u00f8 3.8325 give the positions of the first zeros of the superresolved and the Airy disk patterns, respectively. The adimensional coordinate h localizes a point at the observation plane and is defined as h pDrylf , where r is the usual radial coordinate, D is the diameter of the lens aperture, l is the wavelength of the radiation, and f is the distance from the aperture to the observation plane", + " Although it is not obvious that there should be a fundamental limit to the performance of optical-superresolution systems, we know of no method that combines arbitrary values of Strehl ratio with a specified level of resolu- 0146-9592/97/090582-03$10.00/0 tion. We show that this is indeed impossible and that there is an upper bound to the values of Strehl ratio that can be achieved for any degree of resolution, whatever the f ield of view and independently of the technique used to design the superresolution f ilter. This result establishes the ultimate goal that limits the best performance attainable by any method, existing or yet to be developed. Consider the configuration shown in Fig. 1(a), which we refer to as the conventional imaging mode. We assume a rotationally symmetric superresolution f ilter operating in the realms of scalar diffraction theory. 1997 Optical Society of America The normalized field C measured at the observation plane is given by2 Cshd 2 Z 1 0 AsrdexpfiFsrdgJ0shrdrdr , (1) where r is the normalized coordinate at the exit pupil. The filter is characterized by an amplitude transmittance Asrd and a phase function Fsrd. Superresolution techniques differ in the way that the functions A and F are defined and optimized" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002824_ias.2007.4347776-Figure11-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002824_ias.2007.4347776-Figure11-1.png", + "caption": "Fig. 11. The manufactured stator and IPM rotor.", + "texts": [ + " The more depth of dead zone, B, is increased, the more cogging torque is decreased. 0 5 10 15 20 -4 -3 -2 -1 0 1 2 3 4 initial model \u03b4 1 = 4.10 \u03b4 1 = 5.50 \u03b4 1 = 7.00 \u03b4 1 = 8.50 \u03b4 1 = 10.00 To rq ue (k gC m ) Mechnical angle (deg.) (a) cogging torque according 1\u03b4 ( 2\u03b4 sets to 1.85 ) IV. VERIFICATION OF SHAPE DESIGN BY COMPARING WITH EXPERIMENTS From the rotor shape design, the optimized rotor parameters are shown in Table . The manufactured stator of each of initial IPM rotor and shape optimized rotor for experiment are shown in Fig. 11. The property of the proposed shape design method is verified with experiment. Fig. 12 and 13 show the back electromotive force (EMF) of an initial IPM motor and a shape optimized IPM motor at 1000 (rpm). The back EMF by using FE analysis is nearly preserved in comparison with the back EMF of the experimental results. The frequency analysis of back EMF is shown in Fig. 14. Besides the reducing of the fundamental component, the fifth harmonics component of the shape optimized IPM motor is large reduced" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000429_tmag.2013.2276092-Figure9-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000429_tmag.2013.2276092-Figure9-1.png", + "caption": "Fig. 9. Temperatures for some typical positions within the cross-section of the motor.", + "texts": [], + "surrounding_texts": [ + "In this study, the lumped parameter thermal model is carried out using the Motor-CAD software package to assess the rated operation condition and thermal behavior of the motor [16]. To reduce the temperature rise, the prototypes of the motor use a closed, forced fluid cooling system using constant temperature emulsified water. The water flows at the rated value of 0.6 L/m inside the chamber that spiral surrounds the stator. The temperature of inlet water is set at 30 C. Based on the losses obtained in the previous sections, temperatures for some typical positions within the motor, operating at the rated operation condition, are plotted in Figs. 9 and 10. It can be seen that the normal operating temperature of motor is below critical temperature." + ] + }, + { + "image_filename": "designv6_24_0001388_s1644-9665(12)60163-0-Figure13-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001388_s1644-9665(12)60163-0-Figure13-1.png", + "caption": "Fig. 13. A conventionally produced bumper stay (a), hydroformed bumper stay (b), bumper stay hydroforming process (c) with detailed shape of tube wall (d) [46]", + "texts": [ + " There exists a considerable interest to reduce vehicle weight through the adoption of lightweight materials, such as aluminium alloys, while maintaining energy absorption and component integrity under crash conditions. The interaction between tube hydroforming and behaviour during crash events was studied using lightweight automotive structural members [43]. There was used a high-pressure hydroforming process in which tubes with various corner radii in the tube cross-section were produced, Figure 12. Next the tubes were subjected to axial upsetting. The tubular hydroforming process can be used to produce a bumper stay, which secures the bumper beam to the vehicle body (see Figure 13b). When the automobile is hit from the front or behind, the bumper beam collapses and the impact force is transmitted to the left and right front frames, respectively, through the bumper beam and bumper stays. The impact energy is absorbed by plastic deformation of the bumper beam and bumper stays. The conventional bumper structure is assembled from several parts (Figure 13a), so several manufacturing processing steps are needed, and the structure is somewhat complex. Most research work on bumper stays has focused on using reinforcing members that have complicated shapes [44\u201345]. Hydroformed bumper stay (Figure 13c,d) is rather simple in the shape but its ability to absorb energy through plastic deformation is relatively high. In most of the tube hydroforming processes, the decrease in wall thickness is prevented by compressing the tube in the axial direction simultaneously with the action of the internal pressure (see example in Figure 1). If the internal pressure is too small, the axial compression causes the wrinkling of the tube wall. Hence the paths of internal pressure and axial compression in the tube hydroforming are keys to prevent the occurrence of these defects" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000234_icelmach.2016.7732762-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000234_icelmach.2016.7732762-Figure2-1.png", + "caption": "Fig. 2. mcA: for 2, 4 and 6 poles according to the PM magnetization direction", + "texts": [ + " An analytical model combined with FE simulations has been considered to compute the minimum synchronization voltage of LSSMs [11]. The rotor parameters of the LSSM are computed by means of FE as a function of magnetizing currents and rotor frequency. Then, the parameters achieved with FE are used in the analytical dynamic analysis. The rotor cage of all the LSSMs shown in this paper is designed to synchronize a load inertia at least 5 times the motor one at full load at rated voltage for all the considered number of poles. A. mcA: for 2, 4 and 6 poles Fig. 2 illustrates the lamination of the mcA, whose geometrical data are shown in Table I. It is a lamination suitable to be used in 2-, 4- and 6-pole LSSM, respectively, by changing the magnetization direction of the PMs in the rotor and the stator winding. The slots occupied by a phase of the stator winding are highlighted in Fig. 2 for the different number of poles. The stator lamination is a standard 4-pole IM one. The rotor exhibits symmetry over 60 mechanical degrees and there are 30 rotor slots. The considered PMs are high energy NdFeB. They are oriented so as to form 1, 2 and 3 pole pairs, respectively. The machines with different number of poles have been analyzed by means of FE analysis. Fig. 3 illustrates the no load flux lines and the flux density. In the 4-pole machine, 3 barriers over 6 contain PM with different magnetization direction. In order to avoid PM flux to be short-circuited between any other pole, the rotor bar at the center of the PM is deeper with respect the other bars, as in Fig. 2. As illustrated in Fig. 3(a), the 2-pole machine exhibits a high flux density in the back iron (1.8T ) and the associated iron losses will be relatively high. As reference, the flux density in the tooth and in the back iron at no load in the corresponding IM is 1.6T and 1.5T , respectively. The performance of mcA are shown in Table II as a function of the number of poles. The highest power at which the IE4 efficiency is reached is associated to the rated power. The losses components (Joule losses Pj and iron losses Piron) varying with the number of pole of the machine are highlighted" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000307_el.2010.1791-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000307_el.2010.1791-Figure2-1.png", + "caption": "Fig. 2 Photographs of fabricated balun a Top b Bottom", + "texts": [ + " Therefore, when the signals are transmitted to the I1-I1\u2032 and I2\u2032-I2 section, the phase shift between the signals at I1-I1\u2032 and I2\u2032-I2 section is 1808 or -1808. Owing to the inherent wideband characteristic of the microstrip to CPW and microstrip to CPS transitions, a wideband 1808 phase shift with desirable performance can be implemented. Simulated and measured results: The proposed balun is modelled by using 3D EM simulation software Ansoft HFSS. It is fabricated on the Rogers RT/Duroid 5880 dielectric substrate with a relative dielectric constant of 2.2 and a thickness of 0.254 mm. The photographs of the fabricated balun are shown in Fig. 2. The balun is measured by a network analyser Agilent E8363B. Its simulated and measured results are compared in Fig. 3. As can be seen from Fig. 3, good performance is achieved. Its simulated return loss, isolation, amplitude imbalance, and phase imbalance are better than 16.4, 10.4, 0.8 dB, and 58, respectively, from 19.5 to 40 GHz. Its measured return loss, isolation, amplitude imbalance, and phase imbalance are better than 13, 10.8, 1.1 dB, and 88, respectively, over the frequency range from 16" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001242_jcmsi.12.116-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001242_jcmsi.12.116-Figure3-1.png", + "caption": "Fig. 3 Power receiving antenna of automatic electrical field mapping system.", + "texts": [ + " This oscillator is connected to a desktop PC at which the frequency changes according to the serial command sent from the PC. The 2DC sheet is fixed on a table, and the coaxial cable transmits the generated microwave to the 2DC sheet. The output power of the microwave is 9 W. The size of the 2DC sheet we use is 300 mm \u00d7 300 mm. A power receiving antenna, consisting of a rectifier circuit, a current sensor (ACS712 (Low Current)), and a microcontroller (Arduino Uno R3), senses a current and transmits the sensor values to the PC after being stabilized through a low pass filter in the microcontroller (Fig. 3). We created a 47 \u00d7 47 mm antenna that has four electrodes with a rotationally symmetrical arrangement (Fig. 4) to receive stable power from the 2-DC sheet (Fig. 5). Both the microcontroller and the oscillator work synchronously. After switching the frequency of the oscillator by 0.01 GHz, the system saves the sensor value to the memory of microcontroller at some time intervals. After repeating up to 2.50 GHz, the microcontroller sends the sensor value to the PC. The current sensor has built-in semi-fixed resistance and can adjust the value of the sensor" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002378_tasc.2010.2092745-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002378_tasc.2010.2092745-Figure6-1.png", + "caption": "Fig. 6. Simulation results of PFM: (a) ; (b) ; (c) ; and (d) (Hx and Hy are magnetic fields on x axis and y axis respectively.).", + "texts": [ + " The pulse current is discharged from the capacitor bank (15000 /2000 V) which is pre-charged by a DC power supply. The IGBT as a power switch is to make PFM start or stop by turning on or off. To avoid resonance of capacitor and coil inductor, 6 diodes array parallel to capacitor. Fig. 5 shows the photograph of PFM system. The experiment parameters are list in Table I. A simulation of PFM of rotor is also processed in COMSOL. The superconducting model in [3] is used. 4 magnetizing coils which are successively applied with a pulsed current each pair are used in simulation. Fig. 6 shows simulation results. It only applies one pulse in simulation. From streamlines distribution as shown in Fig. 6, it indicates trapped magnetic field on 4 poles by superconducting bulks distributes asymmetrically. This is because To investigate the distribution of rotor trapped field, it had totally recorded 36 positions by rotating rotor in 10 degrees for each. The rotation angle is as shown in Fig. 3. Fig. 7 shows the applied field and trapped field in original (0 degree) position. The peak of applied field (measured without YBCO bulks) is 850 mT and pulse width is 0.75 s approximately. The trapped field shown in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002782_mop.24669-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002782_mop.24669-Figure3-1.png", + "caption": "Figure 3 PIFA dimensions (in mm)", + "texts": [ + " Attempts to reduce the ground plane size have been done, and the working bandwidth is 800 MHz. Ground plane reduction modified the expected radiation pattern (monopole radiation type). Maximum gain is now found in the direction collinear to the feeding line and is around 2 dBi. As this antenna is placed along body, near field energy is radiated into it. This makes the antenna sensitive to its environment. The dimensions of the antenna are as follows: s 2 mm, l 10.5 mm, and d 3.5 mm. The PIFA (Fig. 3) is an improved thin antenna with coplanar feeding to facilitate its use in BAN context [4]. The height of the antenna is /40. The size of the ground plane has been reduced to the smallest possible dimensions. This causes the energy to radiate normally to the ground plane instead of along it, as for a monopole. The working bandwidth is around 100 MHz. The antenna gain is around 3.5 dBi. Because of its broadside radiation pattern, PIFAs might not be the most suitable antenna in case of a direct link on the body surface, chest to chest or waist to chest communication for instance" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000431_j.engappai.2004.09.003-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000431_j.engappai.2004.09.003-Figure4-1.png", + "caption": "Fig. 4. Initial stator/rotor lamination.", + "texts": [ + " Using the annealing strategy\u2014a linearly decreasing mutation probability rate with each new generation\u2014the effects of a too high or too low mutation rate can be overcome. The operator is useful if the value of the probability is 0.1%, while in the annealing strategy it starts with 1% and ends at 0.1%. We evaluated the proposed evolutionary design approach by estimating the actual improvement in the efficiency of an initial UM that was designed using the conventional direct design approach: 1. With the ANSYS software, we calculated the efficiency of an initial UM. An outline of the rotor/ stator lamination of this motor is shown in Fig. 4. The power losses of this motor were calculated to be 313W and the output power was calculated to be 731W (Table 2). In the outline, the levels of magnetic flux density through the rotor/stator lamination are shown, expressed as Tesla (T). The darkest gray color indicates areas with the highest level of magnetic flux density, which results in high iron losses. The copper losses are not shown in this area. 2. After several runs of the DOptiMeL software, a set of promising solution candidates was collected", + " It took around 3000 runs for the optimization to converge, and since one finiteelement analysis took about 7min on a Pentiumbased computer, the whole optimization lasted for 2 weeks. Most of the solutions that were given by the DOptiMeL program show a significant reduction of the iron and the copper losses, in comparison with the losses in the initial motor. The best solution results in a power-loss reduction of 24%, and gives us a motor with iron and copper losses of 239W (see Fig. 5). The main differences between the initial design (Fig. 4) and the optimized design (Fig. 5) are: (a) the height of the rotor-and-stator laminations is increased by 13%, (b) the rotor radius is increased by 5%, (c) the slot (copper) areas in the stator and the rotor are larger, and (d) the iron area in the rotor is larger. A comparison of the magnetic flux densities in the initial and the optimized motor shows a clear reduction of the areas with the highest levels of magnetic flux density in the optimized motor. In the optimized lamination, the copper losses in the rotor and the stator are significantly lower than in the initial lamination" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001602_icelmach.2012.6350262-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001602_icelmach.2012.6350262-Figure6-1.png", + "caption": "Fig. 6. Flux density distribution (a) Constant linear \u03bcr=3000 (b) Non linear \u03bcr as function of flux density", + "texts": [ + " The analytical model compares well with these two linear simulations shown by the deviation of the results compared to the FE models. B. Non-linearity effect In this simulation, analytical model compare to finite Element simulation with non-linear material in the stator. BH curve of the stator material show on Fig. 5. The change of flux density in stator while the material changed from linear material with constant permeability \u03bcr=3000, to material with permeability as function of flux density show in Fig. 5. The resultant flux density distribution is shown in Fig. 6(b). The maximum flux density is 2.19 T. This value decreased from 5.9 T in the linear stator model shown in Fig. 6(a). Table III shows the loss comparison between three conditions in machine II while running at rated speed and rated current. The deviation between analytical and FEM with non linear material is around 50 %. C. Slot opening effect The slot opening has big influence on no load-losses. These losses are present due to flux pulsations in the magnetic field due to slotting of stator. The no load losses are influenced by the change of flux density due to the change of reluctance. This change of reluctance is caused by the geometry of slot-pole combination" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002755_tvlsi.2012.2227848-Figure16-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002755_tvlsi.2012.2227848-Figure16-1.png", + "caption": "Fig. 16. SOI FinFET 6T SRAM (113) configuration. (a) (FEOL+BEOL) and (b) FEOL only. Dielectric regions are not shown.", + "texts": [], + "surrounding_texts": [ + "Owing to the width quantization property, multigate FETs with large electrical widths need to have multiple fins. We synthesized multifin FinFETs using the bulk and SOI FinFETs generated earlier at the 22-nm/14-nm/10-nm nodes. They consisted of four fins each, with shared raised source/drain epiregions that are via-contacted and connected using metal-1, as shown in Fig. 8(a) and (b). We varied the fin pitch, FP, which is the distance between the centers of consecutive fins, and computed the parasitic (FEOL+BEOL) capacitances for each layout using the setup described in Fig. 7(b). From Fig. 9(a), we can see that the trends in CDRAIN,TOT are in stark contrast to the single-fin results in Section III. While moving from SOI to bulk FETs, there is a 11.5%, 10.8%, and 8.8% increase in CDRAIN,TOT for the 22-, 14-, and 10-nm nodes, respectively, which can be attributed to the shared drain-to-bulk fin capacitances in bulk FETs. However, in the case of CGATE,TOT [Fig. 9(b)], there is only a 2%\u20134% increase from SOI to bulk FETs. An increase in FP from 40 to 70-nm results in a 20%, 31%, and 36% increase in CDRAIN,TOT for the 22-, 14-, and 10-nm nodes, respectively, while CGATE,TOT increases by 16%, 26%, and 28%, respectively. These results suggest that gate-toepi-source/drain/metal-1 capacitances begin to dominate as FP increases or the technology node decreases, and they highlight the need to model the entire (FEOL+BEOL) structure." + ] + }, + { + "image_filename": "designv6_24_0002153_amm.813-814.915-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002153_amm.813-814.915-Figure5-1.png", + "caption": "Figure 5: Mode shapes of curved 3 spokes design (Steel)", + "texts": [], + "surrounding_texts": [ + "Static Structural Analysis. A pressure of 6 MPa is applied on the face of rim and all DOF are fixed for the hub inner face as shown in Fig 2b. The response of the wheels in terms of Total deformation, von mises stress, structural stiffness and specific structural stiffness are calculated and compared the results for all 6 design with three materials discussed in previous section. Tabular Values Spoke type/Result type Inclined spokes Curved spokes Y-shape Spokes Deformation [mm] 0.4257 0.2956 0.2327 Von mises stress [Mpa] 574.76 384.85 791.4 Stiffness [N/M] 813.26 1171.3 1487.9 Mass[Kg] 9.3295 9.8311 10.831 Specific structural stiffness[N/M-kg] 87.171 119.14 137.38 stiffness of 3 spoke curved design made up of steel is better than others. \u2022 All 5 spokes designs produced the higher von mises stress than the 3 spokes designs. It can be observed that the von mises stress does not vary much based on the type of the material chosen for manufacturing. All three spoke designs lower von mises stress than the other designs. Modal Analysis Results. It is done to find out the natural frequencies of the structure. All degrees of freedom are fixed at the hub as shown in fig 2a and performed the modal analysis for all designs. The typical contours are shown in below section. Contours Tabular Values Type/Mode Number Inclined spokes Curved spokes Y- shape 3 5 3 5 3 5 1 126.7 178.9 134.6 196.5 118.2 166.1 2 132.7 179.2 141.9 195.6 125.2 166.4 3 168.5 245.9 172.2 257.1 173.0 269.9 4 222.9 248.5 230.7 262.8 207.5 283.9 5 225.0 249.3 234.7 267.9 270.1 327.0 6 233.1 342.2 276.9 380.5 270.4 328.2 Mass (Kg) 1.914 2.139 1.985 2.258 2.123 2.483 Observations. All 5-spoke designs have higher natural frequency than the 3-spoke designs. For three shapes of the spokes, curved spoke designs produced the higher natural frequencies. Designs produced the nearly same natural frequencies irrespective of the material. Even though aluminum and magnesium have less mass than the steel, they produced the nearly same frequencies." + ] + }, + { + "image_filename": "designv6_24_0001347_ijabe.v10i6.3142-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001347_ijabe.v10i6.3142-Figure2-1.png", + "caption": "Figure 2 Configuration of pneumatic rice seed metering device", + "texts": [ + " The speed of the seed metering device and the advance speed of the transplanter are shown in Equation (1): 6000 v n Kl (1) where, n is speed of the seed metering device, r/min; v is advance speed of the transplanter, m/s; K is the number of group holes on the sucking plate; l is hill spacing, cm. According to Equation (1), with the same hill spacing, the faster the advance speed of the seeder is, the higher the rotation speed of the seed metering device. With the same advance speed of the seeder, the bigger the hill spacing, the lower the rotation speed of the seed metering device. The core component of pneumatic rice direct-seeder is pneumatic rice seed metering device. The seed metering device is shown in Figure 2. The precision rice seed metering device mainly consists of seed box, sucking chamber shell, stratified room, seed chamber shell, cleaning device, sucking plate, shaft and seed tube. When the seed metering device is operated, the seeds are getting into the seed filling room of the seed chamber shell through seed box and stratified room. The seed filling room is filled with seeds and then seeds accumulate on the surface of sucking plate. Seeds are sucked on the holes of sucking plate under the effect of vacuum of sucking chamber shell" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000957_1.1559581-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000957_1.1559581-Figure5-1.png", + "caption": "Fig. 5 Correction functions for point indication", + "texts": [ + " The correction function for the surface modification by a specified point is the product of the correction functions for parameter directions u and v . The correction function gm p (u ,v) for the m-th specified point is gm p ~u ,v !5 f m ,u~um ,u ! f m ,v~vm ,v !, (4) where f m ,u(um ,u) is the correction function for the u direction, f m ,v(vm ,v) is that for the v direction, and (um ,vm) is a parameter value at the specified point. We can specify the region where a surface is to be modified by the range of parameter values. Examples of correction functions for the point indication are shown in Fig. 5. A correction function is generated as a surface by multiplying the correction functions according to Eq. ~4!. Figure 5 shows the case for Type 1, where the indicated point is located in the central region. The left side of the figure shows the correction function for the modification over the entire area ~@0,1#,@0,1#! and the boundary condition is C0. The right side of the figure shows the function over the specified region (0.2 d, rotation angle 6 < 7~12. In such case, by utilizing the geometrical relationship between the beams in Fig.S(a), we can obtain the following equation. S = dsin8+ Scos8, (1) where d is a half of the gap between the parallel beams and 8 is the rotation angle as shown in Fig.3. When the contact length 0 < S 5 d, the equation obtained from Fig.3(b) is S = dsin(7r - 8) - S c o s ( ~ - 8) (2) where 6 2 ~ / 2 . From eq.(l) and (2), we can easily compute the contact distance S as follows. (3) d sin 13 s=- Fig.4 shows an example of sensing curve (contact length vs rotation angle) computed from eq.(3). It should be noted that the contact length is a function of the rotational angle alone, while the sensitivity depends on the contact position. 1 -case 3 Interaction between the beams After the first beam makes contact with an object, the beam is deformed by the pushing force" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001668_1460412.1460439-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001668_1460412.1460439-Figure5-1.png", + "caption": "Figure 5: The Epic family includes hardware specifically designed for (a) making platform prototyping possible in a classroom setting by novice designers (b) interfacing with the popular Phidgets analog and digital sensors (c) empowering module designers to construct, probe, and debug intricate circuits on-the-fly, both only using (d) off-the-shelf parts such as jumpers, sensors, solar power packs, and surfboards.", + "texts": [ + " Unfortunately, several factors increase time-to-result. Issues like sensor and power supply selection, electrical wiring, and device driver development dominate engineering efforts while more novel aspects like application software, performance characterization, and end-user data collection are routinely back burnered during the initial stages. To improve productivity, we created a Development Board that can be easily and inexpensively integrated with off-the-shelf sensors, displays, and solar packs to improve time-to-result. Figure 5(a) shows the Development Board, which benefits from the choice of an industry-standard LCC-68 footprint by including an off-the-shelf socket for easily swapping modules. Adhering to the principle that all signals should be available to the platform designer, breakout pins allow access to every signal, simple shorting shunts allow each signal to be individually connected to power or ground, and jumper wires allow a signal to be easily connected to off-the-shelf parts like the ones shown in Figure 5(d). The Development Board also incorporates a USB module for programming, alkaline and lithium battery connections for supplying power, and LEDs and buttons for feedback, debugging, and control. This flexible platform enables quick prototyping and exploration of novel development elements while circumventing the complexities of module and carrier design. The board has already been used by undergraduate students to develop application-specific platforms and a second version, shown in Figure 5(b), was used to teach a summer school on wireless embedded systems. Debugging is an often frustrating aspect of prototyping. Effective debugging requires the developer to probe signal voltages to verify circuit operation and measure currents to identify unexpected draws and verify expected ones. Unfortunately, many systems can make probing signals and debugging painfully difficult: signals are buried under chips, routed through to intermediate layers of the printed circuit board, and never exposed through any header", + " In most systems, directly measuring the individual draws of the microcontroller, radio, flash, or other peripherals is impossible since the individual power supply lines are buried in the circuit board and a single, global power supply line is exposed. The result is that developers must write test code that isolates different functions, rather than being able to directly observe the system running application code. To address these challenges of hardware debugging, we developed a breakout board, shown in Figure 5(c), that includes an LCC-68 socket, pins for easily accessing and jumpering each signal, and an Epic programming port. With access to the full array of signals, hardware developers can easily probe every point in a design, connecting the circuit, multimeters, oscilloscopes, and other monitoring equipment as they see fit. Carriers are circuit boards that act as substrates to glue together general-purpose modules with application-specific sensors, power supplies, and mechanical constraints. To evaluate the utility of our proposed architecture, we designed and implemented several different pilot-stage carrier boards" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000489_tgrs.2021.3051727-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000489_tgrs.2021.3051727-Figure3-1.png", + "caption": "Fig. 3. Geometric diagram of GMTI for the spaceborne BiSAR system, LT-1 [11], [13].", + "texts": [ + " Moreover, the required baseline Bcross for the XTI of the spaceborne M-SAR can be described by (3). It is obvious that the cross-track interferometric baseline increases with the improvement of XTI accuracy Bcross = \u03bb \u00b7 R \u00b7 sin \u03b8 2 \u00b7 hamb . (3) ATI SAR is based on the time difference between two moving SAR instruments separated along the along-track direction to image the same terrain, so as to measure the speed of targets on the ground [18]. The geometric diagram of ground moving target indication (GMTI) performed by the spaceborne BiSAR system, LT-1 [11], [13], is shown in Fig. 3. ATI SAR adopts two antennas, and their azimuth separation distance (also known as along-track baseline) is L. For SAR system operating at an altitude of several hundred kilometers, the time lag between the two antennas that observe the same region is t = L/vsat, where vsat is the satellite speed. Besides, it can be assumed that the distance difference between the two antennas for observing the targets with a velocity component of vr along the radar LOS is \u03c1, which is expressed as \u03c1 = vr \u00b7 t . Therefore, the phase difference of the moving targets in two images can be described by (4) [18] \u03d5 = 4\u03c0 \u03bb \u00b7 L vsat \u00b7 vr " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002367_iros.2009.5354127-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002367_iros.2009.5354127-Figure2-1.png", + "caption": "Fig. 2 Internal structure of the motor bearing", + "texts": [], + "surrounding_texts": [ + "A. Rotating ball The rotating ball is arranged so that its center coincides with the origin of the orthogonal X0, Y0 and Z0 axes. Also, the configuration is such that there are three mutually orthogonal bore holes, and the drive shafts which have the hollow shaft motors applied to them are inserted into these bore holes and fixed with pins, etc., so that the motors are transmitted via the bore holes, and combine to produce a rotating ball. The output shaft to the link is fixed to the rotational shaft of the Zm axis motor. The movement of the rotating ball is output to the link via the rotational shaft of the Zm axis motor, passing through its hollow bore. B. The ball\u2019s bearing and clamp The bearing of the ball is constituted by the hollow sphere formed by the rotating ball and its concentric sphere. It functions as a spherically-sliding axis-bearing supporting the rotational motion of the rotating ball in an arbitrary direction. According frames of latitude and longitude in terms corresponding to the Earth, the ball\u2019s bearing is such that around the Xm, Ym and Zm output shafts a reduced quadrilateral aperture is formed. This bearing is supported by means of the clamps. In addition, the clamps fix the bearing, and function to support the motor bearings which support each motor by means of the attachment extremities located on the 3 orthogonal axes stated above. C. Motor bearings We now focus on and explain the motor bearing supporting the Xm axis motor shown in 1. This motor bearing is formed by the great circle of the rotating ball and the concentrically curved hollow cylinder component. Also, for the surface on the side of the Xm axis motor, and the surface on the side of the rotating ball, a long hole is formed in a longitudinal direction with the insertion width of the hollow motor output axis. Also, a component known as a slider is incorporated into the hollow part of the motor bearing (see Fig.3). There are a number of bearings located in the slider. The dimensions of these bearings are set so that there cannot be any gap between the inner surface of the hollow part of the motor bearing, the inner surface on the motor side, and the inner surface of the rotating ball. The slider ensures that the internal components do not jiggle in any direction. In this way the slider, by means of the rolling motion between each inner surface of the internal components yielded by the bearings, makes it is possible for the inner part of the motor bearing to slide with low friction in a longitudinal direction while keeping the rotating ball at a constant distance. The anterior end of the Xm axis motor is fixed via a support plate coupled with the slider. In this case, the output axis extending from the anterior end of the Xm axis motor passes through the slider and is joined to the rotating ball. According to the structure stated above, there is no jiggling in the Xm axis motor between the slider and the hollow part of the motor bearing, so the rotation is regulated around the axial direction of this part itself (see Fig. 3). Moreover, due to the slider there can be no reciprocation along the motor bearing, so among the other 2 orthogonal axes the Xm axis motor can turn around the Y0 axis (see Fig.4). With this type of motor bearing, the ends are fixed to the rotation axes, and there is axial support in such a way that rotation is possible at the support-end located on the Z axis of the clamp. There are two bearings installed internally in the support-end, and by aligning the two bearings in the axial direction of the rotation axis, even when an external force is applied to tilt the rotation axis the two bearings act together to support the rotation axis and prevent it from tilting. According to the structure stated above, among the other 2 orthogonal axes, the motor bearing, i.e., the Xm axis motor, can rotate about the Z0 axis (see Fig.5). We focused on and explained the motor bearing supporting the Xm axis motor, but the explanation regarding the motor bearings supporting the Ym axis and Zm axis motors is the same. According to the structure stated thus far, the device has the functionality of a 3-DOF active rotational joint between a pair of links. According to these structural principles, if the Xm, Ym, and Zm axis motors are rotated simultaneously and independently, rotational motion in an arbitrary direction is possible as a composition of motion in the direction of each DOF. Also, having a structure similar to a spherical joint, we can present a 3-DOF active rotational joint with a simple structure." + ] + }, + { + "image_filename": "designv6_24_0000008_icelmach.2018.8506886-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000008_icelmach.2018.8506886-Figure3-1.png", + "caption": "Fig. 3. Design of the prototype motor: 1 - stator, 2 - rotor, 3 - original hub with car brake system", + "texts": [], + "surrounding_texts": [ + "In order to assess the efficiency of the cooling system, an analysis based on finite element method (FEM) and computer analysis of fluid dynamics (CFD) was performed which is slower than other methods, such as thermal diagrams, and requires high computing power, but its big advantage is that allows analyzing devices of any geometry using any cooling systems. The only limitation of the method are computational capabilities of computer hardware [1],[3],[5],[6]-[9]. In order to conduct a thermal analysis, based on the finite element method, a simplified three-dimensional model of the stator of the engine was developed (Fig.4). The model has been prepared in such a way as to simplify the geometry that does not affect the efficiency of the cooling system and the thermal state of the machine. The applied model includes: an aluminium support element with a water jacket (1), a simplified stator core (2), a simplified winding model (3), a thermally conductive resin filling the space between the winding and the supporting structure (4). In the CFD analysis program, the model (shown in Fig.5) was additionally supplemented with a cooling medium in the water jacket channels. The thermal resistance substitute parameters have also been assumed: Rs - thermal resistance corresponding to the pressure between the core and the water jacket construction, R\u017c - the thermal resistance corresponding to the groove insulation. Then the model was discretized. The discreet model is presented in Fig.6. All models and calculations were performed in Autodesk Inventor and Autodesk Simulation CFD software." + ] + }, + { + "image_filename": "designv6_24_0002850_12.599519-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002850_12.599519-Figure2-1.png", + "caption": "Figure 2. One side clamped elastic beam", + "texts": [ + " The PI-Observer can be applied to estimate and locate an unknown force acting on a structure. The method can be used for model-based fault detection and control and can be applied for arbitrary structures, also for 3D structures, if the model of the structure is known (cf. Eq. (1)). Here, a one side clamped elastic beam and an all side clamped elastic plate are used to illustrate the performance of the observer. The introduced method can also be applied to arbitrary 3D structures, if the dynamical model is known. The one side clamped elastic beam is shown in Fig. 2(a). The scheme of the test rig is presented in Fig. 2(b). An elastic beam clamped on one side gets in contact with the contact surface after a short excursion of the beam (dashed position in Fig. 2(b)). The impact contact force is measured by quartz force sensors, which are mounted between the contact surface and the ground. The displacement of the beam is measured at two points with non-contacting optical measurement systems. There are also 3 strain gages mounted on the beam to measure the strain. This experiment is used to validate and test the PI-Observer for impact forces acting on an elastic mechanical structure. The contact force is estimated using the displacement measurements and (or) the strains", + "18 Using \u22022w(x, y, t) \u2202x2 = \u221e\u2211 i=1 \u221e\u2211 j=1 \u22022wij(x, y) \u2202x2 qij(t) , (32) \u22022w(x, y, t) \u2202y2 = \u221e\u2211 i=1 \u221e\u2211 j=1 \u22022wij(x, y) \u2202y2 qij(t) (33) the displacements can be calculated by the strain gage measurements. In this experiment, four strain gage measurements are used (see Fig. 3), so only w11(x, y), w21(x, y), w12(x, y) and w22(x, y) can be considered. For the beam, Eq. (31) has only to be considered for the x direction. The beam gets in contact with the contact surface after a short excursion of the beam (dashed position in Fig. 2(b)). The displacement is measured in node 4 and 5. In Fig. 5(a) and 5(b) the measured and the estimated force are compared. In Fig. 5(a) the design parameter q = 103 is too small, the observer can not follow the fast dynamic of the contact force. In Fig.5(b) the design parameter q = 108 is too high, the error caused by the measurement noise and the model uncertainties prevails. Choosing q = 106 the PI-Observers can estimate the contact force very good as can be seen from Fig. 6(a). The contacts are very fast, up to 5 ms for a contact" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003049_mop.22983-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003049_mop.22983-Figure3-1.png", + "caption": "Figure 3 Dual-band filter based on CSRRs Figure 4 Frequency response (a) before adjusted and (b) after adjusted", + "texts": [ + " Since two different resonant frequencies are corresponding to the CSRRs, so taking the CSRRs as resonant unit, dual-band filter can be designed based on the design theory of the coupled resonator filter. Furthermore, the coupling effect between the inner and outer ring in a unit or different units leads to cross coupling, therefore out-of-band transmission zeros are produced. Most of all, different resonant frequencies can be produced by adjusting the size of CSRRs reasonably, as a result, dual-band filters with different frequency ranges can be designed. The structure of the dual-band filter based on the CSRRs unit is shown in Figure 3. The dimensions of the CSRRs used in the filter is a1 9 mm, t1 1 mm, g1 0.5 mm, a2 6 mm, t2 0.5 mm, g2 0.5 mm, d 4.5 mm, e 2 mm. And the two resonant frequencies are 2.55 and 4.05 GHz, respectively. Two symmetric T-shaped feeding lines are used as the input/ output feeding lines of the filter. The size and the position of the T-shaped feeding line determine the coupling between the CSRRs Figure 2 (a) Comparison of the resonant property between CSRRs and the two single rings (b) resonant property of CSRRs with different a2 8 MICROWAVE AND OPTICAL TECHNOLOGY LETTERS / Vol", + " Figure 2 shows simulated S-parameters of proposed DGS unit in dependence on length l. The substrate used for simulation was DiClad870 0.508 mm thick with permitivity r 2.33. S-parameters were simulated using Ansoft Designer planar EM simulator. Parameters w1 1.5 mm, g 0.3 mm, and w2 0.9 mm are constant while length l changes in steps 5 mm, 7 mm, 9 mm, and 11 mm. As the length l is increased, the effective inductance increases as well. Because of this, shift of the cut-off frequency and attenuation pole can be observed. Figure 3 illustrates equivalent network structure for low-pass filter with DGS units. It consists of series branches of parallel resonant 10 MICROWAVE AND OPTICAL TECHNOLOGY LETTERS / Vol. 50, No. 1, January 2008 DOI 10.1002/mop" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002870_1.5122106-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002870_1.5122106-Figure2-1.png", + "caption": "FIGURE 2. Fields of velocities in the cross section of a ball valve with and 1.5 MPa for turned ball scheme", + "texts": [ + " 1, (a)) the total number of elements was about 600 thousand; for scheme No. 2 (Fig. 1, (b)), the total number of elements was about 650 thousand. A numerical study of the working processes in a ball valve was carried out for a full-bore valve used in the chemical industry with a high gas flow: 0.2 m, \u0307 5 kg/s with 1.5 and 2.5 MPa in the angular range 10\u00b0\u202675\u00b0. During computational research of the working process in a ball valve, the gas pressure and velocity fields were obtained for two computational schemes (Fig. 2 and 3). Let us consider the results of numerical simulation in more detail. As a result of transition from scheme No. 1 to scheme No. 2, the pressure losses in the ball valve increases 1.3 ... 2.9 times, smaller values correspond to greater 030056-4 values of . However, the absolute values of flow resistance for small angles differ slightly. Higher gas velocities are realized for scheme No. 2 in a ball valve (Fig. 3), which leads to higher pressure losses compared with scheme No. 1. So, for 45\u00b0 and 1" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002295_b978-0-08-011216-9.50010-9-Figure7.2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002295_b978-0-08-011216-9.50010-9-Figure7.2-1.png", + "caption": "FIG. 7.2", + "texts": [ + "6), putting ay = 0, will be which may be expressed in terms of the bending moment and torque as \u03c3\u03b9,\u03c32 = ^\u039c\u00b1{\u039c2 + \u03a4 \u03c8 ^ (7.3) Since the third principal stress is zero and \u03c32 is negative, the maximum shear stress is The above expression may be used when considering the criterion of yielding of ductile shafts, and if \u03c3 \u03b3 is the yield stress in simple tension, then according to the Tresca theory at yield 2 nd*K + } or M* + T>= { ^ ^ y ) 2 . (7-4) Alternatively if the von Mises criterion, section 3.15.2, is used, it can be shown that M2 + \u00ccT2={\u00cciaY}2\u00b7 ( 7 \u00b7 5 ) EXAMPLE 7.1 The hollow circular shaft shown in Fig. 7.2 is subjected to the combined action of axial thrust, bending and torsion. Determine the principal stresses and maximum shear stress at point A on the surface of the shaft. For the cross section shown Area =~(122 - 42) = 100-5 in2, / = A ( 1 2 4 - 4 4 ) = 1005 in4, 64 Ip = \u00cf2 (124 \" 44) = 2 0 1 \u00b0 in4> 10,000 Axial stress = \u2014 = \u201499-5 lb/in2, 100*5 \u201e \u039b. \u039b My 10,000 x 48 x 6 Bending stress at A = \u2014\u2014 = \u2014\u2014 = -2865 lb/in2, . , , \u039b Tr 400,000 x 6 Torsional shear stress at A = \u2014 = \u2014 /\u201e 2010 = 1194 lb/in2" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002878_eej.22684-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002878_eej.22684-Figure1-1.png", + "caption": "Fig. 1. Magnetic-geared motor in this study.", + "texts": [ + " To solve these problems, we have proposed a magnetic-geared motor with permanents magnets arranged only on the high-speed rotor, and have investigated its cogging torque characteristics and torque ripple characteristics [3\u20135]. However, it was still unclear how the slip characteristics in the presence of overload affected theN\u2212T characteristics. Thus, in this paper we report the N\u2212T characteristics of magnetic-geared motors obtained by coupling magnetic field analysis with vector control, and the results of experiments using a prototype. 2.1 Structure A magnetic-geared motor with permanent magnets arranged only on the high-speed rotor is shown in Fig. 1. Eight magnet pole segments (Br = 1.3 T) are attached to the high-speed rotor. The low-speed rotor is composed of 20 pole pieces consisting of laminated electric steel plates. The 12-slot stator is provided with concentrated three-phase windings. 110 turns of copper wire (diameter 0.5 mm) are provided at each tooth, with 4Y connection (quadruple parallel star connection). The resistance is 1.0 \u03a9. 2.2 Working principle The coils are wound on the stator so as to obtain four winding pole pairs producing magnetomotive forces when a three-phase AC current flows" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003860_iembs.2009.5332578-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003860_iembs.2009.5332578-Figure6-1.png", + "caption": "Fig. 6. (a) New model design GNO version 2 (b) New GNO mounted on a wheelchair.", + "texts": [ + " Instead, he propped up his left hand to the rim of the can using his right hand and transferred the items from one hand to the other (Fig. 5a). Within the device however, he was able to complete the task with one hand even though it took him longer (Fig. 5b). Patient feedback on how to improve the GNO was recorded and discussed within multi-disciplinary team meetings. Patients recommended a more efficient mounting configuration, and a smaller, less conspicuous frame design. A second modified GNO version 2 was then designed and built (Fig. 6). Although the working mechanisms remain similar, the frame was redesigned to address their concerns. The new GNO frame no longer mounts to the back of the wheelchair, hanging over the shoulder. Instead, it replaces the wheelchair\u2019s existing arm rest. In addition, a new brushless DC motor was used instead of a stepper motor for its higher torque. The motor generates motion directly at the elbow through a worm gear transmission that has the benefit of preventing the patient from unintentionally moving the arm shelf" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001756_8.366381-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001756_8.366381-Figure1-1.png", + "caption": "Fig. 1 . D = 0.5 mm, I I = 2 mm). Microstrip patch geometry and the near field probe (I = 9 mm,", + "texts": [ + " The measured results for the near fields presented in this paper have been obtained using this scheme. For illustration, some examples of microstrip patches of the generalized annular sector type [28], [29] have been chosen. 0018-926X/95$04.00 0 1995 lEEE BOKHARI e / 01.: NEAR FIELDS OF MICROSTRIP ANTENNAS I89 11. THEORY The geometry considered is a single dielectric layer microstrip patch antenna excited by a coaxial cable connected perpendicularly to the patch through the grounded dielectric slab (Fig. 1). The usual assumptions in the Green's function solution are made. Namely, an infinite transverse dimension for the grounded dielectric slab, a perfectly conducting ground plane and both the ground plane and the patch of infinitesimal thickness have been assumed. The coaxial probe is modeled as a uniform vertical current filament of length h, the validity of which has been well established for electrically thin dielectric slabs. The unknown electric current distribution [ J ( x . y ) = 2 J, (.c,:y) + ;ij ", + " It is necessary to restrict the number of unknowns used in the representation of the current distribution to be less than or equal to BOKHARl e! ul : NEAR FIELDS OF MICROSTRIP ANTENNAS 191 (N1/2 x N z / 2 ) so that the error due to circular convolution is eliminated. Equations (12)-(15), and (19)-(20) are now in a form convenient for the application of iterative methods. The biconjugate gradient method for a symmetric operator has been used for the solution of the current distribution. The computation of the near fields is restricted to rectangular tangential components on a plane parallel to the patch surface (Fig. 1). These contain a substantial amount of the far field information and are sufficient to reconstruct it over a relatively large angular extent around the broadside (defined as the 0 = 0) direction. The measurement probe is a short cylindrical dipole with an integrated diode at its center [27]. The assembly is mechanically displaced along the .c and :y directions in the measurement plane and with the orientation of the dipole fixed either along the 2- or the y-axis. The actual problem is that of a center loaded dipole near the microstrip patch the currents on both of which must be treated as unknown", + " The total number of cells for geometries A , B, and C were 2533, 2200, and 2 155, respectively, and the number of unknowns are approximately twice this number. I t is interesting to observe that the sensitivity of geometries B and C shows up in the near fields as discussed below. Before measuring the near fields, a simple verification of assumptions (i-ii, Section 11) was made by measuring the return loss of the antennas with and without the near field BOKHARI er ul.: NEAR FIELDS OF MICROSTRIP ANTENNAS 193 probe in their close vicinity (Fig. 1, d = 2 mm). Geometries '4 and B showed no significant change in their impedance characteristics. However, geometry C showed a dramatic change in impedance values depending on the probe position and orientation. It was therefore not possible to obtain reliable estimates of the near fields of this geometry by measurement. Note that, in academic terms, the insensitivity of retum loss to the presence of the measurement probe is necessary but not sufficient to guarantee unperturbed near fields. Geometry B which exhibits a high impedance level was matched with a coaxial triple stub tuner for the near field measurements" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002779_s11044-016-9540-9-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002779_s11044-016-9540-9-Figure3-1.png", + "caption": "Fig. 3 Flexible slider\u2013crank mechanism", + "texts": [ + " For the precise and efficient derivation of the eigenmodes, however, Nelson\u2019s method [14] can be used as also proposed in [5]. In order to test the gradient computation procedure, the structural sensitivity of a flexible piston rod of a slider\u2013crank mechanism is analyzed. The sensitivity information can be used, for instance, in the topology optimization of the piston rod; see [10]. The flexible slider\u2013crank mechanism, which is used as an application example, consists of a rigid crank and a flexible piston rod; see Fig. 3. The eccentricity \u03b5 of the crank is 0.1 m and the distance l between the bearings of the piston rod is 1 m. Since no sliding-block is attached to the system, the loading on the piston rod in motion originates only from its own inertia. As shown in [9], in this case, it is crucial to provide exact gradients in order to obtain viable optimization results. The motion of the system is composed of two phases and applied via a rheonomic constraint of the crank angle \u03d5 as \u03d5(t) = \u23a7 \u23aa\u23a8 \u23aa\u23a9 i=7\u2211 i=0 ait i , 0 s \u2264 t \u2264 2 s, \u03a91t + \u03d51, 2 s < t \u2264 3 s" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001801_j.triboint.2016.11.027-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001801_j.triboint.2016.11.027-Figure1-1.png", + "caption": "Fig. 1. Photograph of prototype nested compression spring gas foil bearing.", + "texts": [ + " Springs are widely used in many types of equipment because it is not only cheap but also easy to obtain. Moreover, the stiffness of the supported structure can be easily controlled by using springs with different diameters. The feasibility of the compression spring foil bearing has been demonstrated and its load capacity is comparative with that of the bump-type GFB. Feng et al. [27] proposed an advanced compression spring GFB by nesting the underlying springs. The nested compression spring GFB is shown in Fig. 1. The advantages of this bearing design is distinctly. Firstly, a high structure stiffness can be achieved by increasing the spring numbers because the springs are nested with one another. The high spring density reduces the top foil sagging between adjacent springs, thus improving ultimate load capacity. Secondly, the damping of NSFB is improved simultaneously as the spring number increases because of the dry friction between two adjacent springs. This means the NSFB can easily achieve high structure stiffness and high damping characteristic simultaneously, which is difficult for the bump-type GFB because the contact area between bump foil and sleeve is reduced by increasing bump numbers" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001939_2015-01-0088-Figure9-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001939_2015-01-0088-Figure9-1.png", + "caption": "Figure 9. Gear Shift Fork - Contact Patch Results", + "texts": [ + " Contact analysis has been carried out to check the contact status using FE Ansys Solver. Surface to surface contact is provided between mating surfaces of forks legs & third pad with synchronizer ring. Contact starts at fork legs that will have a hinge effect before it touches third pad meeting the required deflection. A minimum nominal gap is provided for 3rd pad to share the load when max/abuse load appears on the load shifting Jaw. Gap could be adjusted if the FOS in the legs goes below 1.0 for the Max/Abuse Load as shown in the Fig. 9 and Fig. 10. Loads and Boundary conditions as shown in the Fig. 8. Once the third pad come into contact stress pattern shifts from fork legs to the middle of the fork as shown in the Fig. 11. Stress induced should be below the Yield strength of the material (FOS > 1.0). If not, the web thickness will be increased to meet the requirement. Experimental Verification As per VE Commercial Vehicles Ltd, standard durability duty cycle, rig has been setup as shown in the Fig. 12 and tested the transmission assembly in which Aluminum Gear Shift Fork's (1st & Reverse Fork and 4th & 5th Fork) are the test components" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001320_861355-Figure19-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001320_861355-Figure19-1.png", + "caption": "Figure 19 - Nutating traction drive CVT (Vadetec circa Ig82).", + "texts": [ + " Department of Energy, developed a design for a toric CVT for use in a hybrid pas senger car. Traction Propulsion, Inc., also participated in the project until 1984, when the Department of Energy terminated the contract before the design was fully implemented. A cross section of a transmission representative of that effort is shown in Figure 18 (Reference 20). Vadetec Corporation demonstrated several designs of an ingenious nutating cone-type traction drive CVT designed for automobiles and tractors [21], Figure 19. The exact status of this project is not fully known, but recent comments in the press indicate that the tractor CVT development is continuing at another company under license from Vadetec) while at Vadetec a new kinematic arrangement) more suitable for passenger cars, is being developed [22]. In the last few years) frequent reports of considerable activity on toroidal CVTs have been forthcoming from Japan. Most of the work seems to be centered at the University of Tokyo Insti tute of Industrial Science and the University of Yokohama [23]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001107_j.matdes.2005.08.009-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001107_j.matdes.2005.08.009-Figure2-1.png", + "caption": "Fig. 2. For DP590, test specimens in both the T\u2013L and the L\u2013T directions were cut from a 25.4 mm \u00b7 25.4 mm \u00b7 5.5 mm plate.", + "texts": [ + " Fracture toughness values of the DP590 steel were also obtained through the correlation with the Charpy impact energy data. In addition, fracture surface appearance, effect of chemical composition, and thickness correction are discussed. Both DP590 and AISI-1018 steels were used in this study. Chemical compositions of the two steels are listed in Table 1 and the mechanical properties are in Table 2. The properties are from [9,10], respectively, for DP590 and AISI-1018 steels. Test samples were cut from a 304.8 mm \u00b7 304.8 mm \u00b7 5.5 mm DP590 plate. Specimens in both the T\u2013L and the L\u2013T directions as shown in Fig. 2 were prepared. The in-plane specimen dimensions follow the ASTM Standard E-23, that is, 10 mm \u00b7 55 mm with a 2 mm deep, 45 V-notch having a 0.25 mm tip radius at the center of the specimen. For the AISI-1018 steel, pre-fabricated standard ASTM E-23 CVN specimens are readily available from Laboratory Device Company [11]. Half of the specimens were machined to reduce the thickness from the standard 10\u2013 5.5 mm to study the effect of the thickness. Tests were conducted using a Tinius Olson pendulum Charpy impact tester with a maximum capacity 339 J (250 ft-lb)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002574_s0266-3538(02)00044-1-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002574_s0266-3538(02)00044-1-Figure6-1.png", + "caption": "Fig. 6. Strain distributions for layer 1.", + "texts": [], + "surrounding_texts": [ + "In this paper a mechanical characterisation of pullwound carbon-epoxy spinnaker poles has been presented. The motivation of the study was provided by the exigency of defining a flexible numerical design procedure able to characterise the product performance so avoiding a number of expensive and time-consuming trials. The availability of a tailored numerical model that can allow to predict some of the key mechanical features of a family of structural shapes is, in fact, highly desirable for the manufacturer. In this context, to the authors\u2019 opinion, it is necessary for practical applications to avoid sophisticated theoretical and/or numerical modelling, mainly devoted to the description of the post-elastic phenomena whose effectiveness is often not experimentally confirmed, and to approach the problem making use of standard FE commercial codes strictly related to experimental tests. The proposed numerical model has been \u2018\u2018fed\u2019\u2019 with material data determined through a series of accurate experimental tests carried on a set of spinnaker pole prototypes. In particular in the adopted numerical FE model the composite material was treated as orthotropic and analysed in the elastic regime. The attribution of the elastic parameters to the specific laminate has been sough through a set of laboratory experiments and simple theoretical assumptions. The matrix apparent Young modulus, Em, has been derived from flexural tests carried out on single-layer pultruded flat strips, by applying the r-o-m. The longitudinal and transverse moduli for the multi-layer laminate have been considered changing from layer to layer due to the variation of the fibre volume fraction values in the layers, which, in turn, have been experimentally measured. The G values, have been determined by forcing the numerical model to reproduce the results of the experimental determination of the initial specific ring stiffness for a set of available pole prototypes. The above test appears to be a significant feature for the product performance characterisation. The ability of the model has been finally validated performing a numerical analysis on a different set of pole prototypes (with a different geometry) and by comparing the numerical results so obtained with the ones again determined via experimental tests. The proposed model is applicable to analyse analogous structural elements, i.e. elements made of the same material but having different geometrical dimensions and slightly different shapes, it certainly represents an effective design tool when different loading conditions have to be considered." + ] + }, + { + "image_filename": "designv6_24_0001313_srin.198900277-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001313_srin.198900277-Figure3-1.png", + "caption": "Figure 3. Geometry of fatigue crack initiation specimen", + "texts": [ + " Geometrie der ErmudungsriBeinleitungsprobe ployed. Triangular wave-form was used for all strain am plitudes and the frequency of cycling was adjusted for each strain amplitude to yield a constant strain rate of about 10-3 S-I. To simulate the cyclic deformation at higher strain amplitudes, fully reversed load control fatigue tests at a stress amplitude of 520 MPa with sine wave-form at a frequency of 1.0 Hz were conducted to observe the sur face appearance and fatigue crack initiation of the fatigue crack initiation specimens (figure 3). Thin foil specimens for transmission electron micros copy (TEM) were prepared to identify dislocation config urations of the microstructures before deformation. Experimental results Optical microscopy. As expected from the fact that the microstructure of dual-phase steels is mainly deter mined by their transformation paths, there are remarkable differences in the resulting morphologies, as clearly shown in figure 4. After the specimens of group A were furnace cooled from the austenitizing temperature, the initial mi crostructure was blocky hypoeutectoid ferrite with pearlite along its boundaries", + " The martensite volume fractions of network structure (A), fine fibrous structure (B) and blocky structure (C) were determined by a quantitative metallography and found to be 23.1%,23.4%, and 22.6%, respectively. Ap parently, there is very little variation in the martensite followed by furnace cooling and water quenching, respec tively. After heating at 980\u00b0C for 30 min, group C was transferred immediately into a salt bath at 800 DC for 30 min, together with the pre-treated groups A and B. All blanks were then quenched in water concurrently. The tensile and fatigue test specimens, as illustrated in figure 2, were machined from the heat treated blanks. Figure 3 shows the geometry of fatigue crack initiation specimens, whose surfaces were mechanically polished. Tensile tests were conducted on a 100 kN testing ma chine at a strain rate of about 10-2 SI. The fracture surface was examined by SEM. The S-N curves of three micro structures were determined on a 40 kN testing machine, using a 50 Hz tension-zero sine wave form under load control. Fully reversed low cycle fatigue tests were con ducted on the same machine, with a clip-on extensometer of 10 mm gauge length" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001924_sp.j.2096-5796.2019.0016-Figure13-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001924_sp.j.2096-5796.2019.0016-Figure13-1.png", + "caption": "Figure 13 The force/torque loading test bench.", + "texts": [ + " At present, the static calibration methods of multidimensional force sensors mainly include tensioncompression dynamometer method, standard weight method and other standard force generator consisting of force generating equipment and high precision single-axis sensor. Due to the diversity of sensor structures, there is no universal calibration device. Standard weight method is mainly introduced in this section. The calibration test bench was developed by Robot Sensor and Control Lab of Southeast University. Figure 13 shows the force/torque loading test bench. The six-dimensional F/T sensor to be calibrated is fixed on the scalable and rotatable indexing plate, which is fixed on the base of the test bench, and the calibration shaft passes through the central axis of the sensor. The sliding rods on both sides of the test bench can be adjusted and fixed to a certain height, so that the steel wire suspending the standard weights presents a horizontal straight line between the calibration shaft and the pulley, that is, it is parallel to the Y axis in Figure 13. Aiguo SONG et al: Multi-dimensional force sensor for haptic interaction: a review dimensional F/T sensor, respectively[36], and the output voltage values of the corresponding channels are recorded. Ideally, in the static calibration experiment, each force / torque component can be loaded only once under the condition that the applied standard force / torque is unbiased and the sensor is pure linear. However, due to the deviation of loading force/torque and the non-linearity of six-dimensional F/T sensor in the actual calibration process, the error of data obtained by loading only once is large" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003989_icma.2019.8816285-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003989_icma.2019.8816285-Figure5-1.png", + "caption": "Fig. 5 Physical interpretation of compliance of PLCM", + "texts": [ + "2035 10 0.9363 0.9363 1.3939 1.3939 1.3939 1.3939 0 0 0.0101 0.0325 0.7951 0.8674 0 0 0.9999 0.9995 0.6065 0.4976 1 1 0 0 0 0 0 0 11.5825 11.5458 5.9252 6.9795 0 0 1.5107 1 diag h diag \u03b6 .7696 10.0662 9.3661 0.9363 0.9363 0 0 0 0 (33) The coefficient of egienscrew can be obtained using (27) shown below. 54.1075 4.1075 0.0174 0.0174 0.0174 0.0174 10 \u03b5 (34) It can be seen from Table 2, that C can be interpreted using a body supported by six screw springs ci in the directions of eigenscrews as shown in Fig. 5. Thus, each screw spring is defined by its spring constant ci, and pitch hi, unit vector ei and position vector ri. The first two springs are serial along axis z. The other four springs lie in a common plane perpendicular to axis z. Moreover, the six springs intersect at point Pc (0,0,-11.5972) the center of compliance. There is a compliant axis when a force produces a parallel linear deformation, and a rotational deformation about the line of the force produces a parallel couple. The existence of compliant axis is related to the corresponding eigenscrew" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003887_iemdc.2015.7409094-Figure12-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003887_iemdc.2015.7409094-Figure12-1.png", + "caption": "Fig. 12. Prototype and test set-up. (a) Rotor and magnets, (b) Stator, coils and custom built housing without cooling, (c) Test bed.", + "texts": [ + " It is clear that, a) the cheapest option that fulfills the structural requirements in section A, is the 3% work hardened 316L austenitic steel, the cost of which is 120% of the cost of the magnets, b) work hardened steel is significantly cheaper than using a stronger grade of steel, such as Nitronic 50, and c) the state of the art copper beryllium, [3], is 5 times more expensive than the proposed austenitic steel option, and, can constitute a cost equivalent to 600% of the magnet costs, which might, significantly, compromise the performance per cost competitiveness of the ferrite based designs. Due to difficulties in procuring low volumes of work hardened stainless steel, the more readily available Nitronic 50 was used for the prototype testing. Furthermore, to simplify the prototype manufacturing and customize the testing, a prototype with one fifth of the stack length of the original motor is built, Fig. 12(a) and Fig. 12(b). In the stator, as only a static testing was intended at this stage, only two coils belonging to one motor phase, have been wound, Fig. 12(b). Static torque was measured by injecting a DC current corresponding to the peak inverter current, and varying the rotor position for 360 degree electrical period, via the rotary table, Fig. 12(c). The variation of the static torque against the rotor position is scaled for the full length machine with all the coils included, and compared against the 3D FE calculations in Fig. 13. As seen in Fig. 13, the prototype test results match well with the 3D FE predictions. The 5% difference in the peak torque is, majorly, attributed to, a) the deviation of the magnets and laminations BH data from what was assumed in the FE modelling, and b) manufacturing tolerances such as the small gaps between the magnets and the rotor pole, which were neglected in FE modelling" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000008_icelmach.2018.8506886-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000008_icelmach.2018.8506886-Figure2-1.png", + "caption": "Fig. 2. Installation of the motor", + "texts": [ + " Dukalski is with the Institute of Electrical Drives and Machines KOMEL Poland (e-mail: b.dukalski@komel.katowice.pl). T. Jarek is with the Institute of Electrical Drives and Machines KOMEL, Poland (e-mail: t.jarek@komel.katowice.pl). T. Wolnik is with the Institute of Electrical Drives and Machines KOMEL, Poland (e-mail: t.wolnik@komel.katowice.pl). to obtain the best possible motor parameters while maintaining the smallest possible mass. In KOMEL Institute, a prototype of a motor for mounting in a wheel was developed (Fig.1 and Fig.2). It has 40 kW of continuous power and dimensions of the motor (frame size 200) were chosen in such a way that there was a possibility its installation in the vehicle while maintaining the existing bearing and braking system. The prototype motor is based on the bearing and braking system from the new Fiat Panda III, while the external dimensions have been limited by the dimensions of the 17-inch rim (Fig.2). The motor consists of two main components: the rotor and the stator. The source of heat in this type of machine are losses in the rotor and in the stator (winding, core). A significant part of the losses are losses in the stator. In order to ensure adequate heat recovery from this element in the stator's supporting structure, a water jacket was made, and the empty space between it and the winding was filled with thermoconductive resin (Fig.4). The efficiency analysis of various structural solutions of the wheel motor cooling systems B" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000231_s00170-017-0346-6-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000231_s00170-017-0346-6-Figure2-1.png", + "caption": "Fig. 2 Multi-breakage effect. a Scheme. b Photo", + "texts": [ + " Further, in the contribution, bending characteristics refer to a summarized definition of the following terms: springback, bend allowance, bending force, contact points, and sliding distance. The multi-breakage phenomenon lends its name from a pseudo-polygonal shape that the plate should take when it loses contact with a punch of large radius [17]. For large radius bending, the contact zone is initially concentrated between the plate and the punch at the lowest point of the punch; then, it splits or \u201cbreaks\u201d into two zones during the bending process [18]. Figure 2 shows the schematic and photographic representations of the multi-breakage effect. For conventional air bending, one can observe that during forming, the radius of the inner surface of the formed plate is larger than the radius of the punch. However, Fig. 2 shows that for large radius bending, the actual radius of the inner surface of the plate is smaller than the punch radius. In the current section, the bending moment diagrams are depicted according to rules presented in [19]. Conventional air bending is described as a central loading scheme, namely three-point bending. In the early phases of the forming process, the shape of the bending moment diagram is triangular (see Fig. 3a). During the actual plastic forming process, the shape of the bending moment diagram looks like a triangle with rounded sides (see Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003998_s00202-006-0054-y-Figure13-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003998_s00202-006-0054-y-Figure13-1.png", + "caption": "Fig. 13 Wound stator magnetic structure and one of the two rotors of a Single Stator Double Rotor, axial flux, PM machine, equipped with a double layer winding with concentrated coils (data in Table 1, 3)", + "texts": [ + " 4), depending on the number of coils which magnetize the considered tooth or slot. From (19), (20), (28), the phase equivalent synchronous inductance equals: L = N2 tu.co \u00b7 Wm a2 \u00b7 I2 tc = N2 tu.co \u00b7 Nc \u00b7 Wcycle a2 \u00b7 I2 tc = N2 tu.co Nc 2 \u00b7 a2 [( 4Ntcph \u2212 Nts ) \u00b7 ( \u03c1 + 4 ) \u2212 2 ] (29) Various FEM comparative tests have validated the previously developed model, with small differences (\u00b11\u20132%). In fact: L = 4.50 mH by analytical computation; L = 4.61 mH by FEM analysis. In order to validate the developed model, a SSDR machine has been constructed and tested: Fig. 13 shows the stator of such a SSDR prototype machine and one of the two rotors, during the assembly process. Each tooth has been fixed on a supporting frame, by means of two L-shaped flanges; finally, the whole stator has been enclosed in a epoxy resin shell, in order to give stiffness and robustness to the mechanical structure. It is a machine intended for a limited operation time duration, with a peak power of 40 kW roughly; in fact, it has been designed to convert seismic energy into electric energy, in order to damp the mechanical vibrations induced by earthquakes in industrial and civil buildings [27]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000934_j.still.2017.12.011-Figure10-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000934_j.still.2017.12.011-Figure10-1.png", + "caption": "Fig. 10. Plastic strain distribution with different cutting angles, (a) at 30\u00b0, (b) at 45\u00b0 and (c) at 60\u00b0.", + "texts": [ + " The results were shown at steady-state stage after 4 s of simulation (Figs. 10 and 11). Soil failure can be defined as the permanent deformation of the soil (Stafford, 1984; Rajaram and Gee-Clough, 1988). For experiment, the failure angle was usually measured by connecting the points where the deformation of soil is most remarkable. Therefore, for simulation, the plastic deformation of soil was selected as a measure of soil failure and the failure angle was obtained from plastic strain distribution. As shown in Fig. 10 (a), (b) and (c), the plastic strain increased evidently with the increase of cutting angle. However, as shown in Fig. 11 (a), (b) and (c), the plastic strain kept almost unchanged with the increase of cutting depth. The failure angle \u03b3 is defined as the angle between the direction of the shear deformation zone and the direction of proceeding of the cutting blade. The failure angle decreased as the cutting angle increased, whereas it kept almost unchanged with different cutting depths. These results are in consistent with the experimental ones of Hatamura and Chijiiwa (1976a). The soil chip thickness l reflects the accumulation of soil caused by cutting. Fig. 10 and Fig. 11 show that both the increase of the cutting angle and cutting depth would lead to the increase of the chip thickness. The ratio of the chip thickness to the cutting depth ranges from 1.31 to 1.71 (Table 3). This ratio increased with the increase of cutting angle. However, it remained almost unchanged with the increase of cutting depth. Elijah and Weber (1971) noted that for flow failure with inclined flat blades, the thickness of the soil chip normal to the blade surface was greater than the depth of cut because of soil deformation" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002102_iecon43393.2020.9255013-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002102_iecon43393.2020.9255013-Figure3-1.png", + "caption": "Fig. 3. Schematic drawings of two MLSs. (a) The proposed single-helix MLS; (b) Traditional MLS in [15].", + "texts": [ + " rotor will rotate by 2\u03c0 rad, so the gear ratio can be defined as follows: 2 p G v = = () where \u03c9 is the angular velocity (rad/s) of the rotor, v is the linear speed (m/s) of the translator. It can be seen from (1) that the gear ratio is inversely proportional to the length of one pole-pitch, so a higher gear ratio can be obtained with a narrower pole-pitch. Other parameters such as the air-gap g, pole-pitch \u03c4p (including magnet width wm and iron yoke width wi), magnet thickness hm, as shown in Fig. 3, will also take significant effects on the performance of the proposed topology, which is analyzed in the following section. III. DESIGN ASPECTS OF THE STRUCTURE The performance of the MLS, including the rotor torque and thrust force (also called pull-out force), are significantly affected by structure parameters such as the rotor radius and translator radius, air gap, PM thickness and width, pole pitch. In this section, the performance of the proposed MLS, which is affected by the above parameters, is analyzed by 3-D FEA method" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003064_el:20030070-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003064_el:20030070-Figure2-1.png", + "caption": "Fig. 2 Transfer function of gain-stage", + "texts": [ + " 1 shows the corresponding weights of the raw bits, the ADC architecture, and the digitally calibrated stages. The ten least significant bits come from the overlapped 9 bit-pairs of the 1.5 bit=stage stages. Fig. 1 shows the full-scale range from Vref to \u00feVref divided into 8180 codes by the ADC. The corresponding code of Vref is 0 and that of \u00feVref is 8180. System level Monte-Carlo simulations using Matlab show that including non-idealities the converter achieves <0.05 LSB DNL and INL after calibration. Radix<2 gain-stage architecture and calibration: Fig. 2 shows the input\u2013output transfer function of the first four stages as well as the calibration procedure. The input\u2013output transfer function is as ELECTRONICS LETTERS 9th January 2003 Vol. 3 follows: nout \u00bc \u00f01\u00fe A\u00de nin A Vref in which 1\u00feA is the slope of the transfer function. The over-range value is (1 A) Vref which results in (1 A)=(1\u00feA) Vref tolerable offset budget. In calibration mode, the gain-stage samples the zero input and then connects to the left and right side of the transfer function, respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000298_1.1807415-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000298_1.1807415-Figure3-1.png", + "caption": "Fig. 3 Temperature and viscosity distribution \u201ecase 2, 0.5 of eccentricity ratio, 9550 rpm of speed\u2026: \u201ea\u2026 temperature, \u201eb\u2026 dynamic viscosity", + "texts": [ + " Consider a cylindrical journal bearing as described in example 1 and Fig. 1. Here the lubricating oil is ISO VG32 ~equivalent to DTE-LIGHT!. Fluid Properties. Two cases are calculated in this thermal analysis. The first case has an assumption of constant fluid properties, which is typical in normal lubrication analysis. The second case assumes the fluid properties are a function of temperature, which is a more realistic simulation. These two cases are described in Table 8. Simulation Results. In Fig. 3, the profiles of temperature and viscosity for case 2 at 0.5 of eccentricity ratio and 9550 rpm of speed are displayed. The performances of the bearing both for case 1 and case 2 over the entire range of eccentricity ratios are shown in Fig. 4. For the purpose of comparison, the results by VT-EXPRESS are also displayed. Figure 5 is the maximum and average temperature curves of case 2. Availability for Extended Thermal Analysis. The possibility to conduct further thermal solutions where the model to be analyzed includes both fluid film and solid elements, such as the bearing housing or journal, has been examined" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003952_978-3-642-33457-3_7-Figure7.11-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003952_978-3-642-33457-3_7-Figure7.11-1.png", + "caption": "Fig. 7.11 Stator and rotor geometry with dimensions in mm", + "texts": [ + "9) where le is the sleeve thickness, g is the mechanical air-gap and lw is the thickness of the winding area. 4. The mechanical air gap is set at 0.5 mm 5. Eventually, an additional constraint of equalizing copper and iron losses is added, yielding somewhat balanced distribution of loss in the stator: PFe = PCu,DC . (7.10) 6. Naturally, saturation limit for the stator core is imposed. This limit hardly affected the optimization since saturation flux density in the core is, for the given magnet material, too high to be reached. Optimization results are given in Table 7.1 and presented in drawings in Fig. 7.11. Optimization in the previous subsection determined, practically, magnetic field in the machine, the available space for conductors and the number of conductor turns. In the second optimization step optimal conductors that fit the available winding space were selected so as to minimize copper loss for the estimated phase currents and the maximum current (electrical) frequency (3.33 kHz). The optimization procedure is quite similar to the one presented in [22].2 In order to alleviate skin- and proximity effect losses in a transformer or high-speed motor, parallel/stranded conductors are used" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001853_ee.1948.6444489-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001853_ee.1948.6444489-Figure2-1.png", + "caption": "Figure 2. Cross section of a salient pole a-c generator with recirculating air cooling system and built-in cooler", + "texts": [ + " which could stand abnormal temperatures, it would be unsatisfactory from the standpoint of operating personnel. In such cases the so-called induced draft system of ventilation may be used. An actual shipboard installation of this type is shown in Figure 1. Two d-c machines have a common exhaust duct which draws the hot air out of the front end of the machines and discharges it above deck. Another system used in shipboard installations is to enclose totally the machine and use a recirculating air system. A built-in heat exchanger is used in this case to dissipate the losses. Such a system is shown in Figure 2 which shows a cross section of a salient-pole a-c generator, self-cooled with a recirculating air system and built-in cooler. Auxiliary equipment on shipboard often is located in such a manner that serious limitations are imposed on ventilation. Motors must have enclosures varying from dripproof and splashproof to totally enclosed. Totally enclosed motors are used for deck machinery and for locations where atmospheric conditions are unusually bad. In such cases totally-enclosed fan-cooled motors are used" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002818_mop.32365-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002818_mop.32365-Figure1-1.png", + "caption": "FIGURE 1 The proposed slot antenna including its various parameters", + "texts": [ + " In some applications, the slot antennas are milled directly in the broad wall of the available standard rectangular waveguides. Then, the waveguides are joined to each other by welding. In such applications, the suggested slot antenna can be employed as straightforwardly as the conventional longitudinal slot antennas, which have offsets from the waveguide centerline. In the following sections, the structure of the suggested slot antennas is described along with the method of employing them in the linear and the planar antenna arrays. Figure 1 shows the suggested slot antennas. According to the figure, there is a longitudinal slot in the top broad wall of the rectangular waveguide. The slot does not have any offset from the waveguide centerline. As the figure shows, two parallel posts are placed near each tip of the slot and the posts have offsets from the waveguide centerline. The posts are symmetrically placed about the plane bisecting the slot. It is evident from the figure that the suggested slot antennas have many parameters, which undoubtedly have effects on the radiation characteristics of them" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002382_bf02654110-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002382_bf02654110-Figure1-1.png", + "caption": "Fig. 1--Schematic diagram of an inert atmosphere water-cooled copper reactor for thermite smelting.", + "texts": [ + " We have chosen the EDXRF technique to meet the analytical requirements because of its capability to analyze samples of ore, minerals,[ 41 metal, and alloys in different forms, such as powder, sponge, as-smelted, or as-cast, to obtain rapid multielement analyses with ease. Rapid analyses of thermite feed and product by this technique have aided in the appropriate alterations of the charge constituents to obtain optimum charge composition. Alloy smelting experiments were conducted in a oneend-closed water-cooled copper reactor shown schematically in Figure 1. This kind of specially designed copper reactor was chosen instead of a conventionally used refractory lined reactor to prevent contamination of the product from the refractory lining and to avoid regular maintenance, as well as relining, of the reactor. The inside portion of the reactor was slightly tapered to facilitate the collection of alloy. All the constituents of thermite charge except reductant aluminum were appropriately heated to ensure complete removal of moisture. This stage is essential from the point of view of the safety of operation and the success of the smelting campaign" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001977_978-981-15-2926-9_1-Figure12-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001977_978-981-15-2926-9_1-Figure12-1.png", + "caption": "Fig. 12 Radiation pattern of antenna", + "texts": [ + " The ground plane, acting as an impedance matching circuit, tunes the input impedance and the operating bandwidth, while the feed gap is varied [5]. To summarize, the optimized feed gap (g) is observed to be at g = 1.6 mm. Figure 11 shows the effect for position of hexagonal notches on the ground plane from the center of the feed line to the center of semi-hexagonal notch. From Fig. 11, it is clear that when position is P5 = 6.5 mm it gives wider IBW due to greatly increased capacitive effect. Figure 12 shows the effect for radius of semi-hexagonal notches. When the radius R = 4 mm gives wider IBW as along notches current path length increases and best matching occurrence is compared to other. Figure 13 shows the effect of feed gap betweenCPW-fed ground and quarter-wave transmission line. From Fig. 13, it can be seen that when the feed gap length x is 0.55 mm antenna gives wider bandwidth due to impedance matching between feed and ground. Figure 14 illustrates the S11 curves by varying the value of the length of closedloop square ring P8. As expected, the S11 at lower band has shifted to the lower Fig. 12 Simulated return loss against frequency for the proposed planar inverted C-shaped monopole antenna with various R radii CPW-Fed Triple-Band Circularly Polarized Printed Inverted \u2026 169 Fig. 13 Simulated return loss against frequency for the proposed planar inverted C-shaped monopole antenna with various X lengths Fig. 14 Simulated return loss against frequency for the proposed planar inverted C-shaped monopole antenna with various P8 lengths of loop resonating frequency, as well as higher resonance frequency also shifts toward the lower region as we increase the length" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000637_6.1992-1178-FigureI-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000637_6.1992-1178-FigureI-1.png", + "caption": "Figure I - Passive Disk Suspension System (Side vicw) v", + "texts": [], + "surrounding_texts": [ + "AIM-92-1178\nDYNAMICS AND CONTROL OF A LARGE DISPLACEMENT SUSPENSION SYSTEM FOR GROUND TESTING OF FLEXIBLE SPACE STRUCTURES\nL\nMercedes C. Reaves Aerospace Engineer\nSpacecraft Dynamics Branch NASA Langley Reseach Center\nHampton,Virginia 23665\nMeng-Sang Chew Associate Professor\nDept. of Mechanical Engineering and Mechanics Old Dominion University\nNorfolk,Virginia 23529-0247\nJer-Nan Juang Principal Scientist\nSpacecraft Dynamics Branch NASA Langley Research Center\nHampton,Virginia 23665\nSteven H. Chiu Engineering Student Trainee Spacecraft Dynamics Branch\nNASA Langley Research Center Hampton,Virginia 23665\nL!htG%l\nAs the development in flexible space smctnres expands, the need for an efficient and accurate suspension system for ground tests increases accordingly. In this article, a Disk Suspension System was developed. Analytical results describing the dynamics of a passive and active suspension system were generated given realistic parameters of the system. These results for the passive case, were then compared with the data obtained from an experimental model with initial displacement and random forcing inputs. The dynamics of the system was analyzed. Results show that the experimental model resembles closely with the analytical model and thcrcfore strongly suggests feasibility of such suspension system.\n1. I n t r m ion\nSpacecrafts have generally k e n considered and designed as rigid structures. However in recent years, there is growing interest in large-scale space structures that may no longer be considered rigid. Some of the structures currently under consideration are: the Mobile Satellite,\nthe Large Deployer Reflector, the Geostationary Platform, and the Large Optical Interferometer. This change in focus drives research into the realm of flexible structures and therefore presents the basis for various forms of dynamic analysis and experimental testing on the ground.\nAttainment or a close approximation to a zero-gravity environment is a necessary condition for ground testing of low-frequency flexible space structures. Suspension systems must carry the payload weight without altering the free-free boundary conditions of the structure and in some situations, accommodate large dynamic displacements as well. Several suspension devicesi-l2 have been proposed and developed in recent years including approaches such as long cables, air-pads, drop towers, parabolic flight maneuvers, pneumatic-electric devices, springs and neutral buoyancy techniques. All the techniques and devices have met with varying degrees of success and discussions of thc limitations of some of these approaches are presented in 13-14.\nThe main thrust of this article is the experimental construction and testing of a Disk Suspension\n\" Copyright c1991 by the American Institude of Aeronautics and Astronautics.Inc. No copyright is asserted in the United States under Title 17, US. Code. The US. Government has a royalty-free license to exercise all rights under the copyright claimed herein for Governmental purposes. All other rigths are reserved by the copyright owner.\n1\nD ow\nnl oa\nde d\nby K\nU N\nG L\nIG A\nT E\nK N\nIS K\nA H\nO G\nSK O\nL E\nN K\nT H\no n\nFe br\nua ry\n1 1,\n2 01\n6 | h\nttp ://\nar c.\nai aa\n.o rg\n| D\nO I:\n1 0.\n25 14\n/6 .1\n99 2-\n11 78", + "System13 to determine its feasibility for ground testing of flexible space structures. Analytical results describing the dynamics of a passive and an active suspension system based on realistic parameter values ae presented. These are then compared to experimental results under various initial conditions and random excitations. The dynamics of the system, as well as its conuol, have been investigated and are reported in this articlc. The following section discusses on the suspension system in the absence of feedback control.\n2. Passive Disk Susoe nsion Svste m\nThe investigation presented in this section on the passive disk suspension system is based on experimentally determined paramctcr values for the intended purpose of comparing experimental and analytical results to verify the feasibility of the Disk Suspension System described in l3.\nFigures I and 2 show a passive disk suspension system. The suspension system consists of a noncircular disk with a specially designed profile, a torsional spring, a thin cable, a small pulley and a test article. The test article freely hangs on a cable wound around thc circumference of the noncircular disk and passing through a pulley located a distance Ra vertically below the rotational axis of the disk. A torsional spring is introduced at the disk rotational axis to prevent any unconstrained downward motion of the disk due to gravity. This torsional spring winds up as thc disk rotates so that the torque produced by the weight of the suspended article about the axis of rotation of the disk is exactly counterbalanced by the torque induced in the torsional spring about that Same axis. Therefore, for initial static equilibrium the torsional spring must have an initial pre-wound angle 0 given by : 0 .\n( 1 1 Bo = R b * W I k,\nwhere Rb is the perpendicular distance from thc disk rotational center to the cable, W is the weight of the test article and ks is the torsional spring constant.\nWhen the test article is given an arbitrary displacement, resulting in the disk rotating an angle 0, then to maintain static equilibrium, the moment about the disk axis due to the test article weight must equal the torsion in the spring. Le.,\nwhere R is the moment arm at that new position. Close inspection of equations (1) and (2) shows that as the disk angle increases, the moment arm R also increases. Therefore, to balance the increasing torsional\nspring torque, while keeping the tension on the cable constant, an increase in the moment arm is required which results in a spira-shaped disk.\nThe coordinates of the disk profile arc generated using envcloped theory and the principles of kinematic inversion as described in 13. The design of the disk profile is based on the suspcndcd load as well as the torsional spring stiffness ks chosen, with a convexity constraint placed on the profile so that the cable can wind around the resulting disk profile. Once designed, the disk can bc reused for a different suspended weight W, so long as the ratio W k s , is preserved. This suspension system directly suspends a test article in static equilibrium for all vertical positions within the designed range of travel. The weightless effect of a test articlc can therefore bc simulated on the ground.\nThe problem to be addressed during this investigation is whether the suspension systcm will simulate the weightless effect of the test articlc in a dynamic situation. The other problem to be addressed is whcthcr the presence of inaccuracies in construction, asscmbly, and in design parameters, ccntribute to the scnsitivity and applicability of the suspension system for actual ground testing of flexible space structures.\n-\n2\nD ow\nnl oa\nde d\nby K\nU N\nG L\nIG A\nT E\nK N\nIS K\nA H\nO G\nSK O\nL E\nN K\nT H\no n\nFe br\nua ry\n1 1,\n2 01\n6 | h\nttp ://\nar c.\nai aa\n.o rg\n| D\nO I:\n1 0.\n25 14\n/6 .1\n99 2-\n11 78", + ",Noncircular Disk\nnr\nrJB&ng\nFigure 2 - Passive Disk Suspension System (Front view)\n2.1 Dvnamics of a Disk SnsDe nsion Svste m in the Absence of Feedback Co ntrol\nThe flexible StruCNle in this investigation is represented by a discrete-lumped parameter system consisting of two masses and a connecting spring as shown in Figure 1. The Disk Suspension System is connected to mass #1 by means of a cable. The total system is a nonlinear two degree-of-freedom dynamic system. Although the test article is linear, the nonlinearities arise from the geometry of the disk profile. A derivation of the dynamic equations of the system is presented in l3. The equations of motion of the system in terms of the\ngiven by :\ni/ generalizedcoordinates I 1 and I 2 of the masses, are\n(4) m2f2-k2(f , - 1 2 ) = 0\nwhere m l and m2 are the masses of the two rigid bodies, k2 is the spring stiffness between the two masses, k, is the torsional spring constant, IC is the moment of inertia of the noncircular disk and W is the weight of the test article. Note that in equation (3) with the absence of IC the system equations reduce exactly to those for a system with two masses connected by a spring in space. Therefore this equation shows the nonlinear conmibution of the disk inertia IC to the i d\ndynamics of the suspension system. Equations (3) and (4) can also be writen in terms of the generalized coordinates 2 and 8 as follows : I\nI, + + B,)']8+ [ m k 2\nwhere,\nm2g k,B3 3k,8,B2 \".( 8t2 + k, 8, - - W 2w 2w\nind ksk2 82 f2 = - 2 w\nThis form of the dynamic equations is used in the simulation when damping and/or active feedback control are incorporated into the simulations where the angular velocity and angular displacements of the disk are needed as discussed later in the paper.\nThe following table contains the parameter values used for the dynamic model simulation .\nD ow\nnl oa\nde d\nby K\nU N\nG L\nIG A\nT E\nK N\nIS K\nA H\nO G\nSK O\nL E\nN K\nT H\no n\nFe br\nua ry\n1 1,\n2 01\n6 | h\nttp ://\nar c.\nai aa\n.o rg\n| D\nO I:\n1 0.\n25 14\n/6 .1\n99 2-\n11 78" + ] + }, + { + "image_filename": "designv6_24_0002574_s0266-3538(02)00044-1-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002574_s0266-3538(02)00044-1-Figure2-1.png", + "caption": "Fig. 2. Cross section portion of the pole: (a) geometrical sketch; (b) adopted FE model.", + "texts": [ + " The relevant literature [15\u201317] offers two possible approaches to the mechanical problem: (i) a micro-mechanics approach which takes into account the microscopic nature of the material, looking at the properties of fibres and matrix as well as at the effective nature of the interface between them; (ii) a macro-mechanics approach which analyses the behaviour of the composite as a whole exhibiting, generally, an orthotropic behaviour. Hereafter a micromechanics approach is utilised to predict the moduli of the composite from those of the constituent materials by means of the r-o-m; while phenomena like interaction between contiguous layers are treated, for simplicity, at a macro-scale level. Thereafter we refer to the FE model of the 80 mm pole specimen used to calibrate the numerical simulation. As described in Section 2 (see also Fig. 1) the pole presents five different layers. Referring to Fig. 2, where a portion of the generic pole cross section is schematically depicted, the layers are numbered sequentially going from the outer layer (ply number 1) to the inner one (ply number 5). The plies #1, #3 and #5, with fibres oriented at W=0 with respect to the longitudinal axis of the prototype, are discretised with \u2018\u20183D-solid\u2019\u2019 isoparametric elements with 27 nodes per element, while plies #2 and #4, with wound fibres at W = 76 respectively, are modelled through 2D 9-node isoparametric \u2018\u2018shell elements\u2019\u2019" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001959_aici.2009.90-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001959_aici.2009.90-Figure2-1.png", + "caption": "Figure 2. 1D transducer imaging unit structure and geometry: 1D transducer consists of a linear array of photodetectors and a cylindrical lens.", + "texts": [ + " In order to identify the matching between different image points, at least four dimensions of information have to be provided, for more detailed analysis see [3]. Therefore, we add a fourth linear CCD as shown in figure1 and deduce its constraint for identifying the correspondence among four linear images. 2.2. Correspondence Constraint 978-0-7695-3816-7/09 $26.00 \u00a9 2009 IEEE DOI 10.1109/AICI.2009.90 422 A Linear CCD is composed of a single-line CCD sensor with a cylinder lens. Each image point and its nodal axis determine an imaging plane, as shown in figure 2. We have learned in solid geometry that three arbitrary planes intersect at a point, but four arbitrary planes normally do not intersect. Therefore, the constraint for identifying corresponding image points can be interpreted as whether the four image planes determined by each point intersect in space at one point. As we have recently deduced in paper [3], the plane equation with a linear CCD can be expressed as ( ) ( )0 0 1 X Y f R T Z \u03bb \u239b \u239e \u239c \u239f \u239c \u239f = \u239c \u239f \u239c \u239f \u239d \u23a0 (1) Where, f is the focal length, \u03bb is the linear image coordinate of a image point, and (R T) is the 3D transformation (rotation and translation) from the world coordinate system to the linear CCD coordinate system" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002693_amr.655-657.542-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002693_amr.655-657.542-Figure1-1.png", + "caption": "Fig. 1 2 D finite element model of electric spindle Fig. 2 The model after dividing network", + "texts": [], + "surrounding_texts": [ + "The losses in electric spindle mainly consist of core loss, copper loss, mechanical loss, stray loss and so on. Calculation of these losses has different methods, shown in the literature [6], the most widely used calculation method of the losses in electric spindle is the finite element method. The thought of finite element is to divide the motor model from the relatively large domain into small triangle region, then calculate and process to make the calculation more simple and accurate. In order to fully reflect the influence of harmonics on losses and make the calculation results more accurate\uff0cgenerally the non-sinusoidal output power is from frequency converter when the electric spindle runs. This article bases on time step finite element method coupling with field-circuit to simulate running state. At the same time, the electric spindle uses non-sinusoidal output power from frequency converter. According to the structure parameters of ceramic electric spindle in author\u2019s lab\uff0cestablish the finite element model of motorized spindle based on Maxwell. The motor model in this paper is three-phase four-pole asynchronous ceramic motor, the structure parameters are shown in Table 1. Combined the Maxwell with Simplorer, establish the time step finite element model , shown in Fig.3. Motor\u2019s three-phase voltage is applied by external circuit. Adopt the SPWM to control inverter switch frequency, switch the dc voltage source into three-phase voltage source which phase difference is 120 degrees, then output as the motor incentive. When Simulating, firstly load torque of 1N\u2022m to the motor, then observe the value of electric spindle losses. By changing the value of the torque, record the change of loss." + ] + }, + { + "image_filename": "designv6_24_0001964_iet-map.2017.0644-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001964_iet-map.2017.0644-Figure3-1.png", + "caption": "Fig. 3 Geometries of antennas considered in numerical study (a) Antenna I [28], (b) Antenna II [29], (c) Antenna III [30]", + "texts": [ + " Comparison of the algorithm working with and without this procedure allows us to demonstrate performance improvement of the optimisation process. Achieving such an improvement is the one and only purpose of this work. The purpose of this section is to provide numerical verification and benchmarking of the optimisation approach formulated in Section 2. We consider three examples of planar UWB antennas, which are optimised using both the proposed methodology and the reference technique described in Section 2.2. Our verification case studies include three UWB antennas shown in Fig. 3. Antenna I [28] and Antenna II [29] are implemented on Taconic RF-35 substrate (h = 0.76 mm, \u025br = 3.5, tan\u03b4 = 0.0018); Antenna III [30] is implemented on 1.55-mm-thick FR-4 substrate (\u03b5r = 4.4, tan\u03b4 = 0.018). EM simulation models for all of the considered structures are implemented in CST Microwave Studio and evaluated using its time-domain solver. The list below provides information about design parameters and other relevant details for Antennas I\u2013III: \u2022 Antenna I: adjustable design parameters xI = [L2 L3 L4 L5 w2 w3 w4 w5 w6 w7]T; all in mm, are fixed; EM model: \u223c940", + " The algorithm optimises the antenna for best matching in the vicinity of this point with constraints on the area, cf. (6). In the example given in the picture, the algorithm is terminated after finding point x(5), which could not be improved. Consequently, the lastfound feasible point x(4) is returned at the final outcome of the optimisation process IET Microw. Antennas Propag., 2018, Vol. 12 Iss. 8, pp. 1273-1278 \u00a9 The Institution of Engineering and Technology 2018 1275 The antennas shown in Fig. 3 have been optimised using the reference algorithm of Section 2.2 as well as the proposed methodology of Section 2.3. The maximum in-band reflection threshold was set to Smax = \u221210 dB for Antennas I and II, and to \u22128 dB for Antenna III. The reference designs are those optimised for best matching. The reason for setting up a specific level for Antenna III is that its minimum-reflection (reference) design barely exceeds \u221210 dB limit which does not leave room for subsequent size minimisation. In order to permit size minimisation, certain margin is necessary (in other words, the initial design should be within the feasible region). At the same time, it should be emphasised that a particular reflection constraint selected for a particular antenna structure is irrelevant from the point of view of demonstrating the proposed algorithmic solutions. The results have been gathered in Table 1 and in Figs. 4\u20139. The reflection characteristics of Fig. 3 indicate that the proposed optimisation technique allows for better control of the reflection response than the benchmark technique. More specifically, the maximum in-band reflection is closer to the required level Smax. Consequently, minimum-size designs can be identified more precisely than for the reference algorithm. As a matter of fact, for Antennas I and III, the proposed approach yields designs that are considerably smaller than those obtained with the benchmark method (520 versus 580 mm2 for Antenna I, and 255 versus 287 mm2 for Antenna III)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001472_9781118354179.auto089-Figure23-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001472_9781118354179.auto089-Figure23-1.png", + "caption": "Figure 23. Mechanical clutch control system using cable. (Reproduced from Nunney, 1998. \u00a9 Elsevier/Butterworth Heinemann.)", + "texts": [ + " This must be satisfied throughout the necessary displacement of the pressure plate to ensure complete clutch disengagement in all operating and wear conditions. The ratios must be suitably chosen and efficiency sufficiently high to ensure low pedal force (typically under 200 N) and acceptable pedal travel (typically <150 mm). The schematic shown in Figure 22 relates to the mechanical clutch control system. In today\u2019s road vehicles, mechanical clutch control systems have found applications in lower performance cars. Usually, a steel cable is used to connect the clutch foot pedal or hand lever to the bearing releasing fork. Figure 23 (Nunney, 1998) shows a typical mechanical Encyclopedia of Automotive Engineering, Online \u00a9 2014 John Wiley & Sons, Ltd. This article is \u00a9 2014 John Wiley & Sons, Ltd. DOI: 10.1002/9781118354179.auto089 Also published in the Encyclopedia of Automotive Engineering (print edition) ISBN: 978-0-470-97402-5 solution. Inexpensive to manufacture and install, such designs may suffer from various problems in particular when higher forces are required, for more powerful vehicles. Efficiency is typically low because of friction in contacts and joints, the cable stretches, with cable ends (connections) prone to snapping off" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000718_bmei.2010.5639976-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000718_bmei.2010.5639976-Figure5-1.png", + "caption": "Figure 5. Different curved needle shapes caused by section lengths.", + "texts": [ + " The maximum angles of articulations are sequentially noted as 1m\u03b1 , 2m\u03b1 , and 3m\u03b1 etc. An enlarged articulation and its bent status are shown in Fig. 4. Combinations of different L and different m\u03b1 generate different needle bent shapes. Take the needle with two articulations, as shown in Fig. 3, as an example, and suppose the sections are rigid. 1) Different section lengths with equal articulation angles When L0, L1, and L2 are equal, 1m\u03b1 and 2m\u03b1 are both equal to 10 degree, the needle bent shape is given in Fig. 5(a). When L0, L1, and L2 are different, for example, L0 is half L1, and L1 is half L2, generated needle bent shape is given in Fig. 5(b). It can be seen that equal section lengths results in uniform curvature of bent needle. And different section lengths generate different curvatures. Also, shorter head length benefits larger curvature in the front end of the needle. 2) Different articulation angles with equal section lengths If 1m\u03b1 and 2m\u03b1 are not equal and L0, L1, and L2 are equal, generated needle bent shapes are obviously different, too. Fig. 6(a) gives generated needle bent shape when 1m\u03b1 is larger than 2m\u03b1 , namely 1m\u03b1 and 2m\u03b1 are equal to 20 and 10 degrees, respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000309_910213-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000309_910213-Figure6-1.png", + "caption": "FIGURE 6 VECTOR DIGGRAM & COORDINATES", + "texts": [ + " The least influence to the vehicle acceleration 2. Drivability improvement a. Eliminates shock resulted from the clutch onloff operation b. Constant off-grille air temperature in passenger compartment METHOD OF ANALYSIS AND MODELING Figure 3 shows a schematic flow diagram of the analysis. Meshes are generated to prepare for the vibration analysis using FEM. These mesh data are Cviicvdor black' Redinr$oari\"~(21 Tcnck Sllppar Slrell Figure 2 A summary and a designation o f variable displacement comprSessor shown in Fig. 6. By taking the center ofhinge ball as the cos o( sin4 cos U t origin of the coordinate system, the slipper's reciprocat- ,o,ing motion, which is resulted from the drive hub's rota- cos2o(+ sinLo(cos2w t cos+o(+ sinad co; tional motion, is expressed in a unit vector as in equation (1). - - - / -.- F R I C T I C Y / :C;:4-?1:3 I CCEFPIC;~;~; / UET2drl i FIGURE 5 A SUMMARY OF INPUT DRTR RND ANALYI'ZCRL YODEL where o(= wobble angle R= pitch circle radius of wobble ball t,$ t= angle of rotation CL) = angular velocity (Step 2 in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003274_ijvd.2012.047403-Figure17-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003274_ijvd.2012.047403-Figure17-1.png", + "caption": "Figure 17 (a) Electric vehicle; (b) Detail of the induction motor and (c) Power electronic converter and control electronics", + "texts": [ + " Following the results presented in this study, one can conclude that in some case it is better to not achieve low THD in currents, but to minimise the total losses (taking into account that these losses are produced by the motor and the electronic converter). Because of the similarity of the results of the experimental test and the simulation in PSIM, using the induction motor, it is possible to draw an analogy with the PMSM comparing the test carried out in PSIM platform (Figure 10) to the future experimental results (currently developing). Next steps in this study will be to realise the same trial in a real electrical vehicle (as the one shown in Figure 17) to determine the real influence of selecting this switching frequency on other key factors, as for example, noise emitting and its impact on human comfort when driving electrical vehicles, and also the impact on increasing vehicle autonomy. This work was supported by the Spanish Ministry of Science and Innovation under the research project PSE-370000-2009-23 co-financed by FEDER. Adam, A.A., Gulez, K., Aliskan, I., Altun, Y., Guclu, R. and Metin, M. (2010) \u2018Steering DTC algorithm for IPMSM used in electrical vehicle (EV)- with fast response and minimum torque ripple\u2019, Paper presented at the 11th IEEE International Workshop on Advanced Motion Control, 21\u201324 March, Nagaoka, Japan, pp" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002699_icorr.2015.7281298-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002699_icorr.2015.7281298-Figure1-1.png", + "caption": "Fig. 1. A prototype of the novel PAU for attendant propelled wheelchair", + "texts": [ + " This function can be implemented with the admittance control applied in our previous research [9], in which, the target speed is not directly pre-selected, but is generated by the exerted force of the attendant via a mechanical admittance model (this will be briefly discussed in Section IV). With our control approach, the novel PAU not only can guarantee the safety of the attendant and the seat occupant but also can make the attendant much easier and more comfortable to maneuver the wheelchair. Based on the above development concept, the prototype of our novel PAU in development is shown in Fig.1. The structure of the PAU seems the same as the conventional ones on market, which usually consist of a handle control panel, a control unit, a battery pack, and a drive unit. But in our new PAU, there are some significant differences from the conventional ones. 1) Easy attachment and detachment: To satisfy our first design concept, we consider what is the part existed in almost all of the manual wheelchairs and the place of the part should be easily accessible. We find this part is tipping lever as shown in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001472_9781118354179.auto089-Figure13-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001472_9781118354179.auto089-Figure13-1.png", + "caption": "Figure 13. Typical diaphragm spring clutch design. (a) \u201cEngine side\u201d view. (b) \u201cGearbox side\u201d view. (Reproduced with permission from ZF Friedrichshafen AG.)", + "texts": [ + " With the introduction of engine start-stop technology and mild hybrid vehicles it has become viable to offer an \u201coff the shelf\u201d solution (by ZF) integrating the clutch and electric motor, as shown in Figure 12. Such solution can work well even with manual transmission, with the clutch and motor functions complementing each other and reducing vehicle fuel consumption and emissions. This is the most commonly used design and the presentation here will be concentrating on the typical assembly and component designs and materials. Figure 5 shows a cross section of the typical diaphragm clutch design, with all components of the assembly. Figure 13a show a photograph (from ZF Sachs) from the \u201cengine side,\u201d including the friction disc. Figure 13b shows \u201cgearbox side\u201d view. Figure 14 (from ZF Sachs) shows the power flow\u2014when the clutch is engaged and disruption to power flow when the clutch is disengaged. For the disengaged clutch, it can be clearly seen that all components of the clutch assembly rotate with the engine flywheel, with the exception of the friction disc. The main clutch components shown in Figure 5 are for the \u201cpush-type\u201d clutch. This is the most common type, but it Encyclopedia of Automotive Engineering, Online \u00a9 2014 John Wiley & Sons, Ltd" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001392_j.ergon.2004.06.002-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001392_j.ergon.2004.06.002-Figure1-1.png", + "caption": "Fig. 1. Sitting measurements: (a) arm length; (b) elbow height; (c) shoulder height.", + "texts": [ + " We use the data obtained in a previous work about anthropometric variables (Lobato, 1997) over a randomised sample that involves 327 workers of the South East of Spain (202 men and 125 women), aged between 16 and 64 years. All variables are defined in accordance with the Norm ISO/DIS 7250.2 (1992) and EN/979 (1995). From their definitions, we can derive that the measurements of shoulder height minus elbow height must be equal to shoulder\u2013elbow length for each person, and we define the se index for standing position, similar to relative error definition, as se \u00bc 100 \u00f0\u00f0shoulder height elbow height\u00de arm length\u00de=arm length\u00de and ses index when measures are obtained in sitting position (Fig. 1). We consider a valid measure when the indices se and ses are less than 77% due to the normal distribution curve obtained (Fig. 2). These values are at the end of the tails of the curve. For subjects whose indices exceed the reference values, the measure must be taken again, taking great care over the results. Those threshold values are chosen according to the sample size and experimental conditions of our anthropometric study. Thus, we obtained a confidence interval over 95%. Obviously, in different experimental conditions other limits can be used" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001472_9781118354179.auto089-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001472_9781118354179.auto089-Figure7-1.png", + "caption": "Figure 7. Electromagnetic clutch. (Janic\u0301ijevic\u0301, Jankovic\u0301, and Todorovic\u0301, 1987. Reproduced by permission of Faculty of Mechanical Engineering, University of Belgrade.)", + "texts": [ + " Such high forces cause considerable loads on the clutch components and unnecessarily increase torque carrying ability well beyond the required level, negating the clutch safety function of slipping if the torque is too high (e.g., if locking wheels when braking). Some time ago, such clutch designs were popular with some semi-automatic and CVT transmission (e.g., cars manufactured by Daf); however, they are nowadays not used as main clutch on road vehicles, with exceptions of some drum type clutches on automatic scooters. Figure 7 (Janic\u0301ijevic\u0301, Jankovic\u0301, and Todorovic\u0301, 1987) shows an electromagnetic clutch, where a large electromagnet is used to provide a normal force for (dis)engaging the clutch. The main advantage of this design is that torque transmitted by the clutch can be easily controlled by electric current control, not requiring any physical force. No doubt, there are thermal aspects and other problems that may diminish the gains. As a result, these are old designs, not used on modern cars. The exception is control of the A/C compressor clutch using somewhat different design and requiring to transmit relatively low torque, only in an \u201con\u201d or \u201coff\u201d position without the need for controlling the engagement", + " Reduced number of components increases Encyclopedia of Automotive Engineering, Online \u00a9 2014 John Wiley & Sons, Ltd. This article is \u00a9 2014 John Wiley & Sons, Ltd. DOI: 10.1002/9781118354179.auto089 Also published in the Encyclopedia of Automotive Engineering (print edition) ISBN: 978-0-470-97402-5 reliability and reduces costs. Finally, it is relatively straightforward to fully automate such a solution. In addition to the above-mentioned modern pneumatic clutch actuation system (Figure 28), which can be automated, two other systems shown earlier, centrifugal clutch (Figure 6) and electromagnetic clutch (Figure 7), can also be relatively easily automated. Both of these are older designs not currently used. Instead, sophisticated hydraulic and electric systems have been developed, as shown in Figures 29 and 30. Obviously, such systems require outside power to operate, that is, hydraulic pressure for the (LuK) system shown in Figure 29 and electric current for the (ZF) system shown in Figure 30. Fenton (1996) gives more detailed insight into the evolution of automated clutch control systems. Figure 31 shows a clutch schematic and basic dimensions" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000810_0470871199.ch11-Figure11.14-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000810_0470871199.ch11-Figure11.14-1.png", + "caption": "Fig. 11.14 Scanned beams for an infinite array antenna. a. Plane wavefront direction for a beam scanned to \u03d1 = \u03d10. b. Plane wavefront direction for a beam scanned to \u03d1 = \u2212\u03d10. The polarisation is perpendicular to the plane of scan.", + "texts": [ + " Let\u2019s assume that we have an infinite, regularly spaced, planar array antenna of identical elements. This means that in the two transverse directions there are no edge elements, only centre elements exist. The infinite array assumption allows us to assume all elements to be identical with respect to mutual coupling effects. Assume now that the array antenna is phased in such a way that a beam is directed towards a certain direction \u03d10, different from broadside and that the polarisation is such that the electric field is directed perpendicular to the plane of scan, see figure 11.14a. By measuring the reflection coefficient of an arbitrary element for the beam scanned to \u03d1 = \u03d10, we can calculate the phased array antenna characteristics as explained in the previous section. Next we assume that a phase taper is applied for scanning the beam into the direction \u03d1 = \u2212\u03d10 and we assume that both phase tapers exist simultaneously. Then we get the situation as shown in figure 11.14b. By superposition, the reflection coefficient of an arbitrary element in the array, emitting both plane wavefronts is identical to that of an element in an array emitting a single beam [10]. In the planes where the two planar wavefronts cross, the net electric field is zero due to the opposite phasing of the waves. Since on a perfect electric conductor the tangential electric field is zero, we may place metallic walls in these planes without this wall placing affecting the array antenna characteristics. If we choose the plane wave directions such that the positions of these metallic walls coincide with planes of symmetry of the array antenna (as is the case in figure 11.14), we may analyse a single waveguide containing a repeatable cell of the array antenna, see figure 11.15. The walls of the waveguide simulator act as mirrors and by reflection of the array cell and its images an infinite array antenna is created. The waveguide simulator needs to be sufficiently large and needs to be terminated into a waveguide load to prevent waves to be incident upon the array cell. The array cell as shown in figure 11.15 contains a single element which is the smallest cell that repeats itself to form the infinite array" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001238_ias.1998.732255-Figure16-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001238_ias.1998.732255-Figure16-1.png", + "caption": "Fig. 16. Circle diagram for IPM motor in linear case", + "texts": [ + " MAGNETIC MODELS FOR MOTOR DESIGN PURFQSE The analysis of the steady state operation of the IPM motor above base speed can readily carried out using the circle diagram theory [Ill. In the id-i, plane, the current and voltage limit together with the constant torque locus are reported, verifylng the more suitable current vector control. A. Linear Case In linear case, in the id-iq plane the current limit defines a circle, voltage limit defines a family of ellipses, with the size decreasing with the speed, while constant torque loci are hyperbolae, as Fig. 16 shows. At point B the maximum torque is obtained by operating at the current limit. Over base speed, the motor is operated with constant current, from B to P. Then, at high speeds, the motor is operated at maximum torque-to-voltage ratio from P to C . B. Non Linear Case A model with constant values for h, and ld (as model illustrated in Section VII, point C) is particularly advantageous for FW analysis and, in addition, for design purpose. In fact, it allows simple equations to be used to describe the motor performance" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000501_1.2048638-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000501_1.2048638-Figure1-1.png", + "caption": "Fig. 1. Schematic cross section of the test structure with a large charge collecting antenna.", + "texts": [ + " The process considered was a model double metal CMOS process, without ion implantations and related masking in order to eliminate the effects of charging during implantations. Test structures were fabricated on p-type, production quality epi wafers, and the thickness of the gate oxide, thermally grown at 920~ in 02 + TCE atmosphere, was about 16 nm. The structure was a small, 10 \u2022 10 ~m, capacitor connected to large (1000 \u2022 1000 ~m) poly (thickness 0.4 Win), metal-1 (0.65 ~m) or metal-2 (1.0 ~m) charge collecting antennas (Fig. 1). A reference capacitor (referred to as reference one in further text) was connected only to small metal-1 and metal-2 pads (125 \u2022 100 ~m). Details of the test structures are given in Table I, where antenna factors (i.e., antenna area to gate area, Aa/Ag, or antenna circumference to gate area ratios Pa/Ag [wm 1]) are given. Contact and pad openings also function as charge collecting antennas during the contact, via, and sputter etch processing steps, and therefore their geometrical parameters are included in Table I" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001007_tasc.2018.2801853-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001007_tasc.2018.2801853-Figure5-1.png", + "caption": "Fig. 5. The cooling Dewar for an HTS field winding.", + "texts": [ + " On the other hand, the developed force in this paper has two components, namely an electromagnetic force (w-hich is produced by the interaction between two magnetic fields: one generated by the armature winding and the other by the HTS field winding); and a reluctance force (which is produced by the variation of the armature reactance). However, the electromagnetic force dominates the reluctance force. In addition, the resultant force can be controlled via adjusting the MMF of the HTS field winding. The cooling Dewar of an HTS field winding adopted a modular racetrack dual-body design. As shown in Fig. 5, the Dewar comprises a dual-body structure where the inner Dewar carries liquid nitrogen and the space between the two Dewars is vacuum. Polytetrafluoroethylene supporters are placed between the inner and outer Dewars to achieve heat insulation. The HTS field winding is immersed in liquid nitrogen for keeping the operation temperature at 70 K. In order to assess the performances of the proposed machine, a 7.6 kW machine has been dimensioned and analyzed. The key design data is tabulated in Table I" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000653_s00500-019-04233-7-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000653_s00500-019-04233-7-Figure3-1.png", + "caption": "Fig. 3 Two-link inflexible robot manipulator", + "texts": [ + " (2010) are dissimilar, to clearly highlight our contributions, a theoretical comparative study between the two control methodologies is given in Table 1. The proposed control method is applied to the models of two MIMO systems, namely a two-link rigid robot manipulator and a helicopter model CE150, and the results are demonstrated by computer simulations. Example 1 In this example, a MIMO plant characterizing a two-link inflexible robot manipulator, which moves in a horizontal plane, has been considered. This robot is shown in Fig. 3. The continuous-time differential equations faithfully modelling this robot are given by Boulkroune et al. (2010a, b): M q\u00f0 \u00de\u20acq\u00fe C q; _q\u00f0 \u00de _q \u00bc s; \u00f082\u00de that can be written also in the following form: \u20acq1 \u20acq2 \u00bc M11 M12 M21 M22 1 U u1\u00f0 \u00de U u2\u00f0 \u00de h _q2 h _q1 \u00fe _q2\u00f0 \u00de h _q1 0 _q1 _q2 with M11 \u00bc a1 \u00fe 2a3 cos q2\u00f0 \u00de \u00fe 2a4 sin q2\u00f0 \u00de, M22 \u00bc a2,M21 \u00bc M12 \u00bc a2 \u00fe a3 cos q2\u00f0 \u00de \u00fe a4 sin q2\u00f0 \u00de,h \u00bc a3 cos q2\u00f0 \u00de \u00fe a4 sin q2\u00f0 \u00de.and a1 \u00bc I1 \u00fe m1l 2 cl \u00fe Ie \u00fe mel 2 ce \u00fe mel 2 1; a2 \u00bc Ie \u00fe mel 2 ce; a4 \u00bc mel1lce cos de\u00f0 \u00de; a4 \u00bc a3 \u00bc mel1lce sin de\u00f0 \u00de: Let x \u00bc q1; q2; _q1; _q2\u00bd T is the state vector, y1 and y2 are the system outputs, u1 and u2 are the control inputs, U ui\u00f0 \u00de, for i \u00bc 1; 2, stand for the actuator nonlinearities (i" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002244_isape.2010.5696431-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002244_isape.2010.5696431-Figure1-1.png", + "caption": "Fig. 1. Antenna with transceiver and on platform.", + "texts": [ + " By the image theory, the magnetic field in the far zone due to M with the conducting surface removed is [2, 3] ok -jkr \u00b0kA , H(r) = - ] e fs2M (r')eJ ror ds' (2) 41tTJT The equivalent magnetic current M is given by M=-zxE over the slot region r' (3) IV. GAIN-BANDWIDTH LIMITATION OF ELECTRICALLY SMALL ANTENNAS Classical theory on fundamental gain bandwidth limitation for antennas constrained by their electrical size has been extensively examined, and is collectively referred to here as the Chu theory [4]. However, there are major shortcomings and ambiguities in the Chu theory when applied to real world problems, as pointed out recently by this author [5, 6]. One problem is the case of an antenna on a platform, as depicted in Fig. 1, where the antenna is generally inseparable from the transceiver/platform. In fact, in some designs the main radiator is the platform or transceiver, not the antenna per se. Thus, the extent and size of the antenna become ambiguous. Also, the Chu theory is valid only for high Q (Quality factor). With the platform becoming part of it, the antenna's effective size is increased and its Q can be lowered beyond the realm of the Chu theory. Nevertheless, the Chu limitation on antenna gain bandwidth due to the electric size of an antenna remains a fundamental obstacle against Top View polarization, of the equivalent electric and/or magnetic current over the spatial angles of interest" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002605_973233-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002605_973233-Figure6-1.png", + "caption": "Fig. 6. Clip Deflection During Insertion", + "texts": [ + " (8) can be calculated as where Mj is the moment at any cross-section, s is the distance along the segment, El is the flexural stiffness of the clip, and L and or are respectively the length and the angle of the segment. For the circular segment under the same loads (see Fig. 5), the compliance ej can be calculated as 3 c: + 3c; g2 = r [2c1c2 + (--- )8 - 2ci sine - 2clc2 cose 2 2 1 2 -CIC~ sin 0 + -(c2 - c;) sin281 4 (13) where p is the initial angle, and 8 is the total angle of the segment. The maximum insertion force is known to occur between the points A and B as shown in Fig. 6, where point A (8 = ei) is the beginning of the arc and point B (0 = 0) is where the largest clip deflection occurs. The clip deflection is the amount of the interference between valve body and clip opening. Solving the location of the maximum insertion force of the clip presents two issues: First, the length of moment arm is reduced as the clip is inserted, effectively stiffening the clip; and second, the angle of the contact surfaces changes during clip insertion. During insertion, the angle of the contact surfaces at point A changes from 0, to 0. To facilitate calculations, insertion is modeled as movement of the clip from 0 = ei to 0 = 0 in arbitrary increments, for each of which the forces required to maintain equilibrium are calculated. The maximum deflection, occurring when the valve body moves down and touches clip at point B (see Fig. 6), is and W rc = r4 + - 2 (1 5) where w is the width of the clip. The clip deflection at any angle 8 can be calculated as where The contact force p now can be obtained from Eq. (8) as To compute the insertion force all the forces applied to the valve body are plotted in Fig. 7, where F,, Nand p are respectively the insertion force, normal force and coefficient of friction. The equilibrium of those forces at vertical and horizontal coordinates leads to and where Bi is the insertion angle as shown in Fig. 7. By substituting Eq. (20) into Eq. (19), the insertion force can be solved as and The maximum insertion force is determined by comparing F, at each increment of 0.1' in angle 8, which is the angle between points A and B (see Fig. 6). The maximum removal force can be obtained in a similar manner. To avoid confusion the insertion and removal forces used in the following mean the maximum insertion and removal forces, respectively. OPTIMIZATION PROBLEM Outimization Parameters - All of the 14 geometrical parameters shown in Fig. 3 can be used as optimization parameters. Due to pre-defined geometrical constraints, however, only eight parameters are used for optimization. These optimization parameters are as follows: x,, r,, x, x , d l , x, 8, and I,, where dl , is the difference of y3 and y," + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002691_2017-01-1767-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002691_2017-01-1767-Figure3-1.png", + "caption": "Fig. 3. Stiffness model of nonlinear bearing", + "texts": [ + " It is represented by normal vector and rotational radii of effective mesh point: (13) The mesh force Fm acts at effective mesh point along effective LOA and is derived from dynamic transmission error ( \u03b4 ) and static transmission error ( e0 ): (14) where km is the effective mesh stiffness calculated from loaded TCA and cm is the mesh damping coefficient. Dynamic transmission error \u03b4 is expressed as: (15) where f(\u03b4 - e0) is the clearance type nonlinear function comprised of three stages depending on the relationship between \u03b4 - e0 and backlash bc : The clearance type nonlinear rolling element bearing model is illustrated in Fig. 3 where {Fbxm,Fbym,Fbzm,Mxm,Mym} is the load acting on the bearing assuming it can rotate freely around z-axis of the local coordinate system. Accordingly, {\u03b4xm,\u03b4ym,\u03b4zm,\u03b2xm,\u03b2ym} is the translational and rotational displacement of bearing model. Based on previous studies [18], bearing stiffness matrix can be generated using partial differential methods assuming rigid raceways as: (17) Bearing stiffness variation is caused by changing orbital position of rolling elements in raceway during operation. Assuming that interaction between rolling elements and inner raceway is pure motion, the position of rolling elements in raceway can be expressed as: (18) where rb is element radius, rd is the pitch radius of raceway, \u03b10 is unloaded contact angle and \u03c9z is the shaft speed. Radial clearance type nonlinearity is also considered in bearing stiffness model (Fig. 3) which can be described using non-linear displacement functions fbq(xlq) where l = p,g; q = x,z. Therefore, the equation of motion in radial translational direction can be rewritten as: (19) where the detailed expression of non-linear function based on quasi-static contact analysis can be formulated as [12]: (20) where \u03b3L is bearing radial clearance, \u03c8j is the angular position of jth rolling element, Z is the total number of rolling elements under load conditions, n is the power coefficient defined by Hertzian contact: for ball baring n = 1" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001910_tie.2011.2161067-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001910_tie.2011.2161067-Figure1-1.png", + "caption": "Fig. 1. Structure of a multiphase slice PMBM. (a) Cross-sectional drawing (N = 6 and p = 1). (b) Developed cross-sectional drawing.", + "texts": [ + ", the solution of the optimal power loss problem and the addition of the solution when a six-phase slice PMBM is under normal operation and when its phase 1 is open circuited; 2) the ripples of the levitation forces and the torque are analyzed; 3) simulations and experiments were redone, and the presentation of the results is improved; 4) simulation results are added to show the case when the slice PMBM is not under the faulttolerant control, which verifies the condition obtained; and 5) experimental results when the motor was accelerated from 800 to 2000 r/min are added to further demonstrate the analysis. The structure of a multiphase slice PMBM is shown in Fig. 1(a), and its developed cross-sectional drawing is shown in Fig. 1(b). There are N teeth distributed equally in the stator and N phase coils wound around each tooth. 2p poles of PMs are attached to the surface of the rotor. The motor is designed to work in the linear magnetic region, and the PM shape can be optimized to create a sinusoidal PM magnetic field distribution. In the sequel, the following assumptions are made. 1) The leakage flux is negligible. 2) The fundamental component is considered only. The levitation force and torque models can be derived according to the principle of equivalent magnetic circuits and virtual energy [18], [19]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001901_iceice.2011.5778327-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001901_iceice.2011.5778327-Figure1-1.png", + "caption": "Figure 1. Structure diagram of GIS bus", + "texts": [ + " To improve the present situation of temperature measuring for GIS bus contact, this paper for the first time put forward an online temperature monitoring technology based on FBG. Coupled electromagnetic field and temperature field model was established to get the temperature distribution of GIS bus, which helped confirming the location where FBG sensors fixed. An elaborate FBG sensor with high accuracy and linearity was designed. Wavelength interrogation system was introduced and physical experiment was given at the end of the paper. II. TEMPERATURE DISTRIBUTION OF GIS BUS The bus of GIS consists of three aluminum cylinders, as shown in figure 1. The analysis of this paper is based on the following assumptions: \u2022 The bus is taken as infinitely long, so the model can be simplified as a two-dimensional one. \u2022 Displacement current is ignored. \u2022 All materials have constant electrical properties. According to the assumptions above, both magnetic vector potential and current density have the z-axis component, the eddy current field of GIS bus can be described as the following equations [6], [7]. z z z A A J x x y y \u03bd \u03bd\u2202 \u2202\u2202 \u2202+ = \u2212 \u2202 \u2202 \u2202 \u2202 (1) z zs zeJ J J= + (2) where, \u03bd is reluctance, zA and zJ are the z-direction components of vector magnetic potential and current density respectively, zsJ and zeJ are source current and eddy current" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003174_lawp.2013.2280277-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003174_lawp.2013.2280277-Figure1-1.png", + "caption": "Fig. 1. Structure of the proposed antenna.", + "texts": [ + "7 GHz which is suitable for LTE application. Second, the ( 10 dB) return-loss bandwidth of this antenna is 65.39% from 1.41 to 2.78 GHz, which is larger than the antenna described above. Also, the director array consists of just nine rectangle loops without dielectric; it is easily fabricated for practice use. The rest of this letter is organized as follows. Section II introduces the detailed concept of the proposed antenna. Section III presents the result and discussion. Finally, a conclusion is presented in Section IV. Fig. 1 shows the detailed design of the proposed antenna. It consists of a planar reflector, a driven slot loop antenna, and three director arrays. Although the antenna is made on a planar structure, it is designed on the concept of a conventional Yagi dipole antenna. The slot loop antenna is set as the driven antenna because of its wide impedance bandwidth. The configuration of the driven antenna is shown in Fig. 2. The driven antenna adopts the model presented in [5], which uses FR4 substrate with relative permittivity and thickness mm" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001824_022050-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001824_022050-Figure1-1.png", + "caption": "Figure 1. Front view. Figure 2. Isometric view.", + "texts": [ + " Series: Materials Science and Engineering 912 (2020) 022050 IOP Publishing doi:10.1088/1757-899X/912/2/022050 Upright was designed using 3D modeling software SOLIDWORKS and analyzed using ANSYS workbench for the static structural condition i.e. von-mises stress, total deformation, stress flow. Numerous iterations were performed to get the final optimized model which can encounter various loading condition during dynamics of the vehicle. The parameters required as boundary condition are calculated through logical calculations. Figure 1 and figure 2 shows the front view and isometric view of the upright respectively. 3.1 Static Loads and Boundary Conditions The parameters of the vehicle used for static analysis of the front upright are [4] [6]: Mass of car and driver=300kgs Weight distribution=35/65 Height of center of gravity=350mm Wheelbase=1560mm Braking g\u2019s=1.04g Wheel Rate=38.78 N/mm These are the forces acting on the front upright when the car is accelerating or decelerating. We consider a scenario that the car applies brake to come at a stop and the weight transfer acts on front uprights" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001036_951526-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001036_951526-Figure2-1.png", + "caption": "Figure 2: MPLM overall Configuration", + "texts": [ + " In the area of satellites for example, the earth observation satellite ERS 1 [ I ] has been successfully developed, launched and operated based on the utilization of an existing plat form of the SPOT programme. Another example is the Gamma-Ray satellite of ESA's scientific project INTEGRAL. INTEGRAL is planned to utilise the service Module of the ESA X-Ray satellite XMM, which will be launched first [ 2 ] . T h s paper is dedicated to the ECLS Subsystem (ECLSS) of COF, see figure 1 and MPLM , shown in figure 2 . Figure 1: COF overall configuration The goal is to provide an integrated set of ECLSS equipment to the MPLM which will be common with the corresponding equipment of the COF, to share subsystem verification effort and to lower future operations cost by reductions of types of spares needed and reduced sustaining engineering effort. PRE-REQUISITES FOR COMMONALITY Experience shows, e.g. [2] , that in order to eventually achieve commonality the necessary pre-requisites in general terms are: (i) the projects together with their missions must be Table 1: Synopsis of key requirements sufficiently known in advance, (ii) the full range of requirements must be taken into account, (iii) the common subsystemlequipment is subject to one single contract, (iv) the subsequent project/mission adapts to the (fured) YFs established along with the design, development and qualification of the earlier project/mission" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000104_978-3-319-27338-9_8-Figure8.3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000104_978-3-319-27338-9_8-Figure8.3-1.png", + "caption": "Fig. 8.3 a Block diagram and b circuitry of differential amplifier", + "texts": [ + " The larger the number of stages the larger the power consumption. The optimum number of stages to obtain a high r at a low power consumption is 4 or 5 stages. The ratio r indicates how much the single stage gain bandwidth product can be extended by using a multistage amplifier. For Atotal = 30 the multistage GBW is 6 times higher than the single stage GBW. The maximum value for the ratio r can be fitted by the formula: rmax = 0.36 \u00b7 Atotal 0.8 (8.8) A differential amplifier has advantages over the single-ended amplifier. Figure8.3 illustrate the block and circuit diagram of a differential amplifier. The differential amplifier has two out of phase outputs Vout+ and Vout\u2212. The differential amplifier\u2019s total output voltage is Vout = Vout+\u2212Vout\u2212. So the differential amplifier output swing is two times the single ended output swing. A differential PA has improved immunity to power supply and substrate noise compared to a single-ended PA. The power supply and the substrate noise are coupled equally to both inputs of the differential PA and thus these noises are suppressed as a common-mode signal" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000999_bf00987123-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000999_bf00987123-Figure2-1.png", + "caption": "Fig. 2. E x a m p l e 1 - The spinning top.", + "texts": [ + " Hence, by starting from an initial trial solution ~r, the right solution U can be achieved iteratively by solving the upper linear system to evaluate the increment d U and to update the solution: and the integration can restart for a new time step. The constraint forces have no boundary terms but are only defined inside the finite element. Hence, by considering a multibody formed by N B bodies and N c constraints, the total number of internal and boundary unknowns is: N = 6 . ( N v N B + N v N B + N v N c + N v N c + N v N o ) . (41) The motion of the spinning top in Figure 2 is analysed. The top is modelled as a rigid body constrained at the IRF by means of a spherical joint. The point of suspension is the origin of the IRF and a centroidal principal ERF is considered. Table 1 summarizes the data of the top defined in a coherent system of measure. At the initial time, the embedded axis ~ stays in the plane yz and is inclined by 10 degrees with respect to the z axis. Three typical cases of motion are considered in correspondence to different values of the initial velocity" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001261_ip-h-2.1990.0008-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001261_ip-h-2.1990.0008-Figure4-1.png", + "caption": "Fig. 4 Measured and calculated radiation patterns of a small ASP 0 0 2 5 (a = 0.02 1, R,, = 0.275 I.) for the basic excitation modes V , and V , ~ measured Fig. 5 Calculated current distribution for a large ASP 0 0 0 calculated a = 0.02 I ; R,, = 0.725 1", + "texts": [ + " These properties shall be reviewed here, related to the problem of polarisation stability. The spirals have been analysed numerically using a thin wire model which is treated with the method of moment approach [24]. The calculated current distribution for a small archimedean spiral is given in Fig. 3 for both excitations V , and V , . The geometry of the spiral is defined by a = 0.02 2, R , = 0.025 I , the wire radius R, = 0.004 lb and a maximum radius of the structure R,,, = 0.275 A. The corresponding radiation patterns are given in Fig. 4 which also presents measured values for perfect conductivity of the antenna wire, attenuation of the antenna current can only be caused by the radiation of energy. Consequently, only the forward travelling wave contributes significantly the radiated far-field of the antenna. In this case the right-hand spiral radiates mostly right-hand polarised waves as expected. For V = VI, only the reflected wave is attenuated and the spiral radiates left-hand polarisation in spite of its winding sense. The polarisation of a small spiral for arbitrary excitations is therefore determined rather by the actual excitation than by the antenna shape" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000999_bf00987123-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000999_bf00987123-Figure6-1.png", + "caption": "Fig. 6. Example 2 - The bipendulum.", + "texts": [ + " Three typical cases of motion are considered in correspondence to different values of the initial velocity. Table 2 presents the initial velocities by means of their components in the ERE The test is interesting because of the superposition of a precession motion, characterized by a low angular speed, to the high speed motion of the top around its principal axis ~. This difference in angular speed makes the numerical problem somewhat stiff. The results obtained are in accordance with those of Reference [26] as shown by Figures 3, 4, 5. The second example considered is the bipendulum in Figure 6. It is formed by two rigid bars linked by means of spherical joints at their tips. The origin of the IRF is placed at the first suspension point and centroidal principal ERFs are chosen. Table 3 summarizes the data of the bar defined in a coherent system of measure. At the initial time, the bars are aligned along the negative direction on the y axis and rotate with an angular velocity of 4 rad/s around the Length 1 mass 0 . 4 5 6 5 Jt,-.,,.. 0 . 0 3 0 8 J~,i~ 0 .0001 grav. - 3 2 . 1 7 5 Z Z z axis" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002458_tvt.2012.2186991-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002458_tvt.2012.2186991-Figure4-1.png", + "caption": "Fig. 4. Velocities and forces in a bicycle dynamical model.", + "texts": [ + " b) For any disturbance d(t) that does not take the nonlin- ear system outside Ur(0), the inequality \u2016Lc\u2016\u221e < \u03b3 holds. Numerical results (such as those presented in Section IV) indicate that this neighborhood is sufficiently large, and effective implementation of Theorem 1, for path following of autonomous vehicles in the presence of sliding effects, is feasible. To examine the proposed controller, a model that is not based on rolling without slipping assumption is required. To fulfill this, we use a nonlinear bicycle dynamical model derived from the diagram presented in Fig. 4 (where the coordinates (xc, yc) are attached to the vehicle frame, and (x, y) are inertial). Based on this diagram, the vehicle model that describes its plane motion is given as [28] M(v\u0307x \u2212 vy \u03b8\u0307) =Fx2 cos \u03c6 + Fx1 \u2212 Fy2 sin \u03c6 M(v\u0307y + vx\u03b8\u0307) =Fy1 + Fy2 cos \u03c6 + Fx2 sin\u03c6 Izz \u03b8\u0308 =L2Fy2 cos \u03c6 \u2212 L1Fy1 + L2Fx2 sin\u03c6 (35) where M is the vehicle mass, Izz is the vehicle inertia, vx, vy are the lateral and longitudinal velocities of the vehicle at the center of gravity, Fxi, Fyi are lateral and longitudinal forces acting on each vehicle wheel (i = 1 represents the rear side of the vehicle, and i = 2 stands for the front), and L1 and L2 determine the center of gravity (L = L1 + L2 is the vehicle length). When rear driving is considered (as is the case of this paper), Fx2 = 0, and Fx1 is the vehicle driving force, as transmitted to the rear wheels by the vehicle engine. From Fig. 4, the tire slip angles can be related to the vehicle velocities by tan \u03b11 = L1\u03b8\u0307 \u2212 vy vx , tan (\u03c6 \u2212 \u03b12) = L2\u03b8\u0307 + vy vx . (36) Additionally, the well-known Pacejka Magic Formula is utilized in the dynamical model to characterize the tire behavior [28]. This empirical formula is capable of producing characteristics that closely match the measured data. The coefficient BM is called the stiffness factor, CM is the shape factor, DM is the peak factor, and EM is the curvature factor. Equation (37), shown below, is a non-shift Magic Formula, and each force in (37) results from the two wheels of a common axle (rear or front, represented by the index i = 1, 2)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003450_pedstc.2014.6799443-FigureI-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003450_pedstc.2014.6799443-FigureI-1.png", + "caption": "Fig. I. Configuration of SRM for one pole", + "texts": [], + "surrounding_texts": [ + "to redress Eccentricity Faults consequences. Two main effects of EFs include considerable increase in Unbalanced Magnetic Force (UMF) and torque ripple. Dependeney of UMF and torque ripple on EF level is shown and the possibility of compensating them via controlling currents in facing poles is proved. These objectives are aecomplished through a novel converter which controls currents in poles of every phase such that one pole is responsible for nullifying UMF and the faeing pole is in eharge of torque ripple minimizing. The eontrol algorithm is analyzed in details and through Finite Element Method (FEM) and MATLAB/Simulink is implemented on a 6/4 Switched Reluctance Motor (SRM). Suggested method is designed such that there is no need to detect EF direction and type which makes it cost effective and practical for a wide range of SRMs and applications.\nKeywords- Eccentricity Compensation, Fault Control, Unbalanced Magnetic Force, Torque Ripple, Finite Element Analysis, Switched Reluctance Machines.\nI. INTRODUCTION\nThe advent of power electronics and significant advancements in this area has made SRMs to be a good candidate for household and industrial applications due to their simple structure and fault tolerant nature [1] .\nEFs impair the uniform air gap of the motor and as flux pattern and flux density distribution change, excessive vibration, acoustic noise, bearing wear and UMF will be the inevitable concomitants[2]. EFs can be detected by monitoring several parameters of the motor such as torque, stator current, temperature and inductance [3, 4]. Also mutually induced voltages have introduced as a good criteria for detection of EFs in [5].\nUMF exerted on rotor in eccentric conditions is evaluated in many literatures such as [6]. To compensate this radial force some remedial strategies have been proposed mainly concerning some modifications in windings structure. The substantial effect of parallel stator winding in induction and synchronous machines has been discussed in [7-9].\nThe idea of different winding methods and their features under various type of faults in a 6/4 SRM has been reported\n978-1-4799-3479-9/14/$31.00 \u00a92014 IEEE\nin [10], and in [11] eccentricity effects in a 12/8 SRM has been reduced by some modifications in stator windings. Although these rearrangements are cost effective and simple but they somehow just reduce the UMF and are not fully controllable and are limited to certain configurations.\nThe possibility of establishing some additional windings to apply radial force chiefly for bearingless types has been evaluated in various papers such as [12], but clearly it imposes extra expenditure and some adaptations in fabrication process.\nIn [l3] , it is shown that in a 12/8 SRM if beside energizing the conduction phase, two extra poles from the descending inductance phase be energized as well, it is possible to produce the desired RF. Using sinusoidal current excitation to control RFs presented in [14] and the idea of controlling currents in facing poles to mitigate UMF is just mentioned in [15].\nDoubly salient configuration of SRMs causes considerable torque ripple that necessitates the usage of control algorithms, and EFs exacerbate this ripple considerably [16].\nMany literatures introduced control strategies to minimize torque ripple in healthy SRMs like [17]. In [18] artificial intelligence like Neuro-Fuzzy controller is suggested to compensate manufacturing inaccuracies namely eccentricity. This strategy is simulated for a 6/4 SRM and minimizes the torque ripple; however a reduction in average torque is expected.\nAs it is evident little attention is dedicated to control UMF in eccentric condition. Minimizing torque ripple in faulty conditions seems inadequate.\nIn previous works, authors have assessed different EFs namely static, dynamic and mixed eccentricity and in [19] proposed a method to control UMF by rearrangements in the number of turns in each pole and through controlling phase currents overall torque is maintained suitably.\nIn this paper a novel approach is presented which through a new converter, EFs' effects namely UMF and torque ripple are compensated simultaneously. This control strategy is accomplished by controlling currents in facing poles such", + "that while the current of one pole is adjusted to keep UMF in a permissible range the other pole current is regulated somehow that minimizes torque ripple. Torque ripple control uses a double loop control algorithm that adjusts speed and torque.\nThe proposed converter requires the middle point of each motor winding in all phases be extracted and increases only two power switches namely one diode and one MOSFET for every phase comparing to conventional asymmetric bridge.\nThis paper is organized as follows: Section II briefly talks about problem statement of eccentricity compensation issue. In Section III, analytical model of the new compensation method is approved. In Section IV, the new practical strategy is introduced in details. Implementation on SRM, and FEM results and MATLAB/Simulink are completely analyzed in Section V. Finally, the main achievements of paper are concluded in Section VI.\nII. PROBLEM STATEMENT\nAir gap in a healthy SRM is uniform and comparing to other types is smaller but several parameters can change rotor centricity. Imprecise fabrication process can be the beginning of EFs and through bearings wear during the machine operation its performance will be marred considerably. Rotor misalignment also can be due to static friction which happens when rotor is in standstill position for a long time. EFs not only precipitate bearings deterioration but also exert a huge UMF on the rotor. Significant escalations on torque ripple and machine acoustic noise are other impacts of these faults.\nIn the presence of EFs in SRMs, flux distribution and electromagnetic fields in the air gap will be non-uniform and accordingly motor characteristics will alter considerably. Motor electric torque, stator currents, temperature, inductance profile, acoustic noise and vibration will change dramatically depending on the EF level. These deviations can also impair the efficacy of sensorless control algorithms for rotor position estimation. This unbalance disposition of fields makes flux density in front of facing poles unequal which leads to a tremendous UMF exerted on the rotor in radial direction.\nTwo main consequences of EFs that endanger the performance of motor, in short term and long term are torque ripple and UMF, respectively.\n1) It is highly expected that the motor produces a\nconstant electric torque and as SRMs naturally have\nexcessive torque ripple which can be aggravated by EFs so\nthat it is needed for a control algorithm to minimize torque\nripple as much as possible both in healthy and faulty\noperation to uphold interminable and satisfactory operation\nof the motor.\n2) As mentioned earlier EFs bring about UMF which\naccelerates bearing deterioration and in long term can cause\nthe rotor and stator poles abrade against each other which\nleads to irreparable damages, So that the control algorithm\nshould be in charge of mitigating UMF in order to prevent\nbearing decadence as well. In the following section using an analytical model for the SRM, it is shown how is feasible to control torque and mitigate UMF by adjusting the currents in facing poles.\nIII. THEORETICAL BASES\nThe main objective of this section is just to show the basic principal of the proposed method. Since in SRMs an analytical model which calculates UMF that is both detailed and general will be so extensive and complicated as is developed in [6], here a linear model for SRM is supposed that gives UMF for those rotor positions that rotor and stator poles overlap.\nAssuming that all the magnetic energy is stored in the air gap, authors in [19] have evaluated an analytical model for radial force and torque that each stator pole produces. Considering Fig. 1:\n(1)\n(2)\nWhere J.Lo is permeability of air , N is the number of turns in every stator pole, Ir is the length of rotor, r/) is the arc for overlapping region, i is the current magnitude in every stator coil and finally g is the radial air gap length in\nthe case of uniform air gap in healthy motor.\nSupposing a SE fault which means, rotor is off-center from stator symmetrical axes (Fig. 2) the degree of eccentricity will be given by:\nc=(:)XlOO% (3)\nwhere \ufffd is the displacement of the rotor.", + "As mentioned earlier Fr is the radial force applied by\nevery stator pole on rotor and given that stator opposite poles act independently, overall force exerted on rotor can be calculated by summing these two radial forces. In healthy operation of SRM the magnitude of these forces are equal and their directions are opposite, so the UMF will be zero but in faulty operation like EFs (Fig. 2) as the air gap varies for facing poles the huge radial forces will be disparate and accordingly UMF grows significantly. In eccentric condition considering Fig. 2 UMF is evaluated as follows:\nUMF = F,I + F,2\nfloN\ufffd (i,r, B) , floNi (i,r, B) 2 , II - 2 I, 2(g+L1) 2(g-L1) {g -L1 in front of Co ill\ng + L1 in front of Coil 2\n(4)\nSupposing that all poles have identical number of turns but current can vary, UMF will be:\n1 floN' (VA) UMF= 2 2 2 (g-L1) (g+L1)\nx ((g-L1)' ii' -(g+L1)'ii ) (5)\nGiven that the currents in facing pole can be controlled separately, in order to nullify UMF the currents ratio must be:\nUMF=O\n=> (g _L1)\\ 2 -(g +L1) 2 ii =0 (6)\nHere it is expected that the converter considering (2) controls the current of one pole somehow that torque is set at desired level and by controlling the other pole current such that the currents ratio satisfies (6), UMF become minimized.\nIV. NEW ALGORITHM\nIn accordance with what was mentioned in section II the control algorithm must maintain two vital roles. First is to\ncounterbalance UMF and second is to minimize TR as much as possible.\nFor such purposes a novel converter has been designed as it's shown in Fig. 3. This converter is planned for a three phase 6/4 SRM, but this idea is tenable for other configurations, too.\nConsidering Fig. 4 flowchart of the control algorithm is depicted. During motor operation certain characteristics of the machine must be realized. For example the current of every coil of the active phase, rotor position using whether sensors or position estimation methods, and radial forces exerted on rotor must be sampled.\nA. SJ Switching\nKnowing rotor position (), it is clear that based on a certain phase advancing or retarding which phase/phases must be excited. Once this is determined, Sf of the corresponding phase/phases is being closed in all the period that the active phase is supposed to be energized.\nB. S2 Switching\nS2 is controlled such that torque ripple is minimized. This idea is accomplished by an algorithm which has two control loops. The outer loop is to control rotor speed and the inner loop is responsible for decreasing torque ripple as is shown in Fig. 5. According to the scheme first actual rotor speed" + ] + }, + { + "image_filename": "designv6_24_0002572_978-3-319-76276-0_5-Figure13-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002572_978-3-319-76276-0_5-Figure13-1.png", + "caption": "Fig. 13 Fatigue zone for dynamic middle axle inner suspension spring deflection of 7, 5.27, 4.42 and 3.97 mm", + "texts": [ + " So it is felt necessary to study the effect of combination of both the modifications to increase fatigue life of inner suspension spring and due to this, the advantages of shim provision and modification of damping coefficient can be achieved effectively. From the calculation, it is observed that the modification of damping coefficient from 100,000 N s/m to 250,000 N s/m leads to shear stress of 640 N/mm2 and 590 N/mm2 without and with shim provision respectively as shown in Fig. 12. Hence, provision of shim along with damping coefficient of 150,000 N s/m reduces the shear stress to 598 N/mm2 which will definitely improve the fatigue life of spring and given in Table 3 and the contour plot for fatigue life is shown in Fig. 13. Thus the provision of shim along with modification of damping coefficient of damper is strongly advisable to avoid failures of middle axle primary inner suspension spring. iv. Design Modification of Primary Suspension Spring The fourth modification suggested, for improving the fatigue life of primary springs it is necessary to reduce the load shared by middle axle suspension springs by decreasing the stiffness of spring. This can be achieved by increasing the mean diameter of middle axle suspension spring" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000194_s12239-015-0047-9-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000194_s12239-015-0047-9-Figure1-1.png", + "caption": "Figure 1. Vehicle with transfer case and torque vectoring system.", + "texts": [ + " In order to get more traction under various driving conditions, four-wheel drive (4WD) is common driving system not only for a sport utility vehicle (SUV), but also for a normal sedan. The left and right torque vectoring system of the rear axle is a device that can control both the torque flow amount and its direction between the right and left wheels (Piyabongkarn et al., 2010; Sawase et al., 2006). This paper describes the vehicle dynamics improvement by means of transfer case and torque vectoring system that are shown in Figure 1. The focus of vehicle dynamics is on the agility and stability during handling situation. *Corresponding author. e-mail: hyungjin.kim@hyundai. com Figure 2 shows brief control logic structure for active yaw controller proposed in this paper. The control logic consists of signal process, actuator control logic, and vehicle control logic. This paper focuses on the vehicle control logic which includes estimator, reference model, yaw moment controller, and traction force distribution. In this paper, a simulation was conducted, so signal process was designed for the simulation signal" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002584_1999-01-0782-Figure14-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002584_1999-01-0782-Figure14-1.png", + "caption": "Figure 14. Radial Load Test Fixture", + "texts": [ + " Sample 2, after passing the 30,000 cycle mark exceeded the 10% deflection limit at 74,284 cycles and the test was concluded. On closer examination, 1 crack was found in the spoke weld at the hub and 7 cracks were found in spoke-rim welds. In addition to the dynamic cornering test, a radial fatigue test was carried out on one of the prototypes. The radial fatigue test involves mounting a wheel to a hub attached to the end of a hydraulic cylinder and loading the wheel radially into a large spinning drum as shown in Fig. 14. The radial load to which the wheel was subjected was determined by the SAE formula R=L*K (1) Where, R = Radial Load (N) L = Load rating of the wheel (N) K = Accelerated test factor Results of the radial fatigue test are shown below DESIGN CRITIQUE \u2013 The decision too develop a third version of the wheel resulted from a review of the state of the current design. Positive aspects of the wheel were: 1. The wheel met cornering fatigue test requirements for the target application. 2. The wheel met radial fatigue requirements of the target application", + " Early in this investigation, we concluded that by far, the most desirable situation would be to devise a wheel center which would be made using existing center stamping equipment and to the extent possible, the presently used methods. Similarly, the most advantageous material appeared to be common, low carbon sheet steel as is presently used for most low cost, steel wheels. With this set of targets, we set about finding a way to implement this twin spoke design as a single piece, stamped wheel center. Results of this effort are shown in Fig. 14. The stamping starts with a shaped blanking cut. Spoke pairs as well as most of the hub area details are then formed in stages with the spoke made by wiping the blank edges toward what will ultimately be the outside of the wheel. Finally, the material between the spokes of each pair is removed and discarded. MATERIAL CONSIDERATIONS \u2013 The complexity of this design made further traditional analysis problematic at best, so our design iterations were carried out using intuition and finite element analysis" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000810_0470871199.ch11-Figure11.8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000810_0470871199.ch11-Figure11.8-1.png", + "caption": "Fig. 11.8 Near-field antenna radiation pattern measurement range. a. Planar near-field range. b. Cylindrical near-field range.", + "texts": [ + " Therefore, a planar near-field antenna measurement range is best suited for highly directive antennas where most of the radiated energy will be directed into the forward direction. The spherical nearfield antenna measurement range is best suited for near-omnidirectional radiators. The cylindrical near-field antenna measurement range forms a good alternative for measuring low-directivity antennas. Besides, the cylindrical near-field antenna range is easily combined with a planar near-field antenna range as is shown in figure 11.8. Figure 11.8a shows a planar near-field antenna measurement range. The scanner moves the probe around to take samples of the near-field in amplitude and phase over a regular grid. This scanner is shown here without its lining of RF absorbers (RAM). The probe is a small antenna (often taking the form of an open-ended waveguide), that is kept as small as possible to minimise reflections between the AUT and probe. As figure 11.8b shows, the planar scanner when operated as a vertical linear scanner, combined with the AUT being rotated on its pedestal, allow - in a relatively easy and cost-effective way - the operation of the planar near-field range as a cylindrical near-field range. Although the (planar) near-field antenna measurement range offers the opportunity to obtain the radiation pattern of a large phased array antenna in a limited-size indoor environment, its value within the design path of a phased array antenna lies mostly in the final, verification stage of the complete antenna with feeding network and phase shifters" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003166_cicc.2002.1012887-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003166_cicc.2002.1012887-Figure5-1.png", + "caption": "Fig. 5. (a) Thecenter contact is completely surrounded by another contact. (b) The model agrees with simulations in all cases. The plot is shown for xa = 40pm.", + "texts": [ + " Hence, only one additional parameter is required to model the 2 1 2 dependence on the relative contact position. The coefficients a, b and c are scalable with contact dimensions. Therefore once the parameters are extracted for a specific geometry, they can be scaled for different contact geometries. (5 ) where p = a, or, b or, c and ct and anew are equal to the 211 value of the merged contacts before and after scaling, respectively. This model can now be used to evaluate the substrate resistance when one contact is surrounded by another contact as shown in Fig. 5 (a). Here, the number of interacting (coupling) sides is four. In such cases, the first step in the 2 1 2 calculation is to divide the complex contact shape into rectangular contacts. Then the coupling between each rectangular part and the contact that they surround are calculated separately. Finally, the overall 2 1 2 between the two contacts is a superposition of the calculated 2 1 2 values. Fig. 5 (b) illustrates the model versus simulation results for the example of Fig. 5 (a). a n e w pnew = pold x -a IV. EXPERIMENTAL VERIFICATION The 2-parameter formulation model proposed in the previous sections has been validated on a test chip. This test chip was fabricated in a 0.35pm CMOS TSMC process through MOSIS. The chip has several substrate test structures as shown in Fig. 6. The test structures for validation of the low frequency scalable model have several p+ contacts of different sizes and shapes. All the contacts are connected to 60pm x 6Opm DC probe pads for probing" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000673_tie.2018.2842736-Figure9-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000673_tie.2018.2842736-Figure9-1.png", + "caption": "Fig. 9 Prototype and parts of MFM-BDRM", + "texts": [ + " Therefore, the integer-slot winding is the optimal winding configuration for MFM-BDRM. It is noted that the FSCW analyzed in the MFM-BDRM is standard dual-layer FSCW. With the design method of FSCW developed, the unfavorable effect of FSCW on the MFM-BDRM may be mitigated or solved [25]. To verify the validity of the above analysis in Part I and Part II, a prototype of MFM-BDRM with the integer-slot winding and optimal pole-pair combination of stator, PM rotor and magnetic blocks is designed and manufactured, as shown in Fig. 9. The structure of PM rotor and stator is similar to that of tradition PM machine. The PM rotor is a surface-mounted PM rotor, which is bound by a layer of epoxy material to ensure its safety under high rotating speed. The stator winding employs the integer-slot winding configuration. The difficulty in the manufacture of MFM-BDRM is the modulating ring rotor. To reduce harmonic loss, the magnetic block is manufactured by laminating silicon steel sheets along peripheral direction. To keep enough mechanical strength of modulating ring rotor, a special material (zirconium dioxide) of magnetic block holder is designed and manufactured" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002884_s11082-015-0225-z-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002884_s11082-015-0225-z-Figure1-1.png", + "caption": "Fig. 1 Staring, pushbroom and whiskbroom imaging", + "texts": [ + " Strategic sensors are used for reconnaissance and surveillance and are fielded on high altitude surveillance aircraft, medium and high altitude endurance UAVs, and satellites (Driggers et al. 1996). For tactical sensors, which include electro-optical (EO) and infrared (IR) systems, the traditional system level performance measures are the probabilities of detection, recognition, and identification. Depending on how the sensor acquires and records the incoming signal, imaging systems can be divided into three general categories: snapshot, whiskbroom, and pushbroom imagers (Fig. 1) (Perschy 2000). The method employed in the pushbroom is a line by line scanning technique, while in whiskbroom method image scanning is carried out point by point. Snap-shot acts as conventional imaging cameras; it maps and stores the entire image on a CCD matrix array at once. In this method, imaging was done within short period, therefore this method is less sensitive to satellite instability than other techniques. In other words, this snapshot can be implemented with low cost and complexity stabilizer system" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002440_j.sna.2016.01.048-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002440_j.sna.2016.01.048-Figure6-1.png", + "caption": "Fig. 6. Experimental setup for measuring a focal length of a varifocal liquid lens.", + "texts": [ + " the applied voltage was changed from 0 V to 50 V at 5 V increments, as shown in Fig. 4. The average variation of sag height per volt is about 32 m. Fig. 5 shows images captured through the liquid lens with different sag heights. The captured image was demagnified for the negative sag height and magnified for the positive one. The focal length was also measured with respect to an applied voltage by using a custom-built testing system consisting of a laser, mirror, liquid lens, and detection screen, as shown in Fig. 6. The laser light of wavelength 532 nm (PB256-C, Optoelectronics tech. Co., Ltd.) was transmitted through the liquid lens. The focal length of the liquid lens depends on the sag height of the lens. To find the appropriate focal length, the detection screen mounted on a 1D traverse system was constantly transported until the minimum size of the laser spot was found. The liquid lens had a concave surface profile with a negative focal length (\u221210 mm) at the initial stage. When the applied voltage was increased, the ring-type neodymium magnet pushed the membrane down" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003402_cac51589.2020.9326821-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003402_cac51589.2020.9326821-Figure1-1.png", + "caption": "Fig. 1. Manipulator structure", + "texts": [], + "surrounding_texts": [ + "3086978-1-7281-7687-1/20/$31.00 2020 IEEE\nnonlinearity and strong coupling, which is easily affected by\nexternal factors, making the system instability. Based on the\nabove problems, an adaptive fuzzy sliding mode control method\nis designed to study the trajectory tracking problem of\nmanipulator in the paper. Firstly, a two-degree-of-freedom\ndynamic model of the manipulator is established according to\nthe Lagrangian equation, and then an adaptive fuzzy sliding\nmode control algorithm for the manipulator is proposed and the\nLyapunov function is used to prove the convergence and\nstability of the algorithm. The simulation results show that the\nadaptive fuzzy sliding mode control can make each joint of the\nmanipulator respond to the desired trajectory quickly, and the\ntracking error converges to zero within a certain time, which\nhas achieved excellent control effect and verified the\neffectiveness of the algorithm.\nKeywords Manipulator, Adaptive control, Fuzzy control,\nSliding mode control\nI. INTRODUCTION\nManipulator is a complex system with strong coupling, which is nonlinear and highly time-varying [1]. In addition, there are external disturbances, parameter uncertainties and other factors in the system, which are easily disturbed by the environment and lead to system instability [2-3]. Therefore, it is necessary to design a controller with better performance to control it.\nUsually, the manipulator system is transformed into a two-degree-of-freedom linkage mechanism. To solve this problem, many scholars at home and abroad have studied and designed a variety of control methods. Common manipulator control methods include PID control, adaptive control, fuzzy control, sliding mode control, robust control and so on, but the control effects of these methods are not satisfactory. Literature [4] uses PID control method, which is simple, easy to implement and does not require model parameters, but its robustness is not good enough. Literature [5] puts forward a new design method of robot controller with parameter uncertainty, which adopts adaptive control scheme. Under the controller, the index error converges to zero, and the designed system can effectively reduce the influence of system parameter uncertainty. Document [6] proposes a robot trajectory tracking method based on fuzzy robust control, which combines the advantages of traditional robust control and fuzzy control. Simulation results show that the tracking error of the system is small and the convergence speed is fast. Literature [7-9] applies neural network theory for manipulator control. The simulation results show that the system can track the reference signal quickly with excellent accuracy and robustness. Global stability of the proposed neural network controller is proved theoretically. In reference[10-13], aiming at the inherent problems in sliding mode variable structure control of manipulator, saturation function is used instead of sign function to solve the problem, thereby improving exponential approximation law. The simulation results show that the controller can reduce the chattering phenomenon of 20 20 C hi ne se A ut om at io n C on gr es s ( C A C ) | 9 78 -1 -7 28 1- 76 87 -1 /2 0/ $3 1. 00 \u00a9 20 20 IE EE | D O I: 10 .1\nAuthorized licensed use limited to: California State University Fresno. Downloaded on June 22,2021 at 02:50:53 UTC from IEEE Xplore. Restrictions apply.", + "3087\nthe system, track the ideal trajectory more accurately, and realize the characteristic of convergence in finite time.\nIn view of the problems of the above-mentioned control methods, combined with the highly nonlinear and strong coupling characteristics of manipulator, adaptive fuzzy control combined with sliding mode control is proposed to design a robust control system which can slow down chattering phenomenon, and the convergence and stability of the system are proved by Lyapunov function in the paper. Through the simulation verification of the designed adaptive fuzzy sliding mode controller, it is proved that the proposed method can control the two joints of manipulator to converge to the ideal position in a limited time.\nII. DYNAMIC MODEL OF MANIPULATOR\nFor multi-joint robot, its dynamic model can be established according to Lagrange equation [14], which can be expressed by the following second-order nonlinear differential equation:\n(1)\nIn which is joint angular displacement, is positive definite mass inertia matrix,\nis coriolis force and centrifugal force, is gravity and is control torque.\nIn the practical application of manipulator, and cannot be accurately obtained due to the existence of measurement error and external environmental noise. Therefore, and are divided into two parts: nominal value and disturbance value. The nominal value is the known prior estimated value, while the disturbance value is the uncertain part, as shown in the following formula:\n(2)\nIII. DESIGN OF MANIPULATOR CONTROLLER\nAccording to the above discussion, the structure of the adaptive fuzzy sliding mode controller designed in the paper is shown in Fig. 2. The sliding mode surface is designed according to the obtained difference between the ideal position and the actual position of the manipulator. The uncertainty can be obtained by a series of derivation, and then the fuzzy universal approximation theorem can be applied to approximate . The adaptive law adjusts the ideal position according to the actual position . The specific derivation process is as follows:\nAccording to the established manipulator dynamics model, is defined as the input command (ideal joint\nposition), then error signal is as follows:\n(3)\nThen the derivative and the second derivative error are as follows,\n(4)\nAccording to the principle of sliding mode control, the sliding mode surface can be designed as:\nThe derivative of is as follows:\n(5)\nFor the error system, the following exponential reaching law is designed:\n(6)\nAuthorized licensed use limited to: California State University Fresno. Downloaded on June 22,2021 at 02:50:53 UTC from IEEE Xplore. Restrictions apply.", + "3088\nIn which .\nAccording to the sliding mode surface and reaching law, the control law of the system can be designed as follows:\n(7)\nSubstituting the designed control law into formula (1) can be obtained:\n(8)\nReplace as follows:\n(9)\nSo the formula (8) becomes\n(10)\nSimplified formula (10) can be obtained\n(11)\nSince the disturbance value is uncertain, another quantity can be implemented instead. In addition to the fitting characteristics of multivariate functions, fuzzy systems also have universal approximation characteristics, so can be approximated by the output of fuzzy systems. The specific steps are as follows:\nStep 1: for the variable , define fuzzy sets\nStep 2: According to the above discussion and the definition of input-output membership function, the rule base can be determined as follows:\nThe membership function used to represent fuzzy sets is designed as follows:\n(12)\nAccording to fuzzy mathematics theory, the output of fuzzy system is:\n(13)\nis approximated by fuzzy control, and is the ideal approximation of disturbance value . According to the universal approximation theorem, there is , which satisfies the following inequality\n(14)\nIV. CONVERGENCE AND STABILITY ANALYSIS OF THE SYSTEM\nThen, to prove the stability of the system, Lyapunov function can be selected as follows:\n(15)\nIn which\n(16)\nDue to ,\nAuthorized licensed use limited to: California State University Fresno. Downloaded on June 22,2021 at 02:50:53 UTC from IEEE Xplore. Restrictions apply." + ] + }, + { + "image_filename": "designv6_24_0002102_iecon43393.2020.9255013-Figure14-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002102_iecon43393.2020.9255013-Figure14-1.png", + "caption": "Fig. 14. Assembly of PMs for realizing single-helix translator, a schematic drawing.", + "texts": [], + "surrounding_texts": [ + "[1] G. Mork, S. Barstow, A. Kabuth, M. T. Pontes, Assessing the Global Wave Energy Potential. ASME 2010 29th International Conference on Ocean, Offshore and Arctic Engineering. pp. 447-454, 2010. [2] J. Wang, K. Atallah, W. Wang, Analysis of a magnetic screw for high force density linear electromagnetic actuators. IEEE Trans. Magn. 47(10), pp. 4477-4480, 2011. [3] S. Pakdelian, N. W. Frank, H.A. Toliyat, Magnetic design aspects of the trans-rotary magnetic gear. IEEE Trans. Energy Convers. 30(1), 41- 50, 2014. [4] H. A. Hussain, A Novel Contactless Rotary-to-Linear Magnetic Actuator. 2019 IEEE International Electric Machines & Drives Conference (IEMDC). pp. 1081-1086, IEEE, 2019. [5] N. Ils, A.B. Christiansen, R.K. Holm, P.O. Rasmussen, Design and test of a reluctance based magnetic lead screw PTO system for a wave energy converter. 2017 IEEE International Electric Machines and Drives Conference (IEMDC), pp. 1-8, IEEE, 2017. [6] A. Heya, Y. Nakata, M. Sakai, H. Ishiguro, K. Hirata, Force Estimation Method for a Magnetic Lead-Screw-Driven Linear Actuator. IEEE Trans. Magn. 54(11), pp. 1-5, 2018. [7] Gao, F., Wang, Q., Jibin, Z. Analytical Modeling of 3-D Magnetic Field and Performance in Magnetic Lead Screws Accounting for Magnetization Pattern. IEEE Trans. Ind. Electron., 67, pp. 4785-4796, 2019. [8] Lu, K., Xia, Y., Wu, W., Zhang, L. New helical-shape magnetic pole design for magnetic lead screw enabling structure simplification. IEEE Trans. Magn., 51(11), pp. 1-4, 2015. [9] Gao, F., Wang, Q., Hu, Y., Chen, B., Zhao, B. Performance Evaluation of Magnetic Lead Screws Equipped With Skewed Arc Magnets Instead of Helical Ones. IEEE Trans. Magn., 54(11), pp. 1-5, 2018. [10] Ling, Z., Zhao, W., Ji, J., Liu, G. Design of a new magnetic screw with discretized PMs. IEEE Trans. Appl. Supercond., 26(4), pp.1-5, 2016. [11] Abolhasani, A., Pakdelian, S. Equivalent circuit for the trans-rotary magnetic gear. IEEE Trans. Ind. Electron., 66(10), pp. 8266-8272, 2018. [12] Liu, X.; Liu, Y.; Li, X. Parametric Analysis and Design of Magnetic Lead Screw. 2019 IEEE International Electric Machines & Drives Conference (IEMDC), pp. 1990-1995, IEEE, 2019. [13] Cirolini, M., Flores Filho, A. F., Wu, Y. C., Dorrell, D. G. Design Aspects of a Reluctance-Based Magnetic Lead Screw. IEEE Trans. Magn., 55(7), pp. 1-6, 2019. [14] Holm, R. K., Berg, N. I., Walkusch, M., Rasmussen, P. O., Hansen, R. H. Design of a magnetic lead screw for wave energy conversion. IEEE Trans. Ind. Appl., 49(6), pp. 2699-2708, 2013. [15] Ling, Z., Ji, J., Wang, J., Zhao, W. Design optimization and test of a radially magnetized magnetic screw with discretized PMs. IEEE Trans. Ind. Electron., 65(9), pp. 7536-7547, 2017. [16] Zhao, A., Wu, W., Zhu, L., Chen, H., Lu, K., Blaabjerg, F. Design and Experiment of an Indirect Wave Power Generation Device using Magnetic Lead Screw. IECON 2019-45th Annual Conference of the IEEE Industrial Electronics Society., pp. 6987-6991, IEEE, 2019. 2812 Authorized licensed use limited to: Sungkyunkwan University. Downloaded on December 18,2020 at 16:25:50 UTC from IEEE Xplore. Restrictions apply." + ] + }, + { + "image_filename": "designv6_24_0003257_s11771-015-2507-9-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003257_s11771-015-2507-9-Figure8-1.png", + "caption": "Fig. 8 Position output (set-point hovering flight)", + "texts": [], + "surrounding_texts": [ + "The proposed control strategy has been tested by numerical simulation in order to check the control performance attained for set-point hovering flight and path following flight of a quadrotor helicopter subjected to external wind disturbance. Due to the system uncertainties such as initial state error, attitude error and noise in the actual situation, a Gaussian white noise module is added to all feedback variables. The turbulent wind field fragment is generated by simulation based on the model establish in section 3. Let simulation time be 60 s and the wind velocity be 20, feet altitude be 10 m/s, the flight altitude of quadrotor be 1 m. The parameters of the turbulent wind filed are summarized in Table 1 and the generated turbulent wind filed fragment is shown in Fig. 6. Table 2 lists the dynamic parameter of the quadrotor helicopter." + ] + }, + { + "image_filename": "designv6_24_0003233_cp.2014.0319-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003233_cp.2014.0319-Figure7-1.png", + "caption": "Fig. 7: Equipotential distributions for different Hc of LCF magnets (flux enhanced, LCF magnet Br=1.1T, r=3), (a)40kA/m, (b) -80kA/m, (c) -120kA/m.", + "texts": [ + " By changing Hc of LCF magnets, the variation of open-circuit back EMFs can be seen in Fig. 6. For flux weakened state, the figure shows working points of LCF magnet are quite stable no matter what Hc is except when Hc is extremely low. However, in flux enhanced state, it is obvious that back EMF begins to reduce when Hc is still high. Unless the Hc of LCF magnet is high enough to withstand the strong magnetic field of NdFeB magnet, the open-circuit cross coupling will always exist. These influences can be seen clearly from equipotential distributions as shown in Fig. 7. When Hc is small, the working points of LCF magnet are low, and nearly no flux generated by LCF magnet goes into the airgap. Therefore, in designing the HMMM, Hc of LCF magnet should not be chosen too small in order to reduce the open-circuit cross coupling effect. The armature reaction makes the on-load cross coupling effect more complicated than open-circuit conditions. The magnetic circuit model of HMMM with Id=0 control is shown in Fig. 8, where \u039bq and Farmq refer to permeance of q-axis flux loop and q-axis armature magneto-motive force" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000885_j.apm.2014.04.032-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000885_j.apm.2014.04.032-Figure4-1.png", + "caption": "Fig. 4. The positioning of the suspension control arms: arm axis (a), revolute angle (b).", + "texts": [ + " The driving elements, which correspond to the active degrees of mobility, are defined by the kinematic parameters u5 (coming from the steering wheel rotation) and la (coming from the vertical motion of the wheel). The mobility from the damper travel la is a \u2018\u2018total\u2019\u2019 mobility \u2013 with influence on the whole mechanical system, while the mobility from the pitman arm rotation u5 is a \u2018\u2018partial\u2019\u2019 mobility \u2013 driving only the steering tie-rod and the wheel carrier. In the guiding system, the following kinematic functions can be defined (see Fig. 1): u1(la) cite this article in press a s of mobility, Appl. Math the rotation of the lower control arm (Fig. 4) s: P. Alexandru et al., Modeling the angular capability of the ball joints in a complex mechanism wit . Modell. (2014), http://dx.doi.org/10.1016/j.apm.2014.04.032 u3(la) the rotation of the upper control arm (Fig. 4) a(la) the variation of the kingpin inclination angle of the steer axis (Fig. 5a) b(la) the variation of the caster angle of the steer axis (Fig. 5a) zk(la) the variation of the vertical displacement of the wheel (9) DEk the wheel track variation h(u5) the steering angle of the wheel (Fig. 5b) hb(la) the steering motion induced by the suspension displacement bpr(la,u5) the pressure angle in joint (Fig. 6) Concerning the motion functions u1,3(la) of the suspension arms 1 and 3 \u2013 revolute motion around the local axis x1,3, which are defined by the global coordinates of the points M0 0 M00 0 and N00 N000, the orientation of the motion axis x1,3 is defined by the pair of angles w01\u2013t01, respectively w03\u2013t03, relative to the longitudinal axis x of the car body reference frame (Fig. 4a): tgw01;3 \u00bc yM00 ;N 0 0 yM000 ;N 00 0 xM00 ;N 0 0 xM000 ;N 00 0 ; tgt01;3 \u00bc zM00 ;N 0 0 zM000 ;N 00 0 xM00 ;N 0 0 xM000 ;N 00 0 : \u00f010\u00de The passing from the fixed reference frame xyz (attached to the car body) to the local reference frame x1,3y1,3z1,3 (attached to the suspension arm 1/3) is performed by two rotations w01,3/t01,3. In Fig. 4a, there is also represented the angle t001;3 = arctg (tg t01,3 cosw01,3), corresponding to the rotation around y1,3. The local axis y1,3 is positioned with the angle w01,3 relative to h two \u03c8 z1,3 0 z y y1,3 M\"; N\"0 0 x x1,3 M' ; N'0 0 1,3 \u03c801,3 \u03c501,3 \u03c501,3 M ; N0 0 y1,3 z1,3 \u03d51,3 M; N 2 1;3 a b ' ' \u03c501,3 Fig. 5. The position angles of the steer axis: steer axis angles (a), steering knuckle angle (b). N 2 A 2' CF G M 4 (CMN) vc F4 \u03b2pr Fig. 6. The pressure angle in the ball joint C. the transversal axis y, in the horizontal plane (xy). Consequently, the current rotation angle u1, respectively u3, will be positioned relative to the local axis y1, respectively y3 (Fig. 4b). Practically, it is important to evaluate the numeric values of the functions u1,3(la) in the wheel travel limits (\u2018\u2018 + \u2019\u2019 bump travel, \u2018\u2018 \u2019\u2019 rebound travel). The modeling of the position angles a(la) and b(la) of the steer axis MN is made relative to the vertical axis z of the fixed reference frame xyz (Fig. 5.a). The steer axis MN is not normal on the ground plane (xy), being titled with the angles a (kingpin inclination angle) \u2013 around x, and b (caster angle) \u2013 around y, which are used to realize the passing in the local frame xp2 yp2 zp2 (defined on the wheel carrier)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003233_cp.2014.0319-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003233_cp.2014.0319-Figure3-1.png", + "caption": "Fig. 3: Open-circuit equipotential distributions, (a) flux enhanced, (b) flux weakened.", + "texts": [ + " Basic geometric parameters of the analysis model are shown in Table 1. Fig. 2 shows the demagnetization curves of two permanent magnets which are the same as the original paper about HMMM [9]. It can be seen that much smaller coercive force (Hc) of LCF magnet compared with NdFeB magnet makes the online magnetization by armature coil possible. When LCF magnets are magnetized to enhance the NdFeB magnet field, essential flux focusing exists and the opencircuit equipotential distributions can be seen in Fig. 3 (a). On the contrary, when LCF magnets are magnetized to weaken NdFeB magnet field, majority of flux will short circuit within the rotor as can be seen in Fig. 3 (b). The simplified magnetic circuit models of HMMM for opencircuit condition are shown in Fig. 4. Fnfb and Flcf stand for magneto-motive forces of NdFeB and LCF magnets, while \u039bnfb and \u039blcf stands for the permeances of NdFeB and LCF magnet flux loops in stator. \u039bg and \u039br stand for the permeance of airgap and rotor, which will be shared by both magnets. 2 In flux weakened state, two types of magnets are shortcircuited by each other. No matter what the Hc of LCF magnet is, the strong circulating flux dominated by NdFeB magnet will make working point of LCF magnet stable" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001717_j.biosystemseng.2008.02.010-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001717_j.biosystemseng.2008.02.010-Figure4-1.png", + "caption": "Fig. 4 \u2013 Schematic illustration of the robotic arm for extracting pomegranate arils.", + "texts": [ + " In these tests the nozzle diameter, air pressure, and number of passes of the nozzle were set at 3.5 mm, 700 kPa, and 4 passes, respectively. In this study, the diameters of the nozzles were the same as those used by Khazaei et al. (2005) and Sarig et al. (1985) for extracting citrus juice and juice sacs and pomegranate arils, respectively. Previous studies have used mechanical methods to move and control the route of the nozzle over the surface of fruits (Grosz and Sarig, 1985; Nahir and Ronen, 1992; Sarig et al., 1985). However, in this study a robotic arm was used for this purpose (Fig. 4). An air nozzle was attached to the end of a robotic arm positioned by a hardware system and controlled by a PC. The complete system consisted of a robotic arm and a nozzle, a regulated air compressor with a maximum supply pressure of 800 kPa to supply air to the nozzle, and a fruit holder. The microprocessor-based hardware system was controlled via the PC using AutoCAD 2004 software. For each test, the route of the nozzle was drawn using the AutoCAD software. The robotic arm moved the nozzle over the surface of fruit along the drawn path with a predetermined number of passes. The robotic arm was equipped with two stepper motors which made it possible for the nozzle to travel along the desired routes (Fig. 4). Using this method the route of the nozzle based on the fruit size and shape and the number of passes of the nozzle could be easily changed. For each test, the weight of the fruit half was determined to 70.1 g using a digital balance. Following each test, all the remaining arils inside the fruit half were extracted manually. The weight of the fruit rind, internal partitions, and non-extracted arils were determined using the same balance and the percentage of extracted arils was calculated. In all the experiments, the perpendicular distance of the nozzle outlet from the fruit surface was fixed at 10 mm" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001223_1.c032150-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001223_1.c032150-Figure1-1.png", + "caption": "Fig. 1 Three separate meshes used for the viscous drag computation. Representative strips are shown for the wing and tail grids.", + "texts": [ + " The form factor accounts for the higher flow velocities relative to the flat plate and corrects for the additional pressure drag due to the boundary-layer displacement effect. For fuselage-like bodies, the form-factor correction is Kform 1 1.5 d l 1.5 50 d l 3 (2) where d\u2215l is the ratio of diameter to length. The contribution of a given component to the drag coefficient is then CD KformCf Awet Aref (3) where Awet is the wetted surface area, and Aref is the reference wing area. In our implementation, we use the three-dimensional surface geometries shown in Fig. 1, which allows us to use the same geometric design variables as in the aerodynamic and structural disciplines. The viscous drag for the wing and tail are computed in a stripwise fashion that accounts for local changes in the Reynolds number due to chord modifications and thickness-to-chord changes due to shape changes. Smooth Kreisselmeier\u2013Steinhauser (KS) functions [50] are used to estimate the t\u2215c ratio from the discrete surface data to ensure smooth, continuous derivatives. As discussed in the introduction, the efficient sensitivity analysis of high-fidelity aerostructural systems is a significant challenge" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001470_8.1133-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001470_8.1133-Figure2-1.png", + "caption": "Fig. 2. Coaxially fed monopole in infinite parallel-plate waveguide.", + "texts": [ + " The probe current distribution is determined by the requirement that the total axial field component E&, z) generated by all the probes and the 0018-926X/88/0400-0449$01 .OO 0 1988 IEEE 450 IEEE TRANSACTIONS ON ANTENNAS AND PROPAGATION, VOL. 36, NO. 4, APRIL 1988 parallel plate waveguide. (b) Top and side view of linear array model. magnetic frills vanishes on the reference probe surface. The analysis is carried out in a number of steps. 1) Determine E&, z ) due to the combination of a single probe current Ja(z) and the magnetic frill M@ concentric to it, shown in Fig. 2, where and 6(z) , VO Mob, Z ) = ~ O & O = -40- b p In - 12 < P < b U (1b) with 6 representing the Dirac impulse function. As indicated, Ezo(P, z)=E;o(P, 2 ; vo)+E;o(P, 2 ; Jzo) . ( 2 ) An expression for Ea@, z ) was obtained in [8] in terms of parallel-plate waveguide radial modes (with propagation constants K,) and is reproduced as follows: with n?r h C,(z)=cos - z dn = JO (Kn b 1 - JO (Kn a) (30 Xn=Hf)(Knb) - H f ) ( K n U ) . (3g) Here Jo and Hi2) denote the Bessel function of the first kind and the Hankel function of the second kind, of zero order" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002921_iecon.2011.6119447-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002921_iecon.2011.6119447-Figure3-1.png", + "caption": "Fig. 3. Rotor temperature monitoring instrumentation: (a) block diagram of the assembly parts (b) photos of the instrumentation enclosure.", + "texts": [ + " The following design aspects contribute significantly to a satisfactory overall power design: The use of thermocouples Low power microcontroller Circuit design with fewest possible number of electronic components The use of IrDA in the optical data link The rotating circuit is powered by a single PP3 9V battery. The overall current load is less than 33 mA, hence with a lithium power cell of 1200 mA, an effective continuous operation of more than 30 hours can be guaranteed. The battery is exposed to very high centrifugal forces. These have not led to any notable performance degradation so far. Fig. 3(a) shows a block diagram of the mechanical assembly. There are two printed circuit boards (PCB) with diameter of 60 mm, one for the thermocouple conditioners and the other for the microcontroller and the power electronic. The boards\u2019 dimensions were minimized to reduce the influence of the centrifugal forces as much as possible. Allocating the electronic components on two boards is necessary in order to have sufficient space for all electric components by keeping small the overall diameter of the instrumentation. The boards are fixated into two aluminum bodies separated by a plastic battery container. The assembly is shown in Fig. 3(b). In case of monitoring machines with different diameter of the shaft, we have to manufacture only the back aluminum body, the one with the thermocouple conditioners shown in the left picture of Fig. 3(b). This particular method of mounting requires a hole of at least 5 mm in the shaft that stops at least to the near end of the rotor. This leads the thermocouples from the rotor through to the instrumentation. From thermal point of view, the enclosure is completely mounted outside the motor. It has to be noted here that to avoid eccentricities, the whole enclosure should be manufactured very precisely and balanced in order to withstand the desired maximum speed of 6000 rpm. Machines operated by PWM driven inverters present due to the current switching an extreme electromagnetic hostile environment for rotor temperature measuring systems" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000881_pesc.2008.4592618-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000881_pesc.2008.4592618-Figure2-1.png", + "caption": "Figure 2. Co-integration of vertical and lateral MOS and bipolar devices (view at the surface of the die).", + "texts": [], + "surrounding_texts": [ + "I. INTRODUCTION\nThe context of this work is the monolithic and functional integration within the power switch. Many works are under going in these research and development areas. We can refer to the work carried out in the industry of course with references [1,2,3] but also by the academic community with references [4, 5, 6, 7, 8, 10]. Most of this work can be split in two other categories. This first one tries to address monolithic functional integration with extensive help of modern and new technological processes [1, 2, 4, 9, 11]. On the other way, the second category tries to develop new concepts and functional integration at limited extra cost leading often to reduce use to technological extra steps [6, 7,10]. Both approaches are of interest and lead to complementary results, characteristics and performances. This work considers the second approach as a lead. The approach considers the opportunity to integrate devices and functions all isolated from the power device thanks to reverse biased PN junctions. This approach is not new and has already been considered in the past [19]. Its full development has not been engaged yet by\nindustrials but we believe that there is here an opportunity for the development of low-cost \"intelligent\" vertical power devices, generic and simple to implement. In such way, some lateral devices can be realized through the power device technological process and we are investigating what can be made thanks to these devices and what can be the expected performances of the resulting integrated functions. All electronic environments is of interest in this integration effort [15, 13, 12] but we are going to focus here on the co integration of a power device and a lateral one, both operating at logic level and both sharing the same technological process. All the work is based on the classical double diffused technological process of a 600V power VDMOS.\nThe first part of the paper is going to justify the modeling and analysis work that is carried out in this paper. Especially, it is shown how it may be possible to integrate the gate driver in the power device based on minimum technological process modifications and optimal operating performances. In the second part of the paper, we will introduce the objectives and the concepts that have been developed in order to co integrate both devices at best operating performances. The work will be based on numerical analysis and analytical well known equations. A special care will be paid on static characteristics of both devices. This analysis will allow us to conclude on the opportunity to integrate, at minimum cost, a gate driver element fully functional under any operating conditions. If dynamic and susceptibility considerations will have to be investigated later, this part will allow to conclude on operational characteristics\n978-1-4244-1668-4/08/$25.00 \u00a92008 IEEE", + "The third part of the paper will focus on practical validation of the work. Devices have been realized through a power device technological process and they have been characterized in both static and dynamic modes. Based on these results, an equivalent multilevel model has been developed in VHDL-AMS, allowing to perform time domain simulations of \"complex\" structures. An estimation of possible gate driver characteristics as well as its corresponding surface within a power switch will be provided.\nAs it has been underlined in several works [19], the monolithic integration of lateral devices is possible when based on the power device technological process. For this, both the power vertical device and signal lateral devices must exhibit satisfactory operating ranges and characteristics while sharing the same process. Besides, the electrical compatibility must be guaranteed at any power device operating conditions. This means that the signal devices, which are integrated within the substrate of the power VDMOS, must be correctly isolated from it. The first issue is deeply addressed in this paper, trying to show how must be set the compromises among the devices. Before this part, we must recall and underline the context and the opportunity that are available in our approach. First of all, the gate driver part must be integrated next to the power cells, within the surface voltage edge terminations where power potential are reduced. Even if this seems logical, it introduces restrictions since it is necessary to keep identical power device voltage capabilities. As it has been mentioned above, in order to minimize technology and to comply with the last sentence, the isolation technique considered is based on a reverse biased PN junction. It is the channel region of the vertical power device which is usually considered for this task. An acceptable electrical isolation is obtained no matter what are the operation polarizations of the devices as long as the P type channel regions are maintained to the reference potential of the silicon die, i.e. the source potential of the VDMOS. In such way, the vertical structure is kept unchanged and the power device voltage ratings are guaranteed. However, only a few signal components can be integrated next to the power MOSFET device, within the P type channel regions. The picture below presents a cross section of a power device and, next to it, are presented possible signal devices correctly isolated no matter what are the operating conditions of the power device. The signal components that can be integrated are: -a lateral N-MOS with substrate referenced at power device source potential, -a signal diode with its anode referenced at power device source potential, -polysilicon and N+ type resistors (for N+ type resistors, they must be realized in P type regions which must be referenced at power device source potential).\nOther lateral devices can be considered [16] but their electrical isolation and compatibility can not be guaranteed at any VDMOS operating conditions. Based on these possible devices, we have developed simple gate driver topologies based on the work carried out many years ago on N-MOS technologies [18]. The following part of this section is going to address this issue.\nOn one hand, the state of the art of gate driver is today based on CMOS technology. Specific control and gate driver dies have been developed and industrialized [1]. They offer CMOS benefits with ease of control and reduced static power consumption. A simplified and idealized model of such a CMOS gate driver can be obtained by the association of a DC pulse source and a gate resistor (figure 3.a). Our objective will be of course to match as much as possible with this state of the art. On the other hand, N-MOS technology is well known for its static power consumption making the classical polarization branch (figure 3.b) not suitable for monolithic integration and gate driver purposes. This is especially the case for the output stages of the gate drivers where nominal current levels can be in the range of amperes. The literature presents a set of solutions allowing to minimize the static consumption of circuits based on NMOS transistors [17, 18]. They are mainly based on a double supply voltage in order to allow large scale cascading structures. Based on these references and our own experience, an original structure has been identified. It complies with specifications and relies on the use of a N-MOS push-pull structure plus an inverting polarization branch for the control of the upper transistor. The topology is presented figure 3.c. All lateral transistors are comparable but the supply of the gate driver output stage is increased in order to account for high side transistor VGSth voltage drop.\n0-5V\nR = 50\nCG\nICMOS\na)", + "The integration of all devices complies with the specific isolation technique requirements presented in the previous section. In this case, neither the whole process, nor the characteristics of the vertical power device are affected. Simulations have been carried out in order to check what could be the time responses of such a structure compared with our idealized CMOS model (figure 3.a). Figure 4 presents the gate to source charge and discharge sequences for a CMOS based gate driver and for the topology described in this section. Table 1 presents the main simulation conditions. Models used for the simulation are level 1 MOS models and a physical model based on charge location as far as the vertical power device is concerned.\n0\n5.0\n500.0m\n1.0\n1.5\n2.0\n2.5\n3.0\n3.5\n4.0\n4.5\n0 200n20n 40n 60n 80n 100n 120n 140n 160n\nVCMOS\nVN-MOS\nVN-MOS+R\nIN-MOS *10\nICMOS *10\nIN-MOS+R\na)\nAs it can be seen on these results, the nominal polarization of the VDMOS is the 5V logic level. For this to append, the lateral devices of the output stage operate under 7V.\nThese choices and their possible consequences on the device characteristics are going to be investigated in details in the following section.\nA few comments can already be given. First of all, it appears that it is possible to design in N-MOS technology a gate driver output stage that offers dynamic characteristics comparable to ideal ones. This is obtained at reasonable efficiency level, as far as the output stage is concerned (50% in the ideal case and 25% in our case). The technique can be fully integrated at no extra technological step. However, the gate driver supply must set higher, around 7V considering our simulation conditions, in order to offer a nominal polarization level to the gate terminal of the vertical power device. Besides, the silicon surface consumption for the output stage of the gate driver is reasonable and represents about 1/10th to 1/20th of the VDMOS active region (when both of them share the same process and the same voltage terminations.\nThese encouraging results have pushed us toward deeper analysis in order to clearly identify what can be expected from this approach. The following section of this paper is going to analyze the static and dynamics characteristics of vertical and lateral components trying to extract functional and optimal tradeoff among them." + ] + }, + { + "image_filename": "designv6_24_0001886_iet-map.2010.0020-Figure19-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001886_iet-map.2010.0020-Figure19-1.png", + "caption": "Fig. 19 Layout of the direct-coupled band-stop filter employing the six-pole triple-mode conductor-loaded cavity filter", + "texts": [ + " Combining the conductorloaded and dielectric-loaded resonator results in a filter structure with a reasonable Q-factor that meets the requirement of many practical applications, whereas the presence of the conductor-loaded cavity resonator suppresses the undesired spurious modes of the dielectric resonator close to the operating band of the filter. IET Microw. Antennas Propag., 2011, Vol. 5, Iss. 10, pp. 1136\u20131142 doi: 10.1049/iet-map.2010.0020 Another potential application of the proposed resonator is in the design of a direct-coupled band-stop filter [14]. As suggested in the coupling diagram shown in Fig. 18, the triple mode conductor-loaded cavity filter can be reconfigured to realise a direct-coupled band-stop filter. The layout of the proposed band-stop structure is shown in Fig. 19. The corresponding rejection response of the bandstop filter with a Q-factor of 3000 is depicted in Fig. 20. The simulated frequency response in Fig. 20 indicates that the band-stop filter can create a rejection of 30 dB from 2140 to 2160 MHz, whereas maintaining a better-than0.25 dB insertion loss at the band edges of the pass band at 2125 and 2175 MHz. 1141 & The Institution of Engineering and Technology 2011 The six-pole triple-mode conductor-loaded cavity filter designed in Section 3 is fabricated and tested" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000083_ir-03-2018-0042-Figure26-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000083_ir-03-2018-0042-Figure26-1.png", + "caption": "Figure 26 The wearable vehicle leg model", + "texts": [ + " Industrial Robot: the international journal of robotics research and application Volume 46 \u00b7 Number 6 \u00b7 2019 \u00b7 740\u2013762 The goal of this controller is to make the wearable vehicle links follow the wearer while walking. Considering the wearable vehicle legs as serial manipulators, a PD controller with online gravity compensation (Kelly et al., 2006) whose block diagram is shown in Figure 25 is used to make the wearable vehicle track the human movement. The control torque of each joint is calculated according to equation (27). T \u00bc kp u d u\u00f0 \u00de1 kv _u d _u 1G u\u00f0 \u00de (27) According to the model of the exoskeleton leg shown in Figure 26, the potential energy of each link can be calculated by the following equation: The following is the derivation of the gravity compensation torques G(u ). These torques can be calculated from the total potential energyUof thewearable vehicle leg links equation (29). U \u00bc U1 1U2 1U3 (29) U1 \u00bc m1g L3sin u 3\u00f0 \u00de\u00f0 \u00de1L2cos u 1 1 u 2\u00f0 \u00de1 L1 2 cos u 1\u00f0 \u00de (30) U2 \u00bc m2g L3sin u 3\u00f0 \u00de\u00f0 \u00de1 L2 2 cos u 1 1 u 2\u00f0 \u00de (31) U3 \u00bc m3g L3sin u 3\u00f0 \u00de\u00f0 \u00de (32) Then, the gravity torque for each joint can be calculated as stated by Equation (33) to (35) and these values are used for compensating the effect of gravity on thewearable vehicle leg links" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001472_9781118354179.auto089-Figure22-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001472_9781118354179.auto089-Figure22-1.png", + "caption": "Figure 22. Schematic of the clutch control system.", + "texts": [ + " In such designs, both friction discs and pressure plates are manufactured from special carbon fiber materials, able to withstand high temperatures, providing high coefficient of friction values. These assemblies are extremely light and have low rotary moment of inertia but high costs. Typically, lives of such components are limited and measured in a number of races, with the race start being the most severe operating duty. Fundamental layout, components, and operation of a mechanical clutch control system are shown in Figure 4. The schematic diagram shown in Figure 22 further illustrates operation, giving lever ratios of each component: clutch pedal (ip), transfer mechanism (im), bearing fork (ib), and clutch lever (ic). As indicated, in order for the clutch to be disengaged, the product of the pedal force Fp, all lever ratios and overall mechanical efficiency \u03b7 Encyclopedia of Automotive Engineering, Online \u00a9 2014 John Wiley & Sons, Ltd. This article is \u00a9 2014 John Wiley & Sons, Ltd. DOI: 10.1002/9781118354179.auto089 Also published in the Encyclopedia of Automotive Engineering (print edition) ISBN: 978-0-470-97402-5 of the control system and clutch itself must be higher than maximum normal force FN. This must be satisfied throughout the necessary displacement of the pressure plate to ensure complete clutch disengagement in all operating and wear conditions. The ratios must be suitably chosen and efficiency sufficiently high to ensure low pedal force (typically under 200 N) and acceptable pedal travel (typically <150 mm). The schematic shown in Figure 22 relates to the mechanical clutch control system. In today\u2019s road vehicles, mechanical clutch control systems have found applications in lower performance cars. Usually, a steel cable is used to connect the clutch foot pedal or hand lever to the bearing releasing fork. Figure 23 (Nunney, 1998) shows a typical mechanical Encyclopedia of Automotive Engineering, Online \u00a9 2014 John Wiley & Sons, Ltd. This article is \u00a9 2014 John Wiley & Sons, Ltd. DOI: 10.1002/9781118354179.auto089 Also published in the Encyclopedia of Automotive Engineering (print edition) ISBN: 978-0-470-97402-5 solution" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001924_sp.j.2096-5796.2019.0016-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001924_sp.j.2096-5796.2019.0016-Figure2-1.png", + "caption": "Figure 2 Draper Watson's 6 DOF F/T sensor[25]. Figure 3 SRI 6 DOF F/T sensor[26].", + "texts": [ + " Multi-dimensional force sensor with integrated structure includes cylindrical elastic beam structure, cross elastic beam structure, asymmetric radial elastic beam structure, and composite elastic beam structure. The comparisons of main features of six-dimensional F /T sensors with various mechanical structures are given in Table1. For sixdimensional F / T sensors, the main principle is to test three directional forces and three dimensional moments by measuring the related strains of the elastic beam by strain gauges. A typical cylindrical elastic beam structure of 6 degree of freedom (DOF) F/T sensors are illustrated in Figure 2, which is called Watson force sensor developed by DRAPER Laboratory in the United States[24]. It consists of three vertical thin beams that are milled from a cylinder and distributed in the circumferential direction of 120\u00b0. The measuring circuit of the sensor is placed in the hollow cylinder. The F/T sensor has simple structure and high sensitivity but large interdimensional coupling errors, thus it needs decoupling calculation to obtain precise six-dimensional forces and torques. And the anti-overload ability of the sensor is very low, which means it is vulnerable to damage" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001238_ias.1998.732255-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001238_ias.1998.732255-Figure1-1.png", + "caption": "Fig. 1. Basic configurations of a IPM motor", + "texts": [ + " It combines the advantage of a permanent magnet (PM) motor with those of a reluctance one, featuring high efficiency and power factor, wide speed range, including flux-weakening (FW) operation, low maintenance requirement and low size and weight for a given torque [ 11. Fig.l shows two typical rotor configurations of an IPM motor. In Fig.l(a) the PMs are circumferentially magnetised, so that two magnets contribute to the flux of each pole, increasing the airgap flux density, while in Fig.l(b) the PMs are radially magnetised, each magnet supporting the whole flux of a pole. Both the configurations of Fig. 1 are considered in this paper. Due to rotor saliency the motor is usually modelled and analysed by means of the two-axis theory, which is well known and effective. However the involved rotor structure makes Wicult an analytical estimation of the magnetic parameters as stator inductances Id, 1, and stator flux linkage due to PM L (that are called motor parameters in the following of this paper). In addition, such a difticulty is dramatically increased if effect of iron saturation has to be taken into account", + " Coenergy Variation Magnetic Model obtained by the coenergy variation [IO], defined BP A possible approach is to consider a motor model Applying the separation of the variables, it yields to AW(id,i+=F(id)G(i,), and the two flux linkages can be expressed as d W , ) I , (id ,i,) = hq(O,iq) - F(id)-- di, (13) that can be computed from the characteristics hd(id,O), hd(id,i-), Aq(O,iJ, &(iamax,i&). Then the reciprocity property (6) is inherently satisfied. The model has been applied to a synchronous reluctance motor, yit:lding a good agreement with the measu.rements. It can be properly used for the motor analysis or for motor control, while it is difficult to extend it to the. motor design. Fig. 1 1. -0.018 2 -0.02 Y 0 50 100 150 200 250 300 350 * q [AI Flux linkage hpd vs. id, with different i, values B. Second Simplified Magnetic Model Another possible approach consists in considering constant PM flux linkage Am, obtained by FE field solution with id=O and i 8 . The d- and q-axis inductances are then complex functions of currents, always verifying (6). The corresponding model is given as follows &,(id,i,) =Am+&(id9iq)id (14) &,(id$,) = 1,Gd,iq)iq The drawback of this model is that infinite d-axis inductance occurs at low d-axis current and q-axis current merent from zero, as drawn in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000363_14644193jmbd244-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000363_14644193jmbd244-Figure1-1.png", + "caption": "Fig. 1 The four-node plate/shell element with variable thickness", + "texts": [ + " The method has advantages for dynamics design of complex flexible multi-body systems with flexible plates. The objective of this study is twofold: (a) to develop a new formulation for a plate/shell element with variable thickness based on ANCF continuum mechanics fully parameterized; to the authors\u2019 knowledge, this paper is the first attempt because all the preceding papers have focused on the plate with uniform thickness using ANCF; (b) to verify the results of the proposed element with the analytic results, experimental data, and ABAQUS, ANSYS, and NASTRAN commercial finite-element software. Figure 1 shows a quadrilateral plate/shell element of length a and width b, and it has four nodes (A, B, C , D) of variable thickness (hA, hB, hC, hD), respectively. XYZ represents the global inertial coordinate system. Using finite-element method, the plate is divided into ne elements, where ne is the number of elements of the body i, and xyz represents the element local coordinate system for element j(j = 1, 2, . . . , ne). It should be pointed out that each element j has variable thickness. The spatial coordinate z is assumed to be in a direction perpendicular to the mid-surface of the plate/shell in the undeformed state" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000193_0954406220906246-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000193_0954406220906246-Figure3-1.png", + "caption": "Figure 3. Free body diagrams of (a) total system, (b) rigid shackle, and (c) curved beam.", + "texts": [ + " Kinematic constraint equation of the three links is presented below \u00bdLc \u00fe R cos \u00f0\u2019i\u00de 2 \u00fe \u00bdR sin \u00f0\u2019i\u00de 2 \u00bc \u00f0Li\u00de 2 \u00f03\u00de In the above equation, \u2019i is the angular orientation of the shackle with respect to Xi axis. Expression of \u2019i in terms of the three links is presented below \u2019i \u00bc cos 1 \u00f0Li\u00de 2 \u00f0Lc\u00de 2 R2 2Lc R \" # \u00f04\u00de Relation between the orientation angles i and \u2019i is also obvious from Figure 2 as given below Li sin \u00f0 i\u00de \u00bc R sin \u00f0\u2019i\u00de \u00f05\u00de Kinetics of leaf spring system Free body diagram of the leaf spring system at deformed configuration of current load step i is shown in Figure 3. To explore kinetics of the system in detail, total system kinetics and force and moment balance conditions of its individual components are considered in Figure 3(a) to (c). The constraint in the rotation of the rigid shackle develops an end moment Mi B at point Bi as shown in free body diagram in Figure 3(b). Using the force balance conditions along horizontal and vertical directions, Ri CH \u00bc Ri BH and Ri CV \u00bc Ri BV, the moment balance equation of the shackle about point Bi becomes Mi B \u00bc R sin \u00f0\u2019i\u00deRi BH R cos \u00f0\u2019i\u00deRi BV \u00f06\u00de The genesis of the end moment Mi B necessary for restoring system equilibrium is modeled theoretically by a moment generated by an elastically restraint rotational spring. Incremental restoring moment for the current load step (Mi r) is given by Mi 1 r \u00fe Ki \u2019 \u00f0 \u2019 i\u00de, where Ki \u2019 is deformation dependent incremental stiffness of the rotational spring and \u2019i is the incremental rotation of the rigid shackle given by \u00f0\u2019i 1 \u2019i\u00de", + " As the end moment Mi B counter balance Mi r at current load step i as well as at the previous step (i 1), its expression is given by Mi B \u00bcMi 1 B \u00fe Ki \u2019 \u00f0 \u2019 i\u00de \u00f07\u00de The equation for incremental spring constant Ki \u2019 is obtained through energy balance condition 1 2K i \u2019 \u00f0 \u2019 i\u00de 2 \u00bc Ui e \u00feUi s \u00feUi l, where Ui e, Ui s, and Ui l are incremental energies associated with work done by external load, internal strain, and locked-up stress field, respectively. Expression of energy due to work done by external load is given by Ui e \u00bcWi \u00f0 Yi W\u00de, where \u00f0 Yi W\u00de is incremental deflection of loading point at current load step i as given by Yi W \u00bc \u00f0Y i 1 W Yi W\u00de. Here, Yi 1 W and Yi W are coordinates of the loading point in global frame at previous (i 1) and current load step i (refer Figure 3(a)). As the energies Ui s and Ui l are associated with deform- ation of curved beam part with respect to updated geometry, their expressions are indirectly available in the deformation analysis of curved beam as presented in the next sub-section. However, the expressions of Ui s and Ui l are clearly described in the numerical solution scheme as presented in the \u2018\u2018Numerical solution scheme\u2019\u2019 sub-section. Now putting the expression of Ui e and keeping the notations of Ui s and Ui l in the energy balance equation, the rota- tional spring stiffness is obtained as given below Ki \u2019 \u00bc 2 Wi \u00f0 Yi W\u00de \u00feUi s \u00feUi l \u00f0 \u2019i\u00de2 \u00f08\u00de The rotational spring stiffness Ki \u2019 described above is defined in incremental way through geometry updation, and hence it may be termed as true stiffness of the rotational spring", + " Such definition of spring stiffness is termed as total stiffness K0 \u2019 and its expression is given below K0 \u2019 \u00bcMi B=\u00f0 \u2019 0\u00de \u00f09\u00de However, it should be noted that the total stiffness K0 \u2019 is only considered to observe effect of geometry updation in system characteristics and is not incorporated in the mathematical model. The counter moment of Mi B is developed in the curved beam at the hinge point with rigid shackle (Bi). Free body diagram of the curved beam showing reaction forces at supports beside the end moment at Bi is presented in Figure 3(c). Force balance conditions of the curved beam in vertical and horizontal directions give the following two equations Ri AV \u00fe Ri BV \u00bcWi \u00f010\u00de Ri BH Ri AH \u00bc 0 \u00f011\u00de Similarly, moment balance condition of the curved beam about point A gives \u00f0Lc \u00fe R cos \u2019i\u00deRi BV \u00f0R sin \u2019i\u00deRi BH \u00bcWi Xi W Mi B \u00f012\u00de Resultant of reaction forces Ri BH and Ri BV at point Bi along ABi direction is given by Ti BR \u00bc Ri BH cos \u00f0 i\u00de \u00fe Ri BV sin \u00f0 i\u00de \u00f013\u00de As mentioned earlier to address the sheer impossibility of existence of rotational restraint of the present model, two more models are derived from the main model through post processing" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003234_cp:19970239-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003234_cp:19970239-Figure1-1.png", + "caption": "Figure 1 : Bifurcated waveguide system with T-slot a) Conventional T-slot", + "texts": [], + "surrounding_texts": [ + "1.207\nlower surface (SI) when the slot is closed by perfect conductors so that the scattered electromagnetic field remains unchanged. By applying continuity conditions for the electric field on the surfaces\n(1) E, x ( Z a - E , , ) = ~ on S, E, x (Ec -E,,) = o on s2\nand expressing these fields in terms of magnetic-type dyadic Green\u2019s functions, a pair of simultaneous integral equations can be written as follows. - ny x RI = j a e K, x il,, Fma (r IF,)) + Zmb(r IF,,)]\n\u2018fie solution to these equations is obtained by using the moment method which converts the pair of integral cquations into a matrices equation of the following iFOlTIl\nwhere the [A] matrix represents the self-coupling \u2018between magnetic current elements on the lower slot surface SI , plus in case of a dual-mode bifurcated waveguide the mutual coupling between currents on this surface (SI) and on the surface S3 of the clearance e [B] matrix represents the self-coupling between magnetic current elements on the upper surface S2 and the [C] matrix represents the mutual coupling to the lower surface SI due to magnetic current elements on the upper Sz, while the [Dl matrix represents the mutual coupling in the opposite direction. The individual matrix elements are given as foollows (4)\nwhere gq, gg and 8, are unit vectors in the slot CO-\nordinate system as depicted in figure la. In equation 3 the entries of matrices [a] and [b] represent the unknown amplitudes of the basis function series used to model the slot field distributions on the surfaces SI , S 2 and S3 , wWe the [h] matrix provides the contributions from the incident mode at the aperture.\nConvergence Tests\nIn order to establish the integrity of the algorithms which have been developed to solve the above matrix equation (eq. 3) tests were performed both with increasing numbers of images used in modelling the parallel plate waveguide Green\u2019s function, and with increasing numbers of basis functions. These tests are summarised in tables 1 and 2. They show that the solutions for typical longitudinal and transverse slot examples converge smoothly with no more than 20 images being required for acceptable accuracy. Different convergence rates are however observed with basis function numbers. For typical longitudinal slots convergence occurs at about 5 basis functions whereas it requires 13 basis functions to ensure Convergence for a typical transverse T-slot for which 0.8 (L/2) < L3 < L/2. More recent results, which will be described in the presentation, show that this restriction is less severe when the edge-condition is incorporated into the T-slot computations.\n\u2019The admittance characteristics of a typical longitudinal off-set slot radiating into a poiariser are depicted in Ifigure 3. These were performed for 3 basis functions and 20 images. The measurements were performed on an HP8510 vector network analyser and the TRL method was used to de-embed the coax-to-waveguide transitions. With experimental errors estimated to be of the order of 6-7% good agreement between theory and measurement is indicated.\nFigure 4 is typical of a range of measurements which have been performed on an even-m\ntransverse T-slot (vertical clearance bifurcated rectangular waveguide. The measured results are, again, in good agreement with the predicted values for slot resistance and reactance. Impedance behaviour as a function of slot length is correctly predicted, while the onset of resonance is accurately located. Nevertheless, some relatively large deviations between theory and experiment are shown. These can attributed to additional errors which are inherent in the transverse slot measurements. Firstly any departure from true \u2018eveness\u2019 of the exciting mode in the bifurcated waveguide generates unwanted odd mode scattering, as does any asymmetry in the location of the T-slot with respect to the guide centre line. Also the experimental slots were round-ended rather than square-ended as required by the theory.\nThe above computations on the conventional T-slot show that it is relatively weakly coupled to the\n10th International Conference on Antennas and Propagation, 14-17 April 1997, Conference Publication No. 436 0 IEE 1997", + "1.208\nwaveguide mode by comparison with the longitudinal offset slot. For polarisation agile array design where slots with wide range of coupling may be required, this represents a significant weakness in the applicability of the conventional T-slot in this role. Preliminary theoretical and experimental studies on the modified Tslot indicate that it is much more strongly coupled to the waveguide mode and is therefore more versatile in array implementation terms.\n~ o ~ p ~ ~ a ~ ~ o n s on a polarisation agile slot pair, comprising a longitudinal slot in combination with a transverse T-slot designed to produce circular polarisation in the polariser at 10 GHz, are summarised in figure 5. It shows the axial ratio at the mid-plane within the polariser plotted as a function of frequency. It can be seen that useful C.P. performance ( 0) is that technological treatment do not has different effects. Statistical testing for the significance of the difference after recommended treatment based on small samples. If the data show a statistically significant change in the specimen receiving the treatment, the null hypothesis is rejected. Typically the null hypothesis is rejected, when the p-value is less than the significance level \u03b1, which value typically is \u03b1 = 0.05. The term significance level \u03b1 is used to refer to a pre-chosen probability and it is present maximum (critical) level of the null hypothesis rejection, which corresponds to a certain value of the test statistic. The significance level is the probability of making a type I error, or the probability that we will reject the null hypothesis when it is in the long run true. If this calculated probability of rejecting the null hypothesis (H0) is less than 0.05, that the difference is significant, the difference is not caused by chance. We must interpret the results variously depending on whether the p value is small or large. If the p value is small, then it is highly improbable that the treatment effect observed is due to mere coincidence of random sampling and experimental setting, if p < 0,05 the null hypothesis is rejected. If the p value is large, the data do not show sufficient reason to deduce that the treatment had an effect. The results for our t-test in a succinct notation twosample t(df) = t-value, p = p-value is: t(11) = 0,02, p = 0,985. The results in our experiments are not statistically significant. That is, the difference in the mean strength results of these two groups could likely have occurred by chance. Proc. of SPIE Vol. 8763 876309-7 Downloaded From: http://proceedings.spiedigitallibrary.org/ on 02/10/2014 Terms of Use: http://spiedl.org/terms" + ] + }, + { + "image_filename": "designv6_24_0000885_j.apm.2014.04.032-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000885_j.apm.2014.04.032-Figure8-1.png", + "caption": "Fig. 8. The angular capability angles (r\u2013s): the casing plane (a), the circular diagram (b).", + "texts": [ + " It must be specified (based on the aspects presented in the previous sections) that all motion functions are pre-determined, so the positions and orientations of the elements are known (depending on the kinematic independent parameters la and/or u5). The angular capability is defined by two angles, r and s: r represents the angle between the unit vector up of the nut and the normal axis ns to the casing plane (S), while s is used to define the orientation of the unit vector up, being defined between the characteristic line dc of the casing plane and the projection up p of the unit vector up on this plane (Fig. 8a). In these terms, Please degree r \u00bc \\\u00f0 ns; up\u00de; s \u00bc \\\u00f0dc; up p\u00de: \u00f015\u00de The angular parameters r and s will be represented in a circular diagram, with s in trigonometric direction and r having a radial measure, marking a point J for a current pair rj\u2013sj (Fig. 8b). With the previous notations (see Fig. 2a), the unit vectors up of the nuts rods are uM , uN and uC . They are positioned by the angles eM\u2013kM, eN\u2013kN, eC\u2013kC relative to the local axis z2 in the wheel carrier reference frame x2y2z2 (the nuts are fixed in the wheel carrier). In a similar way, the unit vectors ns of the normal axis to the casing plane will be ns1 , ns3 and ns4 (see Figs. 2b and 3). They are positioned by the angles l1\u2013g1, l3\u2013g3, l4\u2013g4 relative to the local axis z1,3,4 in the control arms reference frames x1y1z1, x3y3z3, respectively in the tie-rod reference frame x4y4z4" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001212_6.2008-5955-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001212_6.2008-5955-Figure1-1.png", + "caption": "Figure 1. Hierarchical Aircraft Decomposition", + "texts": [ + " Thus, designers have the flexibility to decide which design algorithm, and disciplinary analysis tools to employ, the level of fidelity of such analysis, and over what specific mission flight phase they should be applied. The figures throught this section use the standard unified modelling language (UML) representation for class diagrams.26 A. The Aircraft Geometry Class An aircraft layout is defined using a hierarchical component association. Individual aircraft components are identified, classified, and associated based on configuration and utility as shown in Figure 1. Internal to each element are sub-components, which can be included in the design optimization process depending the level of fidelity used by the disciplinary analysis. Components as well as subcomponents have specific sets of parameter definitions, which form the building blocks of the aircraft model based on the physical characteristics that must be specified. For example, we can have a wing with a control surface. In the aircraft hierarchy, the wing will represent a main component and the control surface will be an associated component (subassembly)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000855_0954406220936732-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000855_0954406220936732-Figure4-1.png", + "caption": "Figure 4. The structure of series limb: (a) the series limb; (b) the hinge and thin beam.", + "texts": [ + "( ) I2\u00bc 1 8r\u00f0n\u00f0n\u00fe 2\u00de\u00de3 8n4 \u00fe 32n3 \u00fe 57n2 \u00fe 50n\u00fe 15 6\u00f0n\u00fe 1\u00de2 \u00fe 5\u00f0n\u00fe1\u00de2ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi n\u00f0n\u00fe 2\u00de p arctan ffiffiffiffiffiffiffiffiffiffiffi n\u00fe 2 n s ) I3 \u00bc 1 8r2\u00f0n\u00f0n\u00fe 2\u00de\u00de3 6n5 \u00fe 30n4 \u00fe 70n3 \u00fe 90n2 \u00fe 59n\u00fe 15 6\u00f0n\u00fe 1\u00de3 \u00fe 4n2 \u00fe 8n\u00fe 5ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi n\u00f0n\u00fe 2\u00de p arctan ffiffiffiffiffiffiffiffiffiffiffi n\u00fe 2 n s ) I4 \u00bc 3 2:96r3\u00f0n\u00f0n\u00fe 2\u00de\u00de3 6n5 \u00fe 30n4 \u00fe 70n3 \u00fe 90n2 \u00fe 59n\u00fe 15 6\u00f0n\u00fe 1\u00de3 \u00fe 4n2 \u00fe 8n\u00fe 5ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi n\u00f0n\u00fe 2\u00de p arctan ffiffiffiffiffiffiffiffiffiffiffi n\u00fe 2 n s ) I5 \u00bc 1 8r3\u00f0n\u00f0n\u00fe 2\u00de\u00de3 6n5 \u00fe 30n4 \u00fe 70n3 \u00fe 90n2 \u00fe 59n\u00fe 15 6\u00f0n\u00fe 1\u00de3 \u00fe 4n2 \u00fe 8n\u00fe 5ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi n\u00f0n\u00fe 2\u00de p arctan ffiffiffiffiffiffiffiffiffiffiffi n\u00fe 2 n s ) \u00f010\u00de where, \u00bc t/2 r Stiffness modeling of each branch There are 12 limbs in the heavy-load-bearing flexible parallel force sensor, of which eight are flexible series limbs and the remaining four are heavy-load-bearing limbs. Heavy-load-bearing limbs can be regarded as cantilever beam structure, whose stiffness modeling method is the same as that of flexible link, and its local coordinate system coincides with its reference coordinate system, so there is no need to transform the coordination. The eight flexible series limbs are identical. Each limb is composed of two flexible thin beams, two biaxial circular flexure hinges and an intermediate flexible link, which is the force sensing unit. As shown in Figure 4, L1 is the height of the flexible thin beam, b is the length of the flexible thin beam section, rs is the radius of flexure hinges, t is the minimum square section length of flexure hinges, l is the length of the link, is the deflection angle of the link, and H is the height of the series limb. As shown in Figure 4, the series limb can be divided into five parts, and the local coordinate systems are established at the end of each element, respectively. The {O1} is the local coordinate system at the end of the flexible thin beam, the other end of which is fixed on the fixed platform. {O2} and {O4} are the local coordinate systems at the end of the biaxial circular flexure hinges. The {O3} is the local coordinate system at the end of the flexible link. The {O5} is the local coordinate system at the end of the flexible thin beam, the other end of which is fixed on the moving platform" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000967_aenm.202100038-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000967_aenm.202100038-Figure1-1.png", + "caption": "Figure 1. a) The schematic of one water-tube-based triboelectric nanogenerator (WT-TENG). b) The working mechanism of WT-TENG. c) The schematic of the four working modes of WT-TENG and the application scenarios.", + "texts": [ + " Based on these findings, we fabricate a WT-TENG wristband with 10 tube units and a WT-TENG box with 34 tube units for body motion and water wave energy harvesting, respectively. Based on our proposed strategy of WT-TENGs, it is expected that various advanced TENGs and self-powered systems will be developed to effectively harvest energies from multiple renewable mechanical energy sources, including ocean, wind, vibration, and biomechanical motions. The schematic of the WT-TENG and the corresponding fabrication process are demonstrated in Figure 1a and Figure S4, Supporting Information, respectively. An FEP tube (with the length of 6\u201318\u00a0cm and diameter of 1\u20132\u00a0cm depending on the conditions) was utilized as a negative tribo-material and the container of the water. The FEP surface\u2019s static contact angle (CA), advancing CA, and receding CA of FEP are 99\u00b0, 122\u00b0, and 88\u00b0, respectively (Figure S5, Supporting Information). DI water was filled into the FEP tube as the positive tribo-material. For all the samples in this work, the water was filled half in the tube to achieve a maximum charge output (detailed explanations are shown in Note 1, Figures S1, and S2, Supporting Information). The two ends of the tube were wrapped with Teflon tapes. Two Cu tapes were wrapped side by side outside of the FEP tube as two electrodes. Figure\u00a01b demonstrates the working mechanism of one WT-TENG unit. When DI water contacts FEP, due to the contact electrification effect, charges tend to transfer at the solid and liquid interface. As a result, DI water will be positively charged and the inside surfaces of the FEP tube will be negatively charged.[28,34] Relative displacement between water and tube can be easily triggered by an external mechanical motion. Once the water carrying charges starts to move in the tube, the electrostatic balance of electrode A and electrode B (shown in Figure\u00a0 1b) is broken, and thus leads to the rearrangement of charge distribution in these two electrodes to rebalance the electrostatic status and charge transfer in the external circuit. Consequently, electric currents could be generated by shaking, rotating, swinging, and levering the WTTENG, or applying any other mechanical stimulations, as long as the external mechanical motion could trigger the water fluid moving between regions of the two electrodes. The flexibility of water enables the WT-TENGs to work in various modes. We mainly demonstrate four typical working modes\u2014rotation, swing, seesaw, and horizontal linear modes in this paper (Video S1, Supporting Information). As shown in Figure\u00a0 1c, these four modes are involved in many environmental mechanical motions. Taking ocean waves as an example, the amplitudes and the frequencies of ocean waves are both irregular and complicated. Considering an object that is immersed in the ocean water or floating on the water surface, its vibration and movements must contain several modes, with all the swing, seesaw, and horizontal linear motions involved. As another example, our body movements also contain several motion modes, such as the swing motion occurs when we bent elbows, and the linear motion happens when we shake arms during running", + " This can be understood as that the high ion concentration leads to ion accumulation at the solid/liquid interface, and thus the electron transfer process is hindered due to the screening effect.[36] The influence of the tube diameter on the performance of WT-TENG in the rotation mode is shown in Figure\u00a0 3b. With the tube diameter ranging from 1 to 1.5\u00a0cm, the outputs of the WT-TENG increased with the tube diameter, due to the increase of water/solid contact area and hence the charge transfer. However, we noted that the outputs of WT-TENG decreased when the tube diameter is above 1.5\u00a0cm. According to the mechanism of WT-TENG shown in Figure\u00a0 1b, the amount of transferred charges is highly related to the FEP area wrapped by the Cu tapes and be little affected by distance between the two electrodes. In Figure\u00a03c,d, we verified these assumptions. A 10\u00a0 cm WT-TENG with tube diameter of 1.2\u00a0cm operated in rotation mode was used to conduct these tests. In Figure\u00a03c, the lengths of the two electrodes are both 1\u00a0cm. When we varied the distance between the two electrodes, the output of the WT-TENG unit kept roughly constant. In Figure\u00a03d, we show that the outputs of WT-TENG are positively proportional to the length as well as the area of the electrodes" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000555_amm.486.239-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000555_amm.486.239-Figure1-1.png", + "caption": "Fig. 1 The basic 3D model for experimental and numerical", + "texts": [ + " For real objects (such as road or track vehicles) the third case usually (almost always) occurs [6-8]. All rights reserved. No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of Trans Tech Publications, www.ttp.net. (ID: 130.237.29.138, Kungliga Tekniska Hogskolan, Stockholm, Sweden-11/07/15,03:50:37) For the analysis of the dependence of geometric and production asymmetry as well as the unbalanced excitation on vibration elements, a simple model was created (see Figure 1). The model can be solved experimentally and numerically. The model system consists of steel plate, which is placed on four coil springs. Two weights were used to simulated the geometric asymmetry (total weight is equal to half-weight of the plate). These weights are placed on the plate in different combinations. One option of symmetrical arrangement and four options of asymmetrical arrangement were chosen for our research. For experimental and numerical solution of vertical oscillation of a mechanical system, one case of symmetric excitation (drop of all four springs at the same moment) and four options of unbalanced excitation (drop of one, two or three springs in different combinations) were selected" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001026_tap.2015.2487512-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001026_tap.2015.2487512-Figure2-1.png", + "caption": "Fig. 2. Dielectric plunger positions to maximize (top) and minimize (bottom) the effective dielectric constant of the parallel plate waveguide controlling the phase delivered to each element.", + "texts": [ + " Another approach is to use multiple ports to create beam steering [30], but many ports are needed for fine beam resolution. In this paper, we avoid the aforementioned issues by instead using a movable dielectric plunger within the feedline of the TWA. Figure 1 depicts the proposed TWA configuration. This TWA employs a parallel plate waveguide (PPW) feed supporting a propagating wave that excites slot antennas to the right in a sequential manner. The propagation constant within PPW is controlled using a small, linear mechanical movement (Figure 2). This motion changes the feedline\u2019s air to dielectric ratio and has several advantages: 1) low cost, 2) low complexity, 3) large total bandwidth, 4) large scan ranges (compared to frequency scanned arrays). In [31] we employed a parallel plate waveguide (PPW) transmission line (TL) feed by placing a dielectric coating at the inner side of the planar strips forming the TL. By moving 0018-926X (c) 2015 IEEE. Personal use is permitted, but republication/redistribution requires IEEE permission. See http://www", + " As the level of insertion of the plunger within the coplanar stripline varied, the percent of dielectric filling the coplanar stripline increased or decreased to change the TL feed\u2019s effective dielectric constant and therefore, the propagation constant. However, this design required precise mating of the trapezoidal dielectric plunger and feedline, limiting its realizability. This paper builds on the experiences in [31] and [32] to propose a much simpler loaded transmission line such as that shown in Figure 2. Specifically, the employed TL supports a parallel plate waveguide mode formed by the vertical metallic faces. Cavity-backed slot antennas are also placed at the surface of the metallic slabs forming the radiating elements. The rectangular dielectric is moved up and down along the z-axis to control the propagation constant of the TL. Several simplifications of the design provided ease of fabrication. Specifically, the rectangular plunger and the solid metal plates can be more precisely fabricated for Ka and Ku band applications", + " Below we present the design of the new TWA concept, its excitation, beam steering performance, and validation using a fabricated prototype operating at 13GHz. As noted, beam steering in a TWA is achieved by controlling the propagation constant of the feedline. Here, we employ a dielectrically loaded parallel plate waveguide (PPW) TL to feed an array of cavity-backed slots as depicted in Figure 1. A rectangular dielectric is positioned within the vertical PPW plates. Specifically, a section of the PPW transmission line is depicted in Figure 2 and as the dielectric plunger moves up and down along the z-axis, the propagation constant changes. The key parameters of the dielectrically loaded PPW are depicted in Figure 3. As expected, the maximum and minimum effective dielectric constant (\u03b5eff ) of the PPW depend on the position of the dielectric plunger. Also, the dielectric composition of the plunger plays a crucial role in determining \u03b5eff . Roger\u2019s TMM13i (\u03b5r = 12.85, tan \u03b4 = 0.0019) was chosen for the dielectric plunger because of its low loss and high permittivity" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002224_3527608117-Figure5.18-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002224_3527608117-Figure5.18-1.png", + "caption": "Fig. 5.18 Principle of plasma transferred arc welding (ISAF, TU Clausthal).", + "texts": [ + " Plasma MIG cladding is characterized by high deposition rates up to 25 kg h\u20131 as well as surface performance of 5000 cm2 h\u20131 with relatively small surface energies of 150 J mm\u20132. Layer thicknesses are usually 4 to 6 mm. Typical applications are 1415.3 Cladding armouring with tough or hard steel, cladding with corrosion resistant steel and cladding of copper- or nickel-based alloys of the same kind. However, due to relatively high machine effort the process has not found widespread application so far. 5.3.3.7 Plasma Powder Transferred Arc Welding (PTA) During PTA welding (Fig. 5.18), a plasma jet produced by a nonmelting electrode is used as a heat source. Two arcs are used, which are controllable over separate power sources. The so-called pilot arc burns after ignition by high frequency between a rod-shaped cathode and an anodically polarized ring nozzle. The pilot arc is used for ignition of the transferring main arc between work piece and pin cathode. The thermal load is kept relatively small by a negative polarity. Plasma powder transferred arc welding uses powdered weld additives, which are supplied to the burner by a transfer gas" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000561_cp.2010.0023-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000561_cp.2010.0023-Figure4-1.png", + "caption": "Figure 4: Rim-driven motor FE model.", + "texts": [], + "surrounding_texts": [ + "This paper explores the use of a hybrid 'canned' induction rotor design in which the conventional cage of an industrial motor is replaced by a simple cage with a low radial depth, along with a series of air-gap cans to provide environmental shielding, and a conducting rotor can to simplify the rotor design. Figures 3 and 4 show a cross-sectional view of the 2D finite element (FE) models of a conventional (benchmark) industrial motor and the proposed simplified rim-driven canned motor, respectively. In a conventional industrial motor, the rotor bar is designed to use the 'deep-bar' effect to increase starting torque. The simplified bar design of the rim-driven motor therefore, would result in a significant loss of starting torque. This is partially circumvented, however, by including a simple conducting can directly underneath a containment (environmental) can on the rotor outer surface, as illustrated in Figure 5 . Canned rotor construction has been used successfully in the past in other applications [5]. The conducting rotor can (and also containment can) here acts as a secondary high-resistance cage whose main function is to boost the starting torque of the motor. The unusual combination of a simple rotor bar design and the conducting rotor cans allows both the radial depth and weight of the rotor to be reduced without significant loss of performance. The objective was to design a rim-driven motor with the following specification requirements: a rotational speed of approximately 575 rpm, 440 V 3-phase supply, 60 Hz and a torque of 3,000 Nm. The motor would be located on the outer rim of a nickel-aluminium-bronze (NAB) tube around the propeller blade tips. The outer diameter of the propeller was 0.8 m including a 20 mm radial thick NAB tube/ring for structural support. The inner diameter therefore, of the rim driven rotor was 0.84 m. A conventional industrial induction machine design was modelled using the Flux2D FE analysis software package. This was used as a benchmark to compare the performance between an industrial motor and the subsequent rim-driven motor design. The industrial motor specification was chosen to have approximately the same torque rating at full-load as the rim-driven motor. This was based on the reasonable assumption that the torque largely dictates the size and weight of a motor. One of the key differences between the benchmark industrial motor and the rim-driven design is the air-gap. The philosophy in a conventional induction motor design is to minimise the air-gap, subject to mechanical and thermal constraints. The benchmark industrial design had an air-gap of 1.5 mm. The rim-driven design on the other hand has to accommodate several non-magnetic cans within the air-gap and due to the larger diameter, mechanical requirements dictates a minimum air-gap of 3.5 mm. The rim-driven design therefore, has a relatively large air-gap. Firstly, there is a can on the inner stator surface; this is made of stainless-steel and is utilised for environmental protection. Secondly, there are two cans on the rotor outer surface: one conducting, which is copper in this instance, and one environmental, again stainless-steel. The stainless-steel environmental cans are 0.5 mm thick and the copper conducting can is 1 mm thick, producing a total air-gap length of 5 .5 mm. This has a significant impact on the design of the rim-driven motor and results in a large magnetising current. The environmental cans were chosen to be as thin as possible to essentially minimise the air-gap length and therefore keep the magnetising current down whilst still maintaining their structural integrity. The radial thickness of the copper can was chosen based upon its effects on the motor performance. The large diameter of the rim-driven design means that the axial length can be reduced significantly in comparison to a conventional machine. It is important, therefore, to minimise the overhang length of the end-winding basket to prevent the end-winding impedance having a strong detrimental effect on the performance of the motor. This would also help to keep the weight and volume to a minimum. A conventional 12-pole winding is used to meet the low-speed requirement of the rim driven design. The high pole number has the advantage of reducing the overhang length of the end-winding basket. A double-layer winding is utilised to further reduce the end winding basket. This would also increase the efficiency of the motor by reducing some of the low order mmf winding harmonics, as illustrated in Figure 6. This is additionally beneficial in reducing the eddy-current losses in the stator containment can; these can be significant because the can is stationary with respect to the rotating air-gap fields. Efficiency, however, was not the main driver in the design of this motor due to its low duty-cycle operation. In conventional induction motor design, the air-gap emf induced per phase in the stator winding is assumed to be approximately equal to the terminal phase voltage, which is then used to set the stator winding turns for a desired air-gap flux density. The large magnetising current of the rim-driven design, however, resulted in a lower specific magnetic loading using this approach. This was compensated for by increasing the required number of turns to bring the specific magnetic loading back up to the desired level of approximately 0.5 T, which produced a peak air-gap flux density of approximately 0.8 T, as shown in Figure 7. Furthermore, the electric loading of the machine was increased because of the availability of water cooling. This motor is to be operated in a sea-water environment. Consequently, the motor will be submersed during operation and, as a result, sea-water will be present in the air-gap in direct contact with the environmental cans on the stator and rotor. Furthermore, the increased surface areas of the rotor and stator cores means that substantially improved cooling of the machine is available when compared to a conventional air cooled machine. The stator conductors therefore, were designed to operate at up to 8 Almm2\u2022 This was almost three times more than the full-load current density of 2.89 Almm2 in the benchmark design. The rotor bar load current is related directly to the stator estimated load current. The rotor bars, however, are normally worked at a higher current density than the stator conductors. The rotor bar material was also changed from aluminium, used in the benchmark motor, to copper to further reduced the required bar area. The rotor bars therefore, were designed to operate at approximately 16 Almm2\u2022 Machines are usually designed to be worked just below saturation level to make effective use of the lamination steel in the core-backs. The stator and rotor core-back of the rim driven motor were designed therefore, to operate at a magnetic flux density of approximately 1.45 T. However, the approximation is less reasonable in the case of the conducting cans because in reality the currents are not constrained to flow axially in the solid material. As such, it cannot be identified as a circuit and hence, it is not possible to add an effective external end-ring impedance. As shown in Figure 8, the air-gap field produces circular current paths in the conducting can, which are not restricted and can flow in any direction. Currents in a caged rotor on the other hand are constrained to flow axially down the rotor bars and around the end rings. .... ..0:- \ufffd+-+-\ufffd\ufffd+- \ufffd ..0:- \ufffd - - ...... -4 \ufffd----=..\ufffd\ufffd----\"'----=..\ufffd _ _ _ I ,, <\"'\"\ufffd\ufffd......-+-\ufffd+-\ufffd \"' ..... \" .---\ufffd----=..-+\ufffd\ufffd\ufffd ..... \\ I / ,r<,\ufffd---\"--\" , \\ , I / /' /\ufffd\"\"'-'-410-.\",..-..,.\"\"\">.\" , \\ I I ///_0=>Gi >0 -{KT/J)Zl < 0 if Sxx <0=>Z t > 0 c-(KTG2 + D)/J < 0 if Sx2 > 0 => G2 > (cJ - D)/KT c-(KTZ2 + D)/J > 0 if Sx2 < 0 => Z2 < (cJ - D)/KT -(KTKf/J)S + TJJ < 0 for abs (S) > e => Kf > TJ(KTe) (20) 7.3 Experimental setup The mechanical arrangement for varying the inertia and load disturbance is shown in Fig. 17. The moment of inertia is varied by adjusting the position of the mass with respect to the shaft. The load disturbance at the shaft of the motor follows a sine function (TL = ML sin (6)); thus the load disturbance is nonlinear and dependent on the positon of the rotor. The experiments were conducted with a step reference input from 0\u00b0 (0 rad) to 90\u00b0 (1.57 rad) for both schemes and the three load tests. 7.4 Experimental results and performance evaluation The conditions for the load tests are given in Table 3" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002224_3527608117-Figure11.8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002224_3527608117-Figure11.8-1.png", + "caption": "Fig. 11.8 Schematic overview of the spray forming unit.", + "texts": [], + "surrounding_texts": [ + "Spray forming can be considered as a further development of atomization. Spray formed (spray compacted) materials differ from the RS atomized alloys in the extremely low gas content in the form of hydrogen due to the inert gas atmosphere (nitrogen) during the spray process. This extremely low hydrogen content means that the spray formed materials are weldable by fusion weld techniques. The pure atomization technique is identical to the RS technique. Because the semi-finished product is in the form of a compact billet, cold and hot isostatic pressing is not applied at the spray forming stage, which is different to the RS technique. Figures 11.8 and 11.9 show the scheme of a spray forming plant, as is applied at the PEAK Werkstoff GmbH for special alloys and development concepts. From the Tundish the melt is dispensed onto the two atomization units. The melt is atomized by nitrogen and the resulting drop spectra is accelerated onto the sample. The impinged droplets solidify at the surface and little by little a structure grows, forming an almost perfect cylindrical bolt by constant rotation round the vertical axis and a continuous downward withdrawal rate. 282 11 Spray Forming \u2013 An Alternative Manufacturing Technique for MMC Aluminum Alloys 11.3.3" + ] + }, + { + "image_filename": "designv6_24_0003507_speedam.2016.7525825-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003507_speedam.2016.7525825-Figure3-1.png", + "caption": "Fig. 3: Load magnetic circuit [17]", + "texts": [ + " From (3) to (6), the different inductances can be calculated as follows: Ldd (Id, Iq) = \u03a8d (Id, Iq)\u2212\u03a8d (0, Iq) Id (7) Ldq (Id, Iq) \u03a8d (Id, Iq)\u2212\u03a8d (Id, 0) Iq (8) Lqq (Id, Iq) = \u03a8q (Id, Iq)\u2212\u03a8q (Id, 0) Iq (9) Ldd (Id, Iq) = \u03a8d (Id, Iq)\u2212\u03a8d (0, Iq) Id (10) A proper calculation of the phase magnetic flux linkage is crucial for the correctly determination of the electrical system equations (1) and (2), as well as for other relevant machine performance characteristics, such as electromagnetic torque (11). Tem = 3 2 \u00b7 p \u00b7 (\u03a8d \u00b7 Iq \u2212\u03a8d \u00b7 Iq) (11) Next, different algebraic approaches are described in order to obtain the d-q axes magnetic flux linkage. In this section, the procedure to obtain the magnetic flux linkage employing a magnet pole lumped parameter magnetic circuit is described. The employed magnetic circuit, widely explained in [17], allows the determination of machine magnetic features at load condition. The load magnetic circuit, Fig. 3, arises from no-load magnetic circuit and armature current magnetic circuit, and is capable of predicting the air gap magnetic flux decrease produced by the cross-magnetizing effect. When the machine is fed with only q-component armature current, the resulting air gap magnetic flux density along a pole arc is shown in Fig. 4. Machine saturation causes an increase of the magnetic flux density on the over-magnetizing side of the d-axis that extend beyond the ferromagnetic material saturation knee", + " w eq corresponds to the region from the d-axis to 3, which fulfils the following condition: the air gap flux contribution from point 3 to 4, \u03c6shaded, must be the same as the flux contribution from 1 to 2. The range between point 2 and the \u03c3axis (the overmagnetized side of the local region) is defined as the \u201dplus side\u201d, Fig. 5. Similarly, the minus side covers the range from the \u03c3axis to 3, where the stator current demagnetizes the local region under study. The region between points 2 and 3 is modelled by the magnetic circuit, Fig. 3. The different components of the magnetic circuit, appear with superscript + or - depending on the corresponding local region they are modelling to, i.e., the plus or minus side. The initial position of the \u03c3axis, \u03c3initial, is defined as the point between 2 and 3, which magnetic flux at plus and minus sides have the same value without considering the magnetic saturation. Once the \u03c3axis is identified, the different lumped parameters of the local magnetic circuit, Fig. 3, can be determined. The magneto-motive force sources due to the stator currents are calculated as follows: F + I = 1(\u03c4p 2 \u00b7 \u03b1 PM \u2212 \u03c3 ) \u03c4p 2 \u00b7\u03b1 PM \u2212\u03b3\u222b \u03c3\u2212\u03b3 F I \u00b7 sin ( \u03c0 \u03c4p \u00b7 x ) \u00b7 dx (12) F \u2212 I = \u22121 (weq \u2212 \u03c3) weq\u2212\u03c3\u222b \u2212(\u03c3\u2212\u03b3) F I \u00b7 sin ( \u03c0 \u03c4p \u00b7 x ) \u00b7 dx (13) FPM = hPM \u00b7HC (14) Where: \u2022 F I : Armature current magneto-motive force. \u2022 \u03b3: Armature current argument (d-q reference system). \u2022 hPM : Magnet height. \u2022 \u03c4p: Polar pitch. \u2022 \u03b1 PM : Magnet polar relative arc. In addition, for the \u03c3axis position the different reluctances are calculated by equations: R+ agap = hagap,eff \u03bc0 \u00b7 (w PM 2 \u2212 \u03c3 ) \u00b7 LFe (15) R\u2212 agap = hagap,eff \u03bc0 \u00b7 (weq \u2212 \u03c3) \u00b7 LFe (16) R+ PM = hPM \u03bc PM \u00b7 (w PM 2 \u2212 \u03c3 ) \u00b7 LFe (17) R\u2212 PM = hPM \u03bc PM \u00b7 (weq \u2212 \u03c3) \u00b7 LFe (18) Where: \u2022 hagap,eff : Air gap effective height (including Carter\u2019s factor)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001820_s11432-013-5038-8-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001820_s11432-013-5038-8-Figure3-1.png", + "caption": "Figure 3 The building blocks of metamaterials to design the GRIN slab lenses. (a) A square-ring unit cell; (b) a dielectric block with drilling hole.", + "texts": [ + " SRRs) can realize large dynamic ranges of material parameters, including negative values, but suffer from narrow bands and large losses. On the contrary, non-resonant structures cannot achieve extreme values of material parameters, but can be operated in wideband with low loss. Considering that the index of refraction is gradually changed without extreme values in GRIN lenses, we choose non-resonant structures, a square ring and a drilling-hole dielectrics, as the building blocks for GRIN metamaterials, as shown in Figure 3 (a) and (b), respectively. In the square-ring resonator [9], the width of metal line is 0.16 mm with a thickness of 0.035 mm, which is printed on an FR4 substrate with the dielectric constant of 4.4 and loss tangent of 0.025. The central frequency is designed as 10 GHz, and the size of unit cell is set as 3\u00d73\u00d73 mm3. The retrieved normalized wave impedance and index of refraction of the effective medium are shown in Figure 4 (a)\u2013(d) when the length of square ring varies from 0.2 mm to 2.4 mm. We observe that the effective index of refraction changes from 1" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003536_978-94-009-5063-4_1-Figure29-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003536_978-94-009-5063-4_1-Figure29-1.png", + "caption": "Figure 29.", + "texts": [ + " As might be expected from this behavior, the critical load of a spherical shell subjected to uniform external hydrostatic pressure is highly sensitive to initial geometric imperfections. Spherical caps: The fact that an initial buckle subtends a small solid angle stimulated those initially interested in complete spherical shells to model the problem of buckling of a complete spherical shell with use of a shallow spherical cap clamped at its edge. Over the years the shallow cap configuration evolved into a \"classical\" problem in its own right, studied with almost the same intensity and frequency as the axially compressed cylinder. However, as demonstrated in Fig. 29, the shallow cap problem has certain characteristics not present in the case of a complete spherical shell. These arise from the presence of the edge. In Fig. 29 load-deflection curves are shown corresponding to linear and nonlinear theories for prebuckling axisymmetric deformations of caps clamped at the boundary. The open circles on the linear load-deflection lines indicate bifurcation buckling at the \"classical\" pressure for the complete spherical shell with the same radius-to thickness ratio as the spherical cap. The classical buckling stress is given by the same formula as that for the cylindrical shell subjected to axial compression, Eq. (10). A is a cap shallowness parameter given by (11) where H is the rise of the cap above the plane in which the edge lies, and h is the thickness. For A less than about 7 or 8 the behavior of the shallow cap little resembles that of the complete spherical shell. With A = 0 (flat circular plate) there is no similarity at all: The load-deflection curve exhibits a stiffening characteristic which results from the build-up of in-plane tension as the plate deforms (Fig. 29(a)). With A less than about 3.5 the load-deflection curve has no horizontal tangent and no bifurcation point so that there is no loss of stability on the primary equilibrium path(b). For A less than about 6 there is axisymmetric snap-through, but no bifurcation buckling(c). For A > 6 bifurcation buckling into a nonsymmetric mode occurs at a lower load than either axisymmetric snap-through of the cap or classical buckling of a complete spherical shell (d, e, f). Notice that as A increases above 7 the prebuckling behavior becomes more and more linear. Figure 29(f) corresponds to a configuration in which the cap is no longer \"shallow\" if that word may be used as a means of classifying structural behavior: The nonuniformity of pre buckling behavior occurs in a relatively narrow band or \"boundary layer\" near the edge. Any further increase in A results in no further alteration in the curves or locations of the bifurcation points presented in Fig. 29(f). No matter how high A is, the behavior of the incomplete spherical shell clamped at its boundary will never be the same as that of the complete spherical shell because the presence of the boundary gives rise to edge buckling at a pressure from 80% to 90% of the classical value Pel. For actual spherical shells and shallow caps random imperfections playa major role in the loss of stability under uniform external pressure. Figure 30 demonstrates that the effect of initial imperfections is just as severe as in the case of cylindrical shells subjected to axial compression" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000653_s00500-019-04233-7-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000653_s00500-019-04233-7-Figure5-1.png", + "caption": "Fig. 5 Laboratory helicopter (CE 150)", + "texts": [ + " From these depicted simulation results, it is clear that actual trajectories converge towards desired trajectories and that the closed-loop system stability is guaranteed in spite of the presence of the actuator nonlinearities (i.e. the input dead zone and saturation), as well as the uncertain dynamics. The control inputs are also admissible and bounded, i.e. limited by their physical lower and upper bounds. This proves the effectiveness of the proposed fuzzy adaptive variable-structure control scheme. Example 2 Consider a laboratory model of a helicopter (CE 150), see Fig. 5, where the elevation angle h and the azimuth angle / are controlled by the main rotor (u1) and the secondary rotor (u2). This helicopter system is a MIMO system with uncertain nonlinear dynamics (Humusoft 2002; Haddad and Boulkroune 2016). Its continuous-time simplified model can be given by: M /; h\u00f0 \u00de \u20ac/ \u20ach \u00fe C /; h; _/; _h _/ _h \u00fe G /; h\u00f0 \u00de \u00bc u \u00f087\u00de where M /; h\u00f0 \u00de \u00bc cos /Il2\u00f0 \u00de2 0 0 Il2 \" # ; G /; h\u00f0 \u00de \u00bc 0 mgIc cos h ; C /; h; _/; _h \u00bc cos h sin h _hIl2 cos h sin h _/Il2 cos h sin h _/Il2 0 \" # ; and s \u00bc u1 u2 : with Il2 \u00bc ml 3 l3 1 \u00fel3 2\u00f0 \u00de l1\u00fel2\u00f0 \u00de \u00fe m1l 2 1 \u00fe m2l 2 2, and lc \u00bc ml l1 l2\u00f0 \u00de\u00fe\u00f0 m1l1 m2l2\u00de=m" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001188_speedam.2012.6264408-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001188_speedam.2012.6264408-Figure6-1.png", + "caption": "Fig. 6. Basic machine design with embedded magnets", + "texts": [ + " It can be seen, that the cogging torque is reduced effectively by using the advanced magnet design compared to the sinu soidally shaped magnets. Furthermore, the mean value of the load torque and, therefore, the output power is increased. However, in this case the increase of the load torque is caused by an approximately equal increase of the magnet volume. IV. ApPLICATION EXAMPLE 2: BRUSHLESS PERMANENT Based on section II, the method is illustrated with another example envolving the basic design (see Fig. 6) of a machine with embedded magnets (for details refer to [4]). The rotor's iron contour facing the air gap is controlled by three parameters. These are, the radial height h f of the changed design and again, the magnitudes of the first and third harmonic al and a3. A detailed description of the arrangement is shown in Fig. 7, where the grey area represents the rotor iron contour facing the airgap excluding the saturation bar area. Although the contour cannot be changed on the whole rotor outline because of the saturation bar area, it can be seen in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001252_tmtt.1973.1128110-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001252_tmtt.1973.1128110-Figure8-1.png", + "caption": "Fig. 8. Circular loci of s22 in the complex plane for constant amplifier power gain with a given s, of 0.32 ~jO.48.", + "texts": [ + " S@ecij$cation ofs Parameters for Desired Amplifier Gain Having determined conditions on szzfor amplifier stability, we now wish to specify values for the s parameters which give the desired gain for the amplifier. By using the unitary properties of the the scattering parameters for the lossless two-port, we have I SU[ = [ S2ZI and [ SUI 2=1 \u2013 [ SZZI2, and from (8) we may write , so , = 1- S7\u2019S22* ST \u2014 S22 (14) where s,=s,(w, lAdl 2). For a constant Is.1 and a given value ofs, then (14) is the equation of a circle in the SM plane with center C at ~= 1s.1\u2019-1 \\salZ- 1s,12s\u201d and radius (15a) (15b) In Fig. 8 we have shown some constant IS= I circles for an arbitrary value of sr. Note that all the circle centers lie on a line through the origin and the point s, as can be seen from (15a). Thus any value of s,, on one of these circles will yield the same amplifier gain, although some of these values are not stable solutions. An amplifier with specified gain \\ S. (co) I can be realized if an SZ2(W) can be realized such that for each u over the band of interest szz lies on each corresponding ls~(~)] circle determined from (15) and s,(u)", + "280 in). This required a broad-band waveguide transformer on the circulator port which was built into the circulator, Under these conditions, a quarter-wavelength section of waveguide having a relative height h to the reference height (actual height/O.020 in) and the same waveguide width has s parameters given by h\u2019\u2013l \u201811= S2\u2019=W+l (16a) and 2h s12. \u2014j\u2014. h2+l (16b) We have neglected any discontinuity capacitance occurring at the junction of the different size waveguides. Since s22 is real, then from Fig. 8 it is clear these sections are used most effectively if s~ is real at the quarter-wave frequency fo. In practice s,(fo) and in particular s,O(~O) was made real by physically changing the reference plane (cutting of some waveguide), rotating s, to the real axis. PETERSON : MILLIMETER-WAVE IMPATT AMPLIFIERS 687 w 47\u2019 45 5\u2019 52 204-694 :-E256 t Predicted from theory. * Actual measured values. gains predicted for the available quarter-wave equalization networks. (b) Equalized characteristics of (a) obtained with the addition of the quarter-wave networks" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001293_s42417-018-0033-4-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001293_s42417-018-0033-4-Figure2-1.png", + "caption": "Fig. 2 AAR type H tight lock coupler [15]", + "texts": [ + ", Fn, Fbrake, and Frr. Methodologies to model coach connections with every single forcing inputs are included and discussed in the subsequent subsections. In a train, the adjacent vehicle is connected mechanically using a device known as coupler which plays an important role in the analysis of LTD [9\u201313]. This technique is easier to apply than the AAR and has no significant error consequence unless used for very sharp curves (less than 0.1% at R = 100\u00a0m). In this study, AAR type H tight lock coupler (Fig.\u00a02), with a low preload draft gear has been considered [14, 15]. The coupler system is a connection between two adjacent coaches in a train. It consists of coupler head with drawbar and its guide and draw and buffing gear in draft gear. The (1)m1x\u03081 \u2212 F ( y\u03071, y1, y\u03072, y2 ) = Ft\u2215bd \u2212 Fbrake,1(t) \u2212 Frr,1. (2)mix\u0308i \u2212 F ( y\u0307i, yi, y\u0307i\u22121, yi\u22121 ) = Fti/bd \u2212 Fbrake,i(t) \u2212 Frr,i. (3) mnx\u0308n \u2212 F ( y\u0307n, yn, y\u0307n\u22121, yn\u22121 ) = Ftn/bd \u2212 Fbrake,n(t) \u2212 Frr,n. coupler system is usually shortened into single-element models, so every two draft gears are exhibited in series as a single unit" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001673_pedes.2018.8707501-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001673_pedes.2018.8707501-Figure2-1.png", + "caption": "Fig. 2. Three-phase voltage source inverter (a) circuit configuration, (b) voltage vectors.", + "texts": [ + " 1), it is necessary to develop mathematical models of IM and power converter. These are described briefly in the following sub-sections. A. Mathematical model of IM and power converter The relevant equations depicting mathematical representation of IM in state space are given below [1] dt dRs s s isv += (1) r r r i er j dt dR \u03c9\u2212+=0 (2) rss ii ms LL += (3) rsr ii rm LL += (4) { }ss i ( ).Im 2 3 ConjpTe = (5) ( )le rm TT Jdt d \u2212= 1\u03c9 (6) Since three-phase, two-level voltage source inverter (VSI) as shown in Fig. 2 has been used, the available voltage vectors can be represented as dcVjj A).(SV 3 2)( = (7) where )( jS depicts the switching states: [0 0 0; 1 0 0; 1 1 0; 0 1 0; 0 1 1; 0 0 1; 1 0 1; 1 1 1], ( )21 aa=A and a is equal to 2 3 2 13 2 je j +\u2212= \u03c0 . B. Model predictive torque control scheme Using stator current ( si ) and rotor flux ( r ) as state variables, system model equation can be represented as [13] \u03c3\u03c3 \u03c3 \u03c9 \u03c4 \u03c4 R j R k dt d r r r s r s s v ii + \u2212=+ 1 (8) where \u03c3 \u03c3 \u03c3\u03c4 R Ls= , 2 rrs kRRR +=\u03c3 , r m r L Lk = , s ds s qs jvv +=sv , s ds s qs jii +=si and s dr s qr j\u03c8\u03c8 +=r are in the stationary ss qd \u2212 reference frame (see Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000439_2013.17935-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000439_2013.17935-Figure1-1.png", + "caption": "Figure 1. Side views of the PTO\u2212driven machines and front views of the active tools: (a) spading machine and (b) rotary chisel.", + "texts": [ + " The author is Fabio Pezzi, ASAE Member, Associate Professor, Department of Agricultural Economic and Engineering, University of Bologna, via Ravennate 1020, 47023 Cesena, Italy; phone: +39\u221205476\u221236114; fax: +39\u221205473\u221282348; e\u2212mail: fpezzi@agrsci. unibo.it. machine represents a technique already in existence for many years and has been used for high\u2212income crops or high moisture content soils. Newly designed equipment fitted to high\u2212power tractors was used with the spading machine in the trials. The two selected soil tillage implements, both PTO driven, were an alternatively actuated spading machine and a rotary chisel, manufactured by FALC (Faenza, Italy) (fig. 1). The semi\u2212mounted spading machine was equipped with ten tools operated by an articulated quadrilateral, with a swinging arm in the lower position. These tools were powered by the tractor PTO through a gearbox, which allowed two transmission ratios to be selected (0.133 and 0.173). The machine frame supported two fixed side protections and an adjustable rear projection for confining clod throwing. The implement was 3.12 m wide. Tillage depth was set by positioning the two trailing wheels. The semi\u2212mounted rotary chisel had a transverse rotor powered by the tractor PTO through a mechanical transmission having three different transmission ratios (0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003441_0954405011515299-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003441_0954405011515299-Figure7-1.png", + "caption": "Fig. 7 Model of the circuit in the second stage of the pressure boost", + "texts": [ + " Other constants for the simulation were determined on the basis of the experimental results shown in Fig. 6. The estimated values were \u00ae ne tuned on the basis of the best agreement of the simulated and experimental results. Initial conditions for the simulation were also obtained from the experimentally obtained data shown in Fig. 6. In the second part of the pressure boost stage, the pressure behind the hydromotor piston increases to the value of the accumulator pressure and the multiplicator is brought into action. The situation is shown in Fig. 7. The direct connection between the accumulator and the hydromotor cylinder is interrupted and the accumulator is connected to the multiplicator instead. As can be seen in the \u00ae gure, there are three masses and the situation is therefore a little more complicated. However, the mathematical model is again based on the equilibrium of forces acting on the concentrated masses and on the continuity of \u00af ow of a compressible \u00af uid. Proc Instn Mech Engrs Vol 215 Part B B09099 # IMechE 2001 at Universidad de Sevilla" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001244_tasc.2016.2543267-Figure9-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001244_tasc.2016.2543267-Figure9-1.png", + "caption": "Fig. 9. Iron loss distribution", + "texts": [ + " Magnetic flux density distribution according to the skew angle Numerical analysis on each model is conducted with a sinusoidal current source. Torque characteristics and axial forces at the ends of the rotor in both cases are shown in Fig. 10. Both models produce average torque at the level of the conventional skew model, regardless of the rotating direction. However, case 1 has a much higher torque ripple while case 2 has a slightly bigger ripple compared to the conventional skew model. The result of loss analysis based on FEA is shown in Fig. 9 and Table II. Core loss and copper loss in the V-skew models are higher compared to those of the conventional model; 4.8% of copper loss for both models, and 5.5% and 3.2% of core loss for cases 1 and 2 respectively. Axial force distribution according to the rotating direction is shown in Fig. 10(b). When V-skew is applied, the same amount of axial force is produced in the opposite direction, which makes the net axial force zero. However, case 1 produces higher axial force at each end, which makes the structure in case 1 less robust than case 2 due to deformation and damage on the lamination" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003616_6.2003-2171-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003616_6.2003-2171-Figure4-1.png", + "caption": "Figure 4 The Original Pack Shape", + "texts": [ + " Finger trapped loops were likely to be the only practical way to provide a join for both cords and a keeper was not likely to be viable for the rigging lines. Thus we opted to loop the lines directly around a suitably buffered loop of the Zylon strop. As regards the other end we had as yet no details of the PRM interface and thus could only assume that we would be able to create a suitably lightweight link between the Zylon and the PRM although it was clear that a metal link was likely to be too heavy. The volume available to stow the parachute was a particularly awkward shape. As shown by Figure 4 it is essentially a toroid with cut outs. The outer diameter is some 500 mm and its maximum depth is about 130 mm. The drogue mortar fits into the cylindrical centre cut-out and the wedge shaped cut-out through the whole of the toroid provides space for the mortar cartridges. Three further cut-outs are needed to allow for three rear cover stringers and a final half round one is needed to make space for the HEPA filter. Because of the large size of the parachute and the very long strop it was essential that we maximised the volume utilisation" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002847_lawp.2011.2171310-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002847_lawp.2011.2171310-Figure1-1.png", + "caption": "Fig. 1. Geometry of the proposed near-field edge-shorted slot microstrip an-", + "texts": [ + " Details of the antenna design are described in Section II, and a fabricated prototype of the proposed patch antenna is constructed and experimentally analyzed in Section III. Parametric study for understanding the antenna property will be performed and discussed as well. Experiments for reading dipole-like tags will be implemented by utilizing the proposed patch antenna with a commercial Sunlit 2.4-GHzRFID readermodule. Finally, this letter will be concluded with a brief summary in Section IV. Fig. 1 depicts the whole configuration with detailed design parameters of the proposed near-fieldmicrostrip antenna. It consists of a rectangular slot radiator, a shorting wall, and a pair of shorting strips. The radiator is implemented using a 0.4-mmthick FR4 substrate with dielectric constant and loss tangent . This patch antenna is fed by a probe, and its overall dimension is merely mm to be easily embedded inside an RFID handheld reader as an internal antenna. A ground plane with a suitable size of mm was employed to model a handheld reader device, which was made with a 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003984_aim.2009.5229822-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003984_aim.2009.5229822-Figure8-1.png", + "caption": "Fig. 8. Coordinate transformation of torque measurement", + "texts": [ + " 7, the high-precision standard sized blocks are utilized to support the rotor\u2019s flat bottom surface to secure the rotor\u2019s position. The screws beneath the rotor can be used to adjust the rotor\u2019s position up and down, and then be tightened to fix the rotor position. The standard blocks are removed after the adjustment to allow rotor motions. The counter weight on the rotor bottom is used to adjust the rotor\u2019s mass center to coincide with the bearing\u2019s rotation center. The raw data acquired from the six-axis force/torque sensor is based on the sensor coordinates as shown in Fig. 8; (xs, ys, zs) represents the sensor coordinate system whereas (xr, yr, zr) represents the rotor coordinate system. The measured data obtained from the sensor can be translated into the actual motor torque by Trx Try Trz = \u2212 sin \u03b4 cos \u03b4 0 cos \u03b4 sin \u03b4 0 0 0 \u22121 { Tsx Tsy Tsz + Fsy \u2212Fsx 0 Ls}, (5) where Fsx, Fsy , Fsz and Tsx, Tsy , Tsz are force and torque components based on the sensor coordinates, Trx, Try and Trz are torque components based on the rotor coordinates, Ls is the distance from the origin of the sensor coordinates to the rotor center" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000398_iecon.2010.5675174-Figure9-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000398_iecon.2010.5675174-Figure9-1.png", + "caption": "Fig. 9 \u2013 FEM Simulated field distribution ( \u00ba16=\u03b8 )", + "texts": [ + " const di 'dW \u03b8 \u03a8 = (9) Where di)i,( i 0 )i,('W \u03b8\u03a8\u03b8 \u222b= (10) It was assumed that computed magnetic quantities were constant for different machine sections, since the materials are considered isotropic and therefore allowing a twodimensional analysis. A number of geometries have been analysed, one for each rotor angular position from 0\u00ba to 24\u00ba in 4\u00ba increment steps. Windings are depicted by the rectangular areas adjacent to the stator poles. For a rotor position of 16\u00ba (mech. degrees) the field distribution is shown in Fig. 9. The FEM model and experimental magnetization curves of the 8/6 SRM are shown in Fig.10 for the selected rotor positions. In Fig. 11, model and FEM curves are represented and compared. Both curves show a good accuracy when compared with the experimental one; nevertheless the model curves are generally closer to the FEM magnetization curves as rotor position tends to the unaligned position (linear region). However in rotor positions revealing strong saturation, the model presents an acceptable accuracy" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000978_2004-01-1151-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000978_2004-01-1151-Figure2-1.png", + "caption": "Figure 2. Free Diagram of Clutch Barrel - Roller \u2013 Race", + "texts": [ + " Insufficient torque capacity results in the drive slippage while excessive high hoop stress on the clutch barrel ring causes barrel crack. To eliminate drive slippage failure, the clutch should be designed with high torque capacity. High torque capacity, however, is a cause of high hoop stress on the barrel that may result in the cracked barrel failure. The higher torque capacity and lower hoop stress are two completely opposite design directions. So, an optimal design approach should be applied to the clutch design. OPTIMALITY VERIFICATION Figure 2 shows the free diagram of the barrel-roller-race. The torque capacity of the clutch and the stresses on the race and barrel can be derived from the Figure 2. Since the torque is derived from the contact stress of the race by assuming the material property (compressive strength), the torque capacity (T ) of the clutch and the hoop stress ( h\u03c3 ) on the barrel are two design objectives, and can be simplified and expressed as[2], [3]:. \u03b1sin)(8300 2 Rr rR nlT + = (1) \u03b1 \u03c0 \u03c3 ctg ab ab alR T h )( 6 22 22 \u2212 += (2) Or replace Equation (1) into (2), then the hoop stress can be expressed as: \u03b1 \u03c0 \u03c3 cos)]( )( [49800 22 22 ab ab aRr rnR h \u2212 + + = (3) where R - radius of the race; ,a b - inside radius and outside radius of the barrel; r - radius of the roller; \u03b1 - strut angle; l - length of the roller; n - number of the rollers in a clutch" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003210_icectech.2011.5941668-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003210_icectech.2011.5941668-Figure1-1.png", + "caption": "Figure 1. The structure of PMSM", + "texts": [ + " In this way, all above efforts have investigated controller just in steady state condition and constant load. In this paper, a new controller with fuzzy logic has been proposed for estimation of stator resistance based on both current and stator flux errors. The resistance error observer designed with fuzzy logic is independent of the motor operating conditions. So this observer is very useful in DTC method. It minimizes the problems of stator resistance variation to a very large extent such as when the mechanical load applied. II. PMSM MODEL Fig. 1 shows the simplified 3-phase PMSM motor. Stator windings are shown as lumped winding, but they distributed sinusoidally around the stator. Stator currents build a rotating flux which makes the rotor follow it. The angle between flux linkage vector and rotor flux vector, load angle , produce electromagnetic torque. The motor voltage equation in the \"abc\" reference are given by [13]: (1) where Vabcs, iabcs and abcs are staror phase voltages, stator phase currents and stator flux linkages. The electromagnetic torque and mechanical speed are related by electromechanical motion equation (2) where J, m, B, Te and TL are the moment of inertia in [kg" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003215_6.1988-4153-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003215_6.1988-4153-Figure1-1.png", + "caption": "Fig. 1", + "texts": [], + "surrounding_texts": [ + "A large space antenna-like ground experiment structure has been developed for conducting research and validation of advanced control technology. A set of proof-of-concept adaptive control experiments for transient and initial deflection regulation with a small set of sensors and actuators were conducted. Very limited knowledge of the plant dynamics and its environment was used in the design of the adaptive controller so that performance could be demonstrated under conditions of gross underlying uncertainties. High performance has been observed under such stringent conditions. These experiments have established a baseline for future studies involving more complex hardware and environmental conditions, and utilizing additional sets of sensors and actuators.\nI. INTRODUCTION\nOver the past decade, the control and identification of large flexible structures has drawn intense interest and inspired extensive research activities. Many theoretical approaches to the control of flexible structures have been developed [I]. However, the experimental validation of control laws has progressed more slowly. Some representative ones are described in [2]-[12]. It is noted that most of these experiments were performed to study vibration damping and slewing of flexible beams or plates, and very few have been performed on structures having the size, complexity, and flexibility resembling those of future space missions.\nDevelopmental research on the application of adaptive control to large flexible structures has been evolving in recent years. Adaptive techniques were developed by Sundararajan et. al, based on lattice filter parameter identification followed by the design of modal controllers [13]. This approach is more correctly denoted 'adaptive calibration' since the tasks of identifying the plant and designing the control law are not performed simultaneously, and stability checks are required between control designs. Adaptive\ncalibration has been studied in a more general context by Kosut [I41 who considers an equation error approach to parameter identification followed by an output error characterization of the additive uncertainty reauired for robust control design. In a context more consistent with mainstream adaptive control, the Direct Model Reference Adaptive Control algorithm based on the command generator tracker (CGT) theory has been studied by a number of researchers 1151-[27]. This algorithm has been applied in various forms to several important problem areas including Space Station vibration suppression and payload articulation r18]-[21], flexible robot arm control 1221, and the NASA SCOLE experiment (251 --- in addition to oeing extended rigorously to more general distributed parameter systems in abstract Hilbert space settings [17][27]. The DMRAC algoritnm will be implemented in this experimental study.\nPast experimental work demonstrating the appiication of adaptive control to large flexible structures has been extremely limited. The few studies which are available include the work of Eldred and Schaechter on the Jet Propulsion Laboratory (JPL) Flexible Beam Facility [ 2 8 ] using an Sxtended Kalman Filter followed by a discrete LQ control design; Sundararajan, Montgomery and Williams on the NASA Langley free-free beam and 2D flexible grid [13] using lattice filter identification followed by a modal control design; the work of Sidman on adaptive pole-placement for the Stanford Four Disk System [29], and Rovner and Franklin using a modified Self-Tuning LQ design [301 on a flexible one-iink robotic manipulator at Stanford. These investigations utilize adaptive algorithms which are essentially of the indirect type, i.e., involving online parameter identification followed by controller design. The main restriction for algorithms of the indirect type is the requirement for persistence of excitation to guarantee parameter convergence. In contrast, this study utilizes the DMRAC algorithm which is of the direct type, i.e., the control gains are tuned directly to minimize the observed tracking error. The advantages of this adaptive method\ncopyright@1988 by the American Institute of Aeronautics and Astronautics, Inc.\nAll Rights Reserved. 832", + "include a reduced order model following capability, robustness to spillover dynamics, multivariable formulation, and convergence without persistence of excitation. The main restriction for this algorithm is the requirement that the actuators and sensors be colocated.\nFor the purpose of developing and validating advanced control technology, an antenna-like structure was developed at JPL. This structure was built to exhibit many essential properties of large flexible space structures, including multiple modes, low modal frequencies and low structural damping. A 3-D antenna-like structure was selected so that complex interactions among the structural components and the controi system could be investigated. One of the experimental investigation areas initiated at this facility is the Direct Model Reference Adaptive Control technology currently being developed at JPL [18]-[21]. A series of experiments has been planned and will be conducted in successive stages with increasing complexity. The objective of these experiments is to validate the control algorithms and closed-loop performance under conditions of gross structural and environmental uncertainties, time varying effects, actuator dynamics, time delays, nonlinearities, and system reconfiguration.\nThis paper describes the first two experiments in this series, i.e., the transient and initial deflection regulation problems. The adaptive controller was designed with greatly reduced knowledge of the dynamical system and its environment. The results have been encouraging, demonstrating robust performance and a high rate of convergence.\nThe organization of this paper is as follows. In Section 11, the experiment facility design is presented. The system dynamic model is described in Section 111. In Section IV, the adaptive control algorithm employed here is presented. The adaptive control experiment design is stated in Section V. In Section VI, results of the experiment are discussed. Conclusions are summarized in Section VII.\n11. EXPERIMENT FACILITY DESIGN\n2.1 Background\nPrior to the experiment structure design, a series of planning activities were conducted. Fundamental questions were addressed, e.g., what technologies should be validated? At what levels and complexity? How can these be achieved? and with what sequence of stages? A number of decisions were made including a three-year four-phase experiment plan that covers the validation of the following technology areas: static shape determination and con-\ntrol, dynamic parameter identification, active vibration damping and control, tracking, pointing and reconfiguration controi with both robust control and adaptive control techniques. The experiment facility itself was required to be a 3-D complex structure exhibiting characteristics and properties of realistic large flexible space structures. An antenna-like structure was conceived for this purpose.\n2.2 Configuration\nThe antenna-like structure selected for the experiment is shown in Figs. 1 and 2. The main component of the apparatus consists of a central hub to which 12 ribs are attached. The diameter of the dish structure is 18.5 feet, the large size being necessary to achieve the low modal frequencies desired. The ribs are coupled together by two rings of pretensioned wires. Functionally, the wires provide coupling of motion in the circumferential direction which cannot be provided by the hub. The ribs, being quite flexible and unable to support their own weight without excessive droop, are each supported at two locations along their free length by levitators. Each levitator assembly consists of a pulley, a counterweight, and a wire attached to the counterweight which passes over the pulley and attaches to the rib. The hub is mounted to the backup structure through a gimbal platform, so that it is free to rotate about two perpendicular axes in the horizontal plane. A flexible boom is attached to the hub and hangs below it, and a feed mass, simulating the feed horn of an antenna, is attached at the free end of the boom. The original boom was 12 feet long, but the desirability of conducting experiments at the ground level so that the", + "Fig. 2 Experiment structure\nstructure was accessible resulted in a second, 3-foot long boom being used for the initial set of experiments instead.\n2.3 Actuators\nEach rib can be individually excited or controlled by a rib-root actuator. Each rib-root actuator has a solenoid design which reacts against a mount that is rigidly attached to the hub. In addition, two hub actuators are provided to torque the hub about its two gimbal axes. The hub torquers do not provide torque directly, but rather are linear force actuators which provide torque to the hub by pushing at its outer circumference. The torque provided is equal to the force times the lever arm about the axis of rotation. The placement of these actuators guarantees good controllability of all of the flexible modes of motion. The locations of the actuators are shown in Fig. 3.\n2.4 Sensors\nThe sensor locations are also shown in Fig.3. First, each of the 24 levitators is equipped with an incremental optical encoder which measures the relative angle of the levitator pulley. The levitator sensors thus provide, in an indirect manner, the measurement of the vertical\nmotion of the corresponding ribs at the points where the levitators are attached. There are also four evenly spaced linear variable differential transformers ILVDT) rib-root sensors colocated with four ribroot actuators. The hub angular positions are measured by two rotary variable differential transformers (RVDT) mounted directly at the gimbal bearings.\n2.5 Computer System\nThe computer system selected for the experiment facility is the DEC VAX station I1 workstation consisting of a 71MB hard disk drive, 5 MB of main memory, dual 819.2 KB floppy disk drives, a 9 5 MB cartridge tape subsystem, a 1K by 2K bit-mapped video graphics monitor, and a dot matrix printer. The system software was written in Ada, while the experiment software was written in FORTRAN. Through the data acquisition system, the computer samples the various sensors outputs, uses these data to update adaptive control gains and compute appropriate actuation commands, and then sends the signals to the actuators. This cycle is repeated until the control objectives are achieved.\n2.6 Data Acquisition System\nThe Data Acquisition System (DAS) is a" + ] + }, + { + "image_filename": "designv6_24_0002862_detc2007-34856-Figure11-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002862_detc2007-34856-Figure11-1.png", + "caption": "Figure 11: John Deere IVT [13]", + "texts": [ + " To overcome the non-integer tooth problem, the inventor proposes a method of allowing the sprocket bars to adjust their position. As shown in Fig. 10, the sprocket bars on the cones Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 02/04/2016 T are allowed to float, meaning that they can move radially and adjust a finite amount along the circumference of the cone. Because of the ability to float, the distance between the sprocket bars can adjust to accommodate a chain of fixed pitch. The hydrostatic-mechanical transmission, developed by John Deere [12] (see Fig. 11), includes a hydrostatic unit and a mechanical differential. In this embodiment, the torque supplied by the power input is divided between the hydrostatic branch and the mechanical branch, and then rotational speed and torque are converted and recombined in the mechanical differential (see Fig. 12). The mechanical differential sums the inputs from the mechanical and hydrostatic branches and provides a resulting output. While the input to the mechanical differential from the mechanical branch is constant (\u03c9in), the hydrostatic unit has the ability to provide a continuously varying output, which when summed with the constant input from the engine in the differential, allows the final output of the differential to also be continuously varying" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001385_iccad.2000.896524-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001385_iccad.2000.896524-Figure5-1.png", + "caption": "Figure 5 : Scheduling Conflict-free Rules: (a) Conflict-free graph (b) Corresponding conflict graph and its connected components", + "texts": [ + ", $T~,, can be generated from ~ C T , ~ (s),...,n~,( s) using a priority encoder. In the best case, a transition T is conflict-free with every other transition in X. Hence, Tis in a scheduling group by itself, and $ T = ~ T without arbitration. X can be partitioned into scheduling groups by finding the connected components of an undirected graph whose nodes are transitions T I , ..., TM and whose edges are {(Ti, Tj) 1 -(IT;. <>CF Tj ) } . Each connected component is a scheduling group. For example, the undirected graph (a) in Figure 5 depicts the <>CF relationships in an ATS with six transitions. Graph (b) in Figure 5 gives the corresponding conflict graph where two nodes are connected if they are not <>cF, i.e. two unconnected nodes 7;: and r j imply 7;: <>CF Tj . The conflict graph has three connected components, corresponding to the three < >CF scheduling groups. The $ signals corresponding to T I , T4 and T, can be generated using a priority encoding of their corresponding n\u2019s. Scheduling group 2 also requires a scheduler to ensure $2 and $5 are not asserted in the same clock cycle. However, $ ~ ~ = n q without any arbitration. Enumerated Scheduler: Scheduling group 1 in Figure 5 contains three transitions { T I , T4, T6} such that TI <>CF T6 but neither TI nor is <>CF with T4. Although the three transitions cannot be scheduled independently of each other, TI and T6 can be selected together as long as T4 is not selected in the same clock cycle. This selection is valid because TI and T6 are <>CF between themselves and every transition selected by the other groups. In general, the scheduler for each group can independently select multiple transitions that are pairwise <>CF within the scheduling group. For a scheduling group with transitions Txl ,..., TXn, $T-~ ,...,$T~,, can be computed by a 2\u201d x n lookup table indexed by n~ (s) ,..., n~,(s). The data value dl ,..., d, at the table entry with index b1, ..., b,, can be determined by finding a clique in an undirected graph whose nodes N and edges E are defined as follows: :\u2019 = {zi 1 bi is asserted} t\u20ac = {(Txp TXj> I (TXiEN) A (TXj\u20acN) A (?& <>CF zj)} For each Gi that is in the clique, assert di. For example, scheduling group 1 from Figure 5 can be scheduled by an enumerated encoder (Figure 6 ) that allows TI and T6 to execute concurrently. The construction of an enumerated encoder is not necessarily unique. For example, in this example, row \u201c01 1\u201d in Figure 6 could also contain the data value \u201c001\u201d. 4.4 Performance Gain When X can be partitioned into scheduling groups, the partitioned scheduler is smaller and faster than the monolithic encoder used in the reference implementation. The partitioned scheduler also reduces wiring cost and delay since n\u2019s and 9\u2019s of unrelated transitions are not brought together for arbitration" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000885_j.apm.2014.04.032-Figure10-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000885_j.apm.2014.04.032-Figure10-1.png", + "caption": "Fig. 10. The conditions on the steering tie-rod disposing.", + "texts": [ + " Thus, for the wheel carrier, if a1, a2 and a3 are the position angles of the local axis x2, b1, b2, b3 \u2013 the position angles of the local axis y2, and c1, c2, c3 \u2013 the position angles of the local axis z2, all of them relative to the global axes (x,y,z), the transformation matrix from \u2018\u20182\u2019\u2019 to \u2018\u20180\u2019\u2019 has the well-known form: M2 0 \u00bc cos a1 cos b1 cos c1 cos a2 cos b2 cos c2 cos a3 cos b3 cos c3 2 64 3 75: \u00f023\u00de The director cosines will result based on the differences of global coordinates (in the fixed reference frame Oxyz), in the following way: \u2013 the local axis y2 is on the line GF, the director cosines being obtained from the global coordinates of the points G and F: cos b1 \u00bc xG xF GF ; cos b2 \u00bc yG yF GF ; cos b3 \u00bc zG zF GF ; \u00f024\u00de \u2013 the local axis x2 is normal to the plane (GFM), being normal to GF and FM, and collinear with v: v \u00bc GF FM \u00bc i j k xG xF yG yF zG zF xM xF yM yF zM zF \u00bc vx i\u00fe vy j\u00fe vz k; \u00f025\u00de ) cos a1 \u00bc vx=v ; cos a2 \u00bc vy=v ; cos a3 \u00bc vz=v ; v \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi v2 x \u00fe v2 y \u00fe v2 z q ; \u00f026\u00de \u2013 the local axis z2 is normal to the plane (x2y2), being normal to GF and v , and collinear with w, w \u00bc GF v \u00bc i j k xG xF yG yF zG zF vx vy vz \u00bc wx i\u00fewy j\u00fewz k; \u00f027\u00de ) cos c1 \u00bc wx=w; cos c2 \u00bc wy=w; cos c3 \u00bc wz=w; w \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi w2 x \u00few2 y \u00few2 z q : \u00f028\u00de For the steering tie-rod there are necessary supplementary conditions concerning the orientation around its own axis CE. This involves the existence of a bolt (pin) in the socket of the spherical joint (E) between the tie-rod (4) and the pitman arm (5). From theoretical and practical point of view, this joint can be materialized by a bolt with bushing, which is equivalent with a spherical joint with bolt (pin). In this way, the tie-rod plane (CC0E) can be identified, the local reference frame being C0x4y4z4 (Fig. 10a). cite this article in press as: P. Alexandru et al., Modeling the angular capability of the ball joints in a complex mechanism with two s of mobility, Appl. Math. Modell. (2014), http://dx.doi.org/10.1016/j.apm.2014.04.032 The local axis x4 is the bolt axis, while y4 is oriented towards the normal axis from C on the bolt direction (in C0). In accordance with Fig. 10a, the existence of a bolt that belongs to the tie-rod, in a socket belonging to the pitman arm, allows the relative rotations around the local axes x4 and z4 (the spherical joint with bolt having two degrees of mobility). In this case, a new geometric parameter will appear, xE\u00f0C0\u00de. The local axis x4 of the tie-rod plane (x4y4) is disposed in the fixed plane (xy), being positioned by the angle a4 relative to x, the socket from the pitman arm belonging to (xy) because z5//z. The local axis y4 is positioned relative to the plane (xy) by the angle b4/b04, where b04 = arctg (tg b4 cosa4) corresponds to the rotation around x4", + " Under these terms, the rotation/orientation matrices have the form: Please degree Ma4 4 0 \u00bc cos a4 sin a4 0 sina4 cos a4 0 0 0 1 2 64 3 75; M b04 4 0 \u00bc 1 0 0 0 cos b04 sin b04 0 sin b04 cos b04 2 64 3 75: \u00f029\u00de Based on the relationships between coordinates, xC xE yC yE zC zE 2 64 3 75 \u00bc Ma4 4 0 M b04 4 0 h i xC\u00f0E\u00de yC\u00f0E\u00de 0 2 64 3 75; xC\u00f0E\u00de \u00bc C0E; yC\u00f0E\u00de \u00bc C0C; \u00f030\u00de there are obtained the positioning angles a4 and b4/b04 in current position. The disposing of the steering tie-rod relative to the nut rod in C (Fig. 10b) can be imposed so that the unit vector uC of the nut is positioned in the plane ( ns4 , CE), the projection of the vector on this plane being along the characteristic line dc4. In the sequence of angles used to pass from the casing plane \u2018\u2018S1/3/4\u2019\u2019 of the suspension arms (g01/g03, l1/l3, u1/u3, t001/t003 and w01/w03) and of the steering tie-rod (g04, l4, b04 and a4) to the fixed reference frame \u2018\u20180\u2019\u2019 (attached to car body), as well as from the nut to the wheel carrier (k0M;N;C and eM,N,C), in accordance with the previous figures, the corresponding matrices will be: MA\u00f0x\u00de the angles=matrices Ml1 S1 1; M/1 1 0; Ml3 S3 3; M/3 3 0; Ml4 S4 4; M b04 4 0; MeM;N;C p 2 ; \u00f031\u00de MA\u00f0y\u00de the angles=matrices M g01 S1 1; M t001 1 0; M g03 S3 3; M t003 3 0; M g04 S4 4; M k0M;N;C p 2 ; \u00f032\u00de MA\u00f0z\u00de the angles=matrices Mw01 1 0; Mw03 3 0; Ma4 4 0: \u00f033\u00de Eqs", + " 3): \u00f0ns4 \u00de4 \u00bc \u00bdMs4 4 \u00f0ns4\u00des4 \u00bc Ml4 s4 4 h i M g04 s4 4 h i 0 0 1 2 64 3 75 \u00bc ux4 s4 ;uy4 s4 ;uz4 s4 : \u00f040\u00de The unit vector uC==zC of the rod uses the rotations eC, k0C; ai, bi, ci (i = 1,2,3); a4, b04 for bringing the system in x4y4z4 through the kinematic loop 2-0-4, \u00f0uC\u00de4 \u00bc \u00bdMp 4 \u00f0uC\u00dep; \u00f0uC\u00dep \u00bc \u00f00;0;1\u00de; resulting the following expression: \u00f0uC\u00de4 \u00bc M b04 0 4 h i Ma4 0 4 \u00bdM2 0 MeC p 2 h i M k0C p 2 h i 0 0 1 2 64 3 75 \u00bc ux4 C ;u y4 C ;u z4 C \u00f041\u00de the positioning angles a4 and b04 for the spherical joint with bolt E (see Fig. 10a) being computed with Eqs. (30). To express the unit vector uC in the casing plane (xs4ys4zs4), for calculating the angle sC (21), the angles l4 and g04 will be used in addition to (41) (see Fig. 3): \u00f0uC\u00des4 \u00bc M g04 4 S4 h i Ml4 4 S4 h i \u00bdMp 4 0 0 1 2 64 3 75 \u00bc uxs4 C ;uys4 C ;uzs4 C ; \u00f042\u00de all these angles being considered around the axes specified in figures, and in accordance with the specifications in Eqs. (31)\u2013(33). The numerical simulations for the guiding (suspension and steering) system in study (see Figs" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002148_kem.620.318-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002148_kem.620.318-Figure2-1.png", + "caption": "Fig. 2 Theoretic mechanical model", + "texts": [ + " If the brake shoe and the wheel are concentric and the friction that chute wear plate acts on the slide wear sleeve is ignored during the braking process, then the braking force F1 acted by the floating lever is on the brake beam of the floating lever end, the normal pressure F2 and the friction F3 (F2 and F3 are not applied on the same point but distributed along the brake shoe cambered surface, while, to facilitate analysis, simply regard them as two concentrated force) acted by the wheel are on the brake shoe and the supporting force F4 acted by the chute is on the sliding block. The braking mechanical models are shown in Fig. 2 for wheel revolving forward and backward. From the braking mechanical model of the braking shoe, we can figure out that the brake shoe, under the friction F3 and the supporting force F4, tends to revolve counterclockwise when the wheel is forward revolving, which results that the normal pressure F2 is close to the upper part of the brake shoe, so the pressure on the upper part is larger than the lower part', and the upper eccentric wear will happen. In a similar way, when the wheel is backward revolving, the brake shoe tends to revolve clockwise, which results lower eccentric wear" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002971_robot.2001.932539-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002971_robot.2001.932539-Figure3-1.png", + "caption": "Figure 3. Free body diagram of an object on a device surface with one nozzle actuation", + "texts": [ + " In addition, instead of applying a complex friction model, the simple Coulomb friction model was applied because the detailed friction mechanism is still unknown and the purpose of this study is to drive the object manipulation method with the applied friction force field. 4 Dyanmics of Manipulation 4.1 Sliding Mode When there are two interface surfaces on an object, sliding may occur on either one or both surfaces dcpending on the friction forces applied on thc object. In this manipulation method, an object can slide against the device surface, nozzle or both surfaces. Figure 3 shows a free body diagram of an object on the device surface with one nozzle actuation. During the actuation of the nozzle with the force F, a normal force In is applied on the object generating a friction force J; on the objcct. If f is a friction forcc against the motion Case 1: F <.:;, J; and J; <.:;, /, There is no sliding on either surface. Case 2: F <.:;, J; and J; > /, There is sliding between the object and the device surface. Case 3: F > J; and J; <.:;, J; There is sliding between the object and the nozzle" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002286_j.1525-1594.1995.tb02397.x-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002286_j.1525-1594.1995.tb02397.x-Figure1-1.png", + "caption": "FIG. 1. Computational fluid dynamic analysis shows the surface pressure distribution on the impeller. The point of highest negative pressure is seen at the leading edge of the long impeller blade.", + "texts": [ + " At this time, we are fabricating internal pump parts from polycarbonate using a computer numerically controlled four-axis milling machine and evaluating pump designs, bearing designs, motor designs, pump material, and control systems, 653 654 K . MIZUGUCHI ET A L . Design of a flow inducer-lirnpeller and in vitro evaluation Initial experimental pump models were made of polyether polyurethane using the stereolithography process ( 3 ) . Computational fluid dynamic (CFD) analysis of this model was performed by NASA/ Ames Research Center. Figure 1 shows the results as indicated by the surface pressure distribution on the impeller. This analysis indicated regions of relatively high negative pressure at the leading edge of the impeller. High negative pressure in this region is suspected to be a major cause of hemolysis. As a result of this study, Ames Research Center suggested use of a flow inducer, a separate pumping section spinning in front of the impeller, which prerotates the fluid before i t enters the impeller. The initial flow inducer consisted of 3 blades in front of the impeller, which formed a separate discrete pumping section (i", + " In vitro hemolysis test result of the polycarbonate pump The polycarbonate pump required a 4-h hemolysis test due to the negligible amount of hemolysis it induced. The measured NIH was 0.0029 f 0.0009 g/100 L (n = 4). Figure 10 shows a statistically significant difference between the polycarbonate and the polyether polyurethane model (NIH = 0.018 * 0.007 gil00 L; n = 3). In vivo evaluation of the polycarbonate pump One polycarbonate pump ran continuously in vivo for 8.5 days (203 hf with sufficient flow (3.6- 4.7 L/min) (Fig. 1 1 ) against 80-100 mm Hg mean arterial pressure. Pump rpm was stable at approxi- ArtifOrgans. Vol. 19. N o . 7. 1995 658 K. MlZUGUCHl ET AL. mately 11,000 during this period. The plasma free hemoglobin level remained below 8 mg/dl , typically between 2 and 3 mg/dl, with a hematocrit of 20%. This in vivo experiment was terminated before the pump stopped due to decreased pump flow and rpm. Upon removal and disassembly, thrombus formation was found at the front and the top of the flow straightener blades" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002430_6.2020-3795-Figure22-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002430_6.2020-3795-Figure22-1.png", + "caption": "Fig. 22 Integrated PTS at the servicing payload level.", + "texts": [ + " All of the verification events are documented and auditable in the configuration managed TDMS and tracked in DOORS. Good progress was indicated in the area of payload level integration and testing documentation. Figure 21 presents the PTS integration and test flow through servicing payload handover to SV. Fig. 21 PTS integration and test flow through servicing payload handover to SV. During the PTS servicing payload integration, the PTS assemblies will be welded together and tested. Weld head and x-ray device clearances were checked and verified to the designed manifold standoff height. Figure 22 shows the fully assembled PTS as it would mount to the servicing payload with the payload structure (hidden for clarity). The VTSA is also shown with the protective cover planned to be used during the integration and test campaign to protect it from accidental damage. D ow nl oa de d by C A R L E T O N U N IV E R SI T Y o n A ug us t 2 2, 2 02 0 | h ttp :// ar c. ai aa .o rg | D O I: 1 0. 25 14 /6 .2 02 0- 37 95 Interconnecting lines structural analysis was carried out, and the acceptance testing was documented including the procedure for the electrical mate to flight avionics" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002508_j.hcl.2013.02.002-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002508_j.hcl.2013.02.002-Figure6-1.png", + "caption": "Fig. 6. (A) Percentage contribution of different factors to resistance to tendon motion. (B) Structures causing resistance to tendon gliding. Note that some factors can be eliminated during surgery: restrictive pulleys constitute about 30% of the total resistance, so venting of the constrictive portions of pulleys would decrease the resistance. Influences from other factors, such as extensor tethering or joint stiffness (contributing 30%\u201335% of total resistance), can be reduced by repeated passive digital motion before active tendon motion.", + "texts": [ + " In a chicken model, we sequentially removed the volar subcutaneous tissue, extensor tendons, and flexor pulleys to try to understand the relative contribution of each of these surrounding tissues to tendon gliding resistance. Constrictive pulleys accounted for about 30% of all the resistance to digital flexion when the FDP tendon was injured; the subcutaneous tissue 10% to 25% (about 10% in normal toe and 20%\u201325% in injured FDP tendon); the FDS tendon about 10%; and the remaining portion (35%) should come from the combined effects of extensor tethering, joint stiffness, and mass of the digit (Fig. 6). Finally, we must emphasize that finger joints are inherently small, and development of stiffness is frequent. At any phase along the flexion arc, if stiffness of these joints presents an obstacle to active tendon movement, the entire scale of contributing factors would be changed. Joint stiffness would become the overwhelmingly major resistance to tendon gliding. Each factor\u2019s contribution to the resistance of tendon gliding is a dynamic process, which varies depending on joint position, relative position of the tendon repair site to the pulleys, and preconditioning of the tendon and adhesion tissues" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002755_tvlsi.2012.2227848-Figure21-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002755_tvlsi.2012.2227848-Figure21-1.png", + "caption": "Fig. 21. BEOL metal stack from the 22-nm 6T SRAM (111) bitcell. (a) Without lithography effects and (b) with lithography effects. Dielectric regions are not shown.", + "texts": [ + " 4) Modeling Lithographic Effects: The structures synthesized above do not take into account the lithographic rounding effects on printed features and, hence, our earlier setups are likely to overestimate parasitic capacitances. In order to quantify the latter, we performed simple experiments where feature rounding was introduced into the BEOL metal layer by layer (with 10- and 6-nm radii of curvature for metal and vias, respectively) during structure synthesis, to realistically model printed via/metal shapes. Fig. 21(a) and (b) shows the 22-nm bulk 6T SRAM (111) BEOL metal stack without and with lithographic corner rounding. From Fig. 22(a) and (b), we can see that there is only a 3%\u20134% overestimation error, which suggests that lithographic effects can be ignored without too much loss in accuracy. In this section, we highlight the need to back-annotate 3-D-TCAD-extracted parasitic capacitances into mixed-mode transient simulations, and compare the relative importance of modeling device transport versus device parasitics" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003552_icorr.2007.4428441-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003552_icorr.2007.4428441-Figure4-1.png", + "caption": "Fig. 4. Mechanical design of the prosthesis.", + "texts": [ + " * the system must be capable of controlling joint position during the swing phase. * the prosthesis must provide sufficient shock tolerance to prevent any damage in the mechanism during the heelstrike. The corresponding parameters values of the above design goals are given in Table I. These parameters values are estimated based on the human data from [5][6][20]. The basic architecture of our mechanical design is a physical spring, configured in parallel to a high power output forcecontrollable actuator (Fig. 4). The parallel spring and the force-controllable actuator serve as the spring component and the torque source in Fig. 3(B), respectively. The prosthetic ankle-foot system requires a high mechanical power output as well as a large peak torque. The parallel spring shares the payload with the force-controllable actuator, thus the required peak force from the actuator system is significantly reduced. Consequently, a smaller transmission ratio can be used, and a larger force bandwidth is obtained. The series elasticity is also an important design feature for the ankle-foot prosthesis as it can prevent damage to the transmission due to shock loads during heel-strike. As can be seen in Fig. 4(a), there are five main mechanical elements in the system: a high power output d.c. motor, a transmission, a series spring, a unidirectional parallel spring, and a carbon composite leaf spring prosthetic foot. The first three components are combined to form a force-controllable actuator, called Series-Elastic Actuator(SEA). A SEA, previously developed for legged robots [21], consists of a dc motor in series with a spring (or spring structure) via a mechanical transmission. The SEA provides force control by controlling the extent to which the series spring is compressed", + " The elastic leaf spring foot is used to emulate the function of a human foot that provides shock absorption during foot strike, energy storage during the early stance period, and energy return in the late stance period. A standard prosthetic foot, Flex Foot LP Vari-Flex [14] is used in the prototype. A. Component Selections Broadly speaking, there are three main design decisions in this project: (1) choosing the parallel spring stiffness, (2) choosing the actuator and transmission, and (3) choosing the series spring stiffness. 1) Parallel Spring: A linear parallel spring kp with a moment arm Rp in Fig. 4(c) provides a rotational joint stiffness Kp> Kp = (kp)(Rp) (1) The goal is to properly select the moment arm and the spring constant in order to provide the suggested offset stiffness in Table I. In the physical system, due to the size and weight constraints, kp and Rp were chosen to be 770KN/m and 0.022m, respectively. Consequently, K4=385rad/s. Because this value is smaller than the suggested offset stiffness(550rad/s), the SEA supplements the required joint stiffness (see Fig. 5). 2) Actuator and Transmission: The goal is to select an actuator and a transmission to bracket the maximum torque and speed characteristics of the prosthesis, so as to match the intact ankle torque/power-speed requirements (Fig", + " Therefore, by choosing a stiffer spring, our design goal was to have the large force bandwidth of the SEA much greater than the required force bandwidth in the specifications (Table I) . To analyze the large force bandwidth, we proposed a simple linear model (Fig. 7) for the prosthesis based on [21]. All system parameters and variables were converted to the linear motion of the ballscrew in the prosthesis. We define a transmission ratio R that converts rotary motion of the motor into linear compression on the series spring(See Fig. 4(c)). The effective motor massMe, dampingBe, and linear motor forceFe can be obtained using the following equations: Me = ImR2, Fe = TmR, Be = bmR, where Im, Tm, bm are the rotary motor inertia, motor torque, the damping term of the motor, respectively. Both ends of the prosthesis are fixed for the bandwidth analysis, consequently, the equation of motion for this model becomes a standard second-order differential equation for a spring-mass-damper system. The spring force Fs was considered as the system output" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003937_978-1-4471-5104-3_5-Figure5.1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003937_978-1-4471-5104-3_5-Figure5.1-1.png", + "caption": "Fig. 5.1 Cross-sectional depths in a water stream", + "texts": [ + " This operation must be repeated several times taking care to discard the measures that deviate too much from the average. With the elapsed time so determined and knowing the distance L, the speed is calculated. The continuity equation must be multiplied by a correction coefficient for surface speed, considering the boundary layers of water. In practice, this correction coefficient is equal to 0.8 and the continuity equation becomes: Q \u00bc 0:8LA=t \u00f05:1\u00de At least two areas must be measured and then the arithmetic average of the areas is calculated as depicted in Fig. 5.1. The cross-section areas are limited by the upper water level and by the stream, creek, or small river bottom. A second method to obtain the water flow rate in streams, creeks, and small rivers is to use a rectangular spillway. This method leads to more accurate results than that of ordinary floats, but it requires more work and it is limited to the cases where the morphological conditions of the watercourse allow its use. In this case, the watercourse is dammed with a panel made of wood boards with a rectangular opening on its upper center by where all water to be measured must pass" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003833_aps.2015.7305409-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003833_aps.2015.7305409-Figure1-1.png", + "caption": "Fig. 1. Schematic drawing of the proposed dielectric antenna using a tapper section and dielectric ring resonator", + "texts": [ + " The antenna is excited through standard metallic waveguide WR15. Effects of various antenna parameters such as dimensions of the DIG, and the dielectric resonator, thickness of the substrate, and the dielectric resonator on the radiation performance of the antenna are investigated. The characteristics of the proposed DRA are demonstrated in terms of some simulation results employing CST studio. Fabrication and experimental investigation are in progress. II. ANTENNA GEOMETRY DESIGN AND FABRICATION Figure 1 shows schematic drawing of the proposed antenna. The main segments of the antenna are the tapper section and ring dielectric part. It occupies less than \ud835\udf06 \u00a0 by \u00a0 \ud835\udf06/2. \u00a0The \u00a0excitation \u00a0of \u00a0the \u00a0antenna \u00a0is \u00a0performed \u00a0through \u00a0a \u00a0 dielectric \u00a0 image \u00a0 guide \u00a0 designed \u00a0 to \u00a0 work \u00a0 in \u00a0 the \u00a0 V-\u2010band \u00a0 [2]. \u00a0 The \u00a0 dielectric \u00a0 image \u00a0 guide \u00a0 is \u00a0 excited \u00a0 through \u00a0 a \u00a0 standard \u00a0metallic \u00a0wave-\u2010guide \u00a0WR15. A transition from the WR15 rectangular waveguide to the Si image-guide is realized by a tapered section of the image-guide inside the rectangular waveguide" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000798_j.epsr.2006.03.004-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000798_j.epsr.2006.03.004-Figure1-1.png", + "caption": "Fig. 1. Schematic diagram of the motor winding connection diagram.", + "texts": [ + " That is, operation in the steady tate at synchronous speed, with a capacitor in the auxiliary inding and supplying a non-fixed amount of load. Change in oad is assumed to lead directly to change in load angle, without otor oscillations. The machine is assumed to have been started at he point where the voltage angle is zero and on no load. In other ords, the initial starting conditions have no influence on the erformance of the motor. The main and auxiliary windings are patially displaced from each other by 90\u25e6 electrical radians as hown in Fig. 1. Suppose voltages across the main and auxiliary indings are: a = Va cos(\u03b8e + \u03b4) and vb = Vb cos(\u03b8e + \u03b4) (3) here Va = Vb = V, \u03b8e = \u03c9et and \u03b4 is the voltage load angle. The transformation adopted for all state variables is: Fa Fb ] = [ cos(\u03b8r + \u03b5) sin(\u03b8r + \u03b5) sin(\u03b8r + \u03b5) \u2212 cos(\u03b8r + \u03b5) ] [ Fq Fd ] (4) here F can be current, voltage or flux linkage in their respecive reference frames, \u03b8r = \u03c9rt and \u03b5 is the initial rotor angular osition. The stator axis voltages obtained by using Eq. (4) in q. (3) are: qs = 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003647_j.msea.2014.02.078-Figure11-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003647_j.msea.2014.02.078-Figure11-1.png", + "caption": "Fig. 11. Contour map showing the length fraction of \u03a33 boundaries as a function of deformation temperature and strain. Depending on the processing parameters, different mechanisms may dominate and effect the development of microstructure and grain boundary character.", + "texts": [ + " The formation of denser subgrain networks and following deformation at strains between 25% and 80% further restricts the grain boundary mobilities during annealing as the average grain sizes in these samples are on the order of 22.2\u201325.8 \u03bcm. Despite the smaller average grain sizes, \u03a33 length fractions between 59% and 63% were obtained in these samples. To summarize, a preliminary process map spanning a broad range of temperatures and strains for grain boundary engineering of IN600 was created, Fig. 11. Although the benefits associated with grain boundary engineering are substantial, process routes that rely on strain annealing mechanisms that are enabled via multiple iterations of cold deformation followed by annealing are not commercially viable and only utilized in a limited number of applications. The present investigation was aimed at quantifying the effects of different strain accommodation mechanisms on the formation of \u03a33 boundaries. Consistent with other published studies, strain annealing or deformation to relatively low levels of strain (11%) followed by annealing was noted to significantly increase the population of \u03a33 or twin boundaries" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000206_tmag.2013.2242087-Figure17-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000206_tmag.2013.2242087-Figure17-1.png", + "caption": "Fig. 17. Open-circuit magnetic results using finite element analyses for the radial magnetization pattern.", + "texts": [ + " 15, self-inductance and mutual inductance become constant (independent of rotor position) in the case of machines with surface-mounted magnet, i.e., . Fig. 16 illustrates the analytical results of the open-circuit magnetic flux density distribution under three different magne- tization patterns: radial, parallel, and Halbach. The magnitude and direction of the magnetic flux density are indicated by color and arrows, respectively. As mentioned before, the FEM is employed to evaluate the proposed analytical expressions. Fig. 17 shows the open-circuit magnetic results using finite element analyses for the radial magnetization pattern. Considering identical assumptions for the analytical and numerical results, taking enough number of harmonics into account in the case of analytical calculations, and using the FEM package properly and efficiently should lead to similar analytical and numerical results. To this end, in the case of analytical calculations, the number of harmonics in all regions and for both open-circuit and AR analyses is set to be 100" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002891_s0007-8506(07)61584-4-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002891_s0007-8506(07)61584-4-Figure2-1.png", + "caption": "Fig. 2: Cu t t e r d e f i n i t i o n Fig.: Workpiece d e f i n i t i o n - by t h i n elementary", + "texts": [], + "surrounding_texts": [ + "The analysis is appreciably d i f f e r e n t from the one f o r t he \"d i r ec t problem\". though the same symbols and geometrical r e l a t ions a r e used. Reference is made t o Fig. 5 t o i l l u s t r a t e t h i s method.\nThe f l u t e is defined by a number of po in t s which a re located on a s e t of m e l l i p s e s . O n each e l l i p s e i , PL. and PR. represent the points of i n t e r sec t ion with the given f j u t e p ro$ i l e . Each e l l i p s e represents a normal view of the cross-sect ion of s l eeve RUi. The given points a r e X R . and YR. ( see Fig. 5 ) .\nThe fvllvwing f a c t s should be noted:\n- Several d i s c c u t t e r s may he involved simultaneously i n cu t t i ng a h e l i c a l curve on the s leeve RUi.\n- A given d i s c c u t t e r j may he involved i n simultaneously cu t t i ng seve ra l h e l i c a l curves on d i f f e r e n t s leeves RUi.\n- Since, according t o t h i s method, t he f l u t e t heo re t i ca l ly generates t he too l p r o f i l e , the smallest r e su l t i ng values R. f o r each d i s c c u t t e r w i l l be chosen as the l o c a l c u t t e r s?ze, as they reprsent t he l e f tove r mater ia l .\nFrom the given f l u t e p r o f i l e t he following values a r e def ined:\nxRN . = XR . cos ( aH) (19) J J J zRN. = XR. s i n ( a H ) J J\nA s point PRN. is also located on the h e l i c a l curve, i t can he subs t i t u t ed ?,to equation (13) t o obtain the constant of t he h e l i c a l curve:\nK R j = zRN. - (L/2n)+RN. (20)\nFurther , s u b s t i t u t i n g equation (20) i n t o (13) y i e lds the function of t he p a r t i c u l a r curve\nzRj = zRNj - L/2n ( 4 R N . - O R . ) (21) J J\nwhere\n4 R N j = Arctan(yRN./xRN.) J J (22)\nSubs t i t u t ion of equations (19) and (22) i n t o equation (21) y i e lds the h e l i x funct ions zR. = H(4R.) a s follows:\nEach c u t t e r d i s c j . whose r ad ius R. has not ye t been establ ished a t t h i s s t age , i n t e r s e k s the curve (23) a t point PR.. The general function f o r the c u t t e r d i s c s was given i n eqdation ( 1 2 ) .\nBy s u b s t i t u t i n g equations (10) and (11) i n t o equation (12) and rearranging w e obtain the function zR. = D ( 4 R . ) f o r t he d i s c c u t t e r a s follows: J\nEquations (24) and (23) meet a t point PR. both being funct ions of zRj and OR.. The so lu t ion f o r 4R. id 'given by J J\n4 R j = M - N cos(4R.) J (25)\nwhere\nRUi N=P tan(ac)\nThe value of +R. may now be obtained from equation ( 2 5 ) , by applying numerigal methods. The corresponding c u t t e r d i s c radius R. i s obtained by 4R., using equation (9 ) f o r t he funct ionJof the e l l i p s e . A 2 already mentioned above, only the minimum values of R. w i l l be p lo t t ed aga ins t V . as p a r t of t he c u t t e r p r o f i l e .\n6 . SOFTWARE\nIn the following, only the d i r e c t problem w i l l he discussed.\nThe accuracy of the obtained r e s u l t s , which a r e i n the form of p r in t ed da ta o r p lo t t ed p r o f i l e s , depends s o l e l y on the densi ty of the computed po in t s , i . e . , the number of s h e l l s and d i s c s se l ec t ed by the designer . Both va r i ab le s can be chosen a t w i l l . However, one should note tha t t he required computing time depends very much on these parameters. The other parameters which have t o be se l ec t ed a re : The v e r t i c a l and the horizontal o f f s e t s , t he s i z e s of t he c u t t e r and the workpiece, t he lead of t he f l u t e , or h e l i x angle and the c u t t e r angle .\nThe program includes a menu screen ( see Fig. 6 ) , al lowing t o s e l e c t pr imari ly the constant parameters such as angles and\nlengths . Also, f u r t h e r options such a s a c u t t e r menu, p r i n t i n g procedures and q u i t t i n g the program a r e ava i l ab le . The entered da ta a r e automatical ly saved i n a spec ia l f i l e during q u i t t i n g and reloaded i f needed again a t a l a t e r s t age . For the \"d i r ec t method\" the c u t t e r menu provides a choice between four d i f f e r e n t opt ions f o r t he c u t t e r p r o f i l e s : Semi-circular , wedge, t r i angu la r and general c u t t e r . After t he choice of t he l a s t opt ion, t he program w i l l ask the use r t o Create o r Load a c u t t e r f i l e , i f such a f i l e already e x i s t s . A t t he r i g h t hand s i d e of the menu sc reen , a small c u t t e r shaped f igu re w i l l i nd ica t e the c u t t e r angle r e l a t i v e t o the ax i s of r o t a t i o n of t he workpiece, a f t e r t h i s value has hccn cntcrcd, thus a id ing t o v i sua l i ze the r e l a t i v e posi t ion of c u t t e r and workpiece. When the program s t a r t s running, a new screen d i sp lays the \"normal view plane\" of the h e l i c a l f l u t e a s given by the coordinates ( X . Y ) . A l l t he other da t a of t he process a r e displayed on the lower p a r t of the screen ( see Fig. 7 ) .\n7 . DISCUSSION AND CONCLUSIONS\nThe CAD program i s , i n p r inc ip l e , a shor t and compact one. I t was, however, increased considerably t o allow f o r a l a r g e va r i e ty of parameters t o be s e t by the designer i n seve ra l screen menues i n order t o obtain an easy and use r - f r i end ly operat ion mode.\nIt was found t h a t t he e a r l i e r generation of Personal Computers was too slow i n execution. Only the more advanced systems using mathermatical co-processors and Turbo compilers made i t possible to have a s u f f i c i e n t l y accurate f l u t e p r o f i l e designed within a reasonable time i n t e r v a l , e . g . , from a few seconds up t o a maximum of a few minutes.\nBoth methods, t he d i r e c t and the i n d i r e c t , can he used one a t a t i m e f o r checking the r e s u l t s obtained by the o the r . This can be seen a s a design s t age and a simulation s t age ac t ing simultaneously, and y i e ld ing improved accuracy and b e t t e r r e l i a h i l i t y . It appears t h a t d i f f e r e n t program packages w i l l have t o he developed f o r d i f f e r e n t appl icat ions, though t h e bas i c mathematics w i l l not he changed. In d r i l l s , f o r example, t he f l u t e shape contr ibutes t o the curvature of the d r i l l po in t c u t t i n g edge 151. The basic ana lys i s of d r i l l s shows t h a t the f l u t e p r o f i l e on the plane perpendicular t o the d r i l l a x i s , i s t o he d e a l t with r a t h e r than the normal plane, which is discussed i n t h i s s tudy. Hence, d i f f e r e n t requirements w i l l , under ce r t a in circumstances, br ing about t he development of a new version of the CAD program.\nfollowing conclusions can he drawn:\nThe geometric problem involved i n undercut t ing during f l u t i n g was def ined and explained. The inverse problem involved i n p red ic t ing the c u t t e r p r o f i l e f o r a required f l u t e geometry was def ined. Two mathematical solut ions were developed f o r t he fundamental p a r t s of t he computer-aided design programs; one f o r designing f l u t e s and the other f o r t h e required c u t t e r p r o f i l e s . Further s tudy is t o he conducted i n order t o have t h e f i n a l CAD programs t e s t ed versus ac tua l machining t e s t s , and i n order t o allow the optimization of d r i l l po in t s by con t ro l l i ng the cu t t i ng edge curvature.\nBIBLIOGRAPHY\nMeister. I . , 1972, Mill ing of a Helical P r o f i l e on a Cylindrical Body, M. Sc. Thesis, Faculty of Mechanical Engineering, Technion, I . I . T . Friedman, M.Y. e t a l . 1971, A Method t o Determine the P r o f i l e of a Disc-Type Cutter When Mil l ing a He l i ca l S l o t , I s r a e l Journal of Technology, 329-332. Dudley, D.W., 1962. GEAR HANDBOOK, F i r s t Ed., McGraw-Hill, U.S.A. AGULLO, J . e t a l , 1983, On the Design of Mil l ing Cutters or Grinding Wheels fo r T w i s t D r i l l Manufacture, A CAD Approach, Department of Mechanical Engineering ETSEIB, University Pol i technica de Catalonia , Spain, 315-320. KALDOR, S . , 1986, A Common Denominator f o r Optimal Cutting Tool Geometry, Annals of the CIRP, Vol. 3511, 41-44.", + "d X\nby t h i n elementary\n~\ngeneral s i z e , workpiece and cutter-work r e l a t i v e o r i en ta t ion\n~ M A l H - M E N U\nt - ) RADIUS OF SHAPE: 50\nIIIGLE OF CUTTER: -30 UERTICRL POSITION: 258 HORIZONTAL POSIIIOH: -12 CUTTER HENU 97 [Y/NI PRINT RESULTS [Y/#l 7 N HlN. NO, OF SLEEUES- 5 QUlI 7 7 [Y /N l '. I .. OF THE FturE: 544\nUEDGE CUTTER PRESS [ESCI TO CONIINUE ' 9 '\nHEDGE CIIIIER\n~ Fig. 6 : Parameters t o be s e t before operat ing the p l o t t i n g procedure by [ E X ]" + ] + }, + { + "image_filename": "designv6_24_0000485_icorr.2013.6650440-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000485_icorr.2013.6650440-Figure3-1.png", + "caption": "Fig. 3. Device mounted to side of user as well as detailed view of the mechanism.", + "texts": [ + " As the wheel rotates counterclockwise the roller attached to the wheel pushes the slider to the left. After the wheel has rotated 180 degrees the slider reaches the end of its range of motion and as the wheel continues to rotate the slider begins to move toward the right. This reciprocating motion continues as long as the wheel continues to turn. When the wheel rotates at a constant velocity the slider reaches its maximum velocities when the wheel is at top dead center and bottom dead center. The mechanism used to actuate the hip joint in the sagittal plane is shown in Fig. 3. This mechanism takes advantage of the cyclically varying transmission ratio of the scotch-yoke mechanism. Periods of peak torque and low speed in the mechanism are lined up with periods of peak torque and low speed of the individual\u2019s leg. Similarly, periods of low torque and high speed of the mechanism, at top or bottom dead center, are lined up with the swing phase of the leg since this phase has the highest speed and lowest torque requirements. A 200W brushless DC motor turns a worm gear through a 36:1 fixed reduction ratio. This worm gear turns the wheel of the scotch-yoke mechanism. Attached to the wheel is a roller that pushes on a track connected to the slider. As the wheel spins the roller pushes on the track and moves the slider back and forth. A roller is used instead of a sliding peg to reduce frictional losses and to ensure smooth motion. The slider is made up of two hollow tubes (transparent in Fig. 3 for clarity) with springs inside them to create a series elastic effect. Series elasticity is commonly used in human-machine interfaces and has many benefits such as reduced control bandwidth and improved user comfort [9]. These springs can also be used to measure the torque currently being exerted at the hip joint. Since the springs have a known stiffness and their displacement can be easily measured, the force can be determined and the moment can be computed. This eliminates the need for strain gauges to be placed along the leg bar" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001137_icelmach.2008.4800194-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001137_icelmach.2008.4800194-Figure8-1.png", + "caption": "Fig. 8. PMSM geometry for 1 pole", + "texts": [ + " The PMSM included in the comparison is based on the same stator as the EMSM and the SMSM. The rotor, however, has V-shaped embedded permanent magnets with air pockets in the extremes of the magnets. The V-shaped magnet arrangement and the inclusion of air pockets in the design is motivated by the increase in reluctance torque achieved, which besides of contributing to a better torque density of the machine, will improve the high speed performance increasing the field weakening capabilities of it. Fig. 8 illustrates the stator and rotor geometry analyzed in this case. Different magnet lengths and angles are scanned in order to find the best magnet configuration for the rotor. Two design parameters, namely dmi (distance from the axial center to the lowest corner of the magnet) and dpn (distance from the symmetry line between the 2 magnets to the closest corner of the magnet) are varied successively in order to evaluate their effect in the performance of the machine. This kind of sensitivity study is similar to that performed in section IV when the rg/ro and ri/ro ratios were to be selected" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001292_j.1525-1594.1997.tb00257.x-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001292_j.1525-1594.1997.tb00257.x-Figure2-1.png", + "caption": "FIG. 2. This diagram of the moving actuator type artificial heart shows the pendulous motion of the energy converter.", + "texts": [ + "uator type Korean TAHs have been described elsewhere in detail (6). Briefly, this type of electromechanical TAH consists of left and right sacs and a moving actuator that includes the motor (Fig. 2). The moving actuator is utilized as an energy converter for a medium speed, medium torque brushless DC motor (S/M 566-18, SierracidMagnedyne, Vista, CA, U.S.A.), generating a pendulous motion through an epicyclic gear train. The blood sacs are double and made of smooth seamless segmented polyurethane (Pellathane, Dow Chemical, U.S.A.). Monoleaflet polyurethane valves were used in the blood pumps. Received February 1996; revised August 1996. Address correspondence and reprint requests to Dr. Won Gon Kim, Department of Thoracic and Cardiovascular Surgery, Seoul National University Hospital, Seoul National University College of Medicine, Yungeun-Dong, Seoul, 110-744, Korea" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002561_ecce.2018.8557396-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002561_ecce.2018.8557396-Figure1-1.png", + "caption": "Fig. 1. Structure of the proposed 6-slot / 2-pole bearingless motor in one unit", + "texts": [ + " Accordingly, a method of enhancing efficiency of the proposed bearingless motor by means of high switching frequency using SiC-MOSFET is examined. The experimental results with the examined method show that the prototype machine can attain the high efficiency of 94.3% at continuous rated operation of 1.70 kW and 100,000 r/min in spite of including suspension input power for magnetic levitation. In addition, operational characteristics comparison of the bearingless motor at switching frequency of 40 kHz, 80 kHz and 120 kHz is discussed in detail. II. STRUCTURE OF THE PROPOSED MOTOR Fig. 1 shows the cross-section view of the proposed ultrahigh-speed BelM, and Table I lists its specification. In order to generate constant suspension force, pole pair number of suspension winding must be p\u00b11 with respect to pole pair number p of motor winding, because of the characteristics of BelMs [8]. The number of poles of the motor winding was set to 2, which is with lowest drive frequency of 1.67 kHz. Therefore, the stator has two kinds of winding: 3-phase 2-pole motor winding for torque generation and 3-phase 4-pole suspension winding for suspension force" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001941_1045389x12451194-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001941_1045389x12451194-Figure8-1.png", + "caption": "Figure 8. Detail of instrumentation placement and dimensions for 3D actuation (all distances in millimeter).", + "texts": [ + " From the clamping point (left) to the tip (right), one finds, respectively, the link segment actuated by the OoP + InP actuators, the link segment with sensors, and the link segment actuated by the Tor actuators. For practical reasons, the cross section of the beam used for sensing is not the same used for actuation. However, it can be seen in Figure 6 that they are close enough so that a quasi-collocated beam vibration control loop can be implemented for the first six modes. The final configuration of the link is represented in Figure 7, while a detail of the link instrumentation is shown in Figure 8. The arrows in Figure 8 represent the fiber direction of the piezoelectric material. MFC actuators from Smart Material Corp., model M8528-P1, were used as actuators for InP and OoP degrees of freedom (fibers aligned along the length of the link)\u2014model M8528-F1 was used as actuator for the Tor degree of freedom (fibers at an angle of 245 with the direction of length) and model M2814-P1 was used as sensor (the MFC itself is bonded at an angle of 245 with the direction of length). The Levy method is a classical method of single-input, single-output (SISO) system identification from experimental results in frequency domain" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001320_861355-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001320_861355-Figure1-1.png", + "caption": "Figure 1 - Examples of kinematic arrangements for traction drive CVTs (taken from", + "texts": [ + ", the 1906 Carter car powered by a 12-hp engine) was in fact produced for a number of years with exactly such a CVT. Some low-power yard equipment successfully utilizes disc/wheel CVTs even today. Traction drives are one of the first CVT concepts ever invented. The high level of crea tivity exhibited by engineers at the dawn of the industrial revolution is reflected in the numer ous, often most ingenious, kinematic arrange ments of rollers, discs, cones, and ather bodies of revolution, ratio control schemes, and mate rials which have been used in order to produce more perfect traction drive CVTs. Figure 1, taken from Reference 1, is but a small sample illustrating the various concepts employed. However, literally dozens of additional configu rations exist. Some of these drives can be traced through patents to the second half of the 19th century. Even today, many of them are being manufactured by a number of companies and successfully marketed for industrial use. Gen erally, these industrial drives are heavy, bulky, limited in power capacity, and utilize ratio control mechanisms which are slow-respond ing and not readily suited for automotive use. Although industrial traction drive CVTs can claim a modest success, it is not so with auto motive applications. The traction drive Con figuration which almost exclusively was experi mented with for automotive applications is the ntoric\" drive, represented schematically in Figure 1 as drive \"12.\" The name comes from the toroidal cavity machined between two discs which are facing each other. Three equally spaced rollers (only one is shown in Figure 1) transmit the torque from the input disc to the output disc in parallel with each other. This increases the power-to-weight ratio of this type of drive above that of other configurations and into a practical range of automotive transmis sions. The concentric and compact geometry of the toric drive, when coupled with a long list of intrinsic desirable attributes of traction drives (to he discussed later in this paper), have provided further incentive to use this particular configuration for automotive transmissions", + " This capability of a toric drive can be best illustrated by plotting the total angle of spiral in race revolutions needed to traverse the full ratio range for various roller inclination or nsteering angles n (Figure 12). For example, at two degrees of inclination and at 3000 rpm, the total ratio range of 6.25 can be traversed in less than five revolutions or in less than one-tenth of a second. This illustrates the great capability for fast response of these units and also points out the potential for serious stability problems if synchronization and dimensional position accuracy of all rollers is not maintained throughout the life of the transmission. 861355 9 arrangements) as is apparent from Figure 1. Consequently, different advantages and limita tions may apply to different drives. However, as was mentioned earlier) automotive traction drive CVTs were usually similar to the ntoric\" geometry; therefore) certain common character istics can be readily identified. Nevertheless, the reader is cautioned that the degree to which a specific drive exhibits certain attributes) may they be positive or negative) has to be judged by analyzing each concept individually. ADVANTAGES - Most of the advantages offered by traction drive CVTs over other types of CVTs stem from the basic mechanism of transmitting power through traction forces existing within lubricated rolling contacts" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001067_tmag.2017.2760800-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001067_tmag.2017.2760800-Figure6-1.png", + "caption": "Fig. 6. Flux distribution of optimal designs at (a) T = 20\u00b0 (b) T = 100\u00b0 (c) T = 140\u00b0 (d) T = 180\u00b0.", + "texts": [ + " In (1), Ci represents the center of the circle shown in Fig. 5 Min[((1/Power factor) \u2212 C1) 2 + ((1/CPSR) \u2212 C2) 2 + (Torque ripple \u2212 C3) 2 + ((1/System \u2212 efficiency) \u2212 C4) 2 + ((1/Starting torque) \u2212 C5) 2 + (Rise time\u2212C6) 2]. Subjected to : Demagnetization proximity \u2264 0 (1) Having applied the constraint which guarantees a safe operation in terms of the demagnetization of the PM material, four different designs, each representing an optimal structure at a specific temperature, are illustrated in Fig. 6. Table I as well as Fig. 6 reveal some interesting temperature-related design guidelines which may be used in the future optimizations. Except for the optimal design at T = 20\u00b0, which is far off a practical temperature for a PM machine, the remaining designs follow a monotonically decreasing or increasing values of the design variables as follows. 1) At higher temperatures, smaller stator yoke thicknesses are sufficient for carrying the magnet flux. 2) The inset depth should be increased with temperature in order to prevent an irreversible demagnetization which is more likely to happen at higher temperatures (Figs" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000467_vlsi-soc.2019.8920335-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000467_vlsi-soc.2019.8920335-Figure1-1.png", + "caption": "Fig. 1: Gate-level arithmetic circuit (FA): a) circuit diagram; b) AIG representation for half adder (HA) extraction.", + "texts": [ + " It is done in a reverse topological order, from the primary outputs (PO) to the primary inputs (PI), hence referred to as backward rewriting [1]. The rewriting transformation simply replaces each variable in the polynomial with an algebraic expression of the corresponding logic gate. Once a given variable (output of a gate) is substituted by an algebraic expression of the gate inputs, it is eliminated from the expression. As a result, the final signature is unique and expressed in the primary inputs only. A set of ordering rules is imposed on the rewriting order to maximize its efficiency [1][11]. Figure 1(a) illustrates the rewriting process for a gate-level circuit to prove that it is a full adder (FA). The output signature of the circuit is Sigout = 2C+S, determined by the weights of the binary encoded output signals, carry C and sum S. During rewriting, the signature polynomial is iteratively modified by moving across the gates and substituting variables representing the gate outputs by the respective logic expressions in terms of their inputs. Two types of simplifications are applied during rewriting: 1) Reduction of the terms with the same monomials, reducing some of them to zero; and 2) replacing the term xk with degree k > 1 by x, since for binary signals xk = x", + " Several attempts have been made to identify the bug(s), either by comparing the result of backward and forward rewriting [13] or by analyzing the difference between the computed input signature and the given specification [14]. With a notable exception of finite field (GF) arithmetic circuits [15] [16], circuit debugging remains an open problem. The rewriting process can be improved by using a functional, And-Invert Graph (AIG) representation of the circuit [17]. This technique identifies half- and full adders in the circuit, shown in Figure 1(b), and performs the rewriting directly over those sub-circuits instead of its logic gate components. It seems at first that such a rewriting model cannot be directly applied to the divider. The characteristic function of the divider can be described by the following expression: X = D \u00b7Q+R, with R < D (1) where X (the dividend) and D (the divisor) are the inputs, and Q (the quotient) and R (the remainder) are the outputs. The problem is that the outputs, Q,R, cannot be directly expressed in terms of the inputs X,D", + " It also searches for XORs and specific logic patterns present in the reference divider. An error is declared if such functions cannot be identified in the circuit. While the CPU runtimes are very good, such a reverse engineering method, based on a strictly structural pattern matching, does not provide the functional verification per se. It may happen, that some components do not match the expected logic, but may work correctly as an ensemble. Or, that some logic is represented without XORs. As an example, Figure 1 shows a non-standard full adder implementation without XORs. In contrast, in our work the divider is modeled in a single functional specification, X = D \u00b7 Q + R, to be compared to the signature computed by algebraic rewriting. This approach works for any combinational divider circuit, regardless of its internal structure. Fractional divider is an essential part of hardware for the floating point division. The dividend X and the divisor D are normalized by preshifting to comply with the IEEE 754 standard" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002010_jae-2008-999-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002010_jae-2008-999-Figure7-1.png", + "caption": "Fig. 7. Impedances of the 8.5[mm] outer diameter RUSM obtained from the simulation and the experimental data.", + "texts": [ + " (77) had to be modified because the paralleled electrodes of the RUSM: 1 Ztot(w) = n\u2211 i=1 1 Zi(w) = Z2Z3 . . . Zn\u22121Zn + Z1Z3 . . . Zn\u22121Zn + . . . + Z1Z2 . . . Zn\u22122Zn\u22121 Z1Z2 . . . Zn\u22121Zn (78) where, Ztot: the total impedance of the RUSM Zi: the impedance which is calculated by Eq. (71) n: the total number of the paralleled input part (in this case n = 2). Hence, the impedance of the RUSM can be expressed by: Ztot(w) = n\u220f k=1 Zk n\u2211 i=1 n\u220f j=1,j =i Zj (79) where, n\u220f j=1,j =i Zj : the accumulated multiplication from j is 1 to n except when j = i. Figure 7 compares the impedance between the simulation and the experiment. The simulation result was obtained by 3-D FEM and Eq. (79), and experimental data by the impedance analyzer. The number of the element for the FEM analysis was 29,568. The tetrahedral mesh with 4-node is used for an element. As shown in Fig. 7, the data from the simulation was almost in agreement with the experimental data. This agreement proved that 3-D FEM routine used in this research was correct. The calculated resonance frequency value, which generated the desired mode properly, was 246,200[Hz] and the experimental result was 248,125[Hz]. Figure 8 shows the calculated result of the 6 th-mode wave at the calculated resonance frequency 246,200[Hz] with 14.14[Vrms] input voltage. Briefly, the vibrator of an 8.5[mm] outer diameter RUSM was analyzed and designed considering the resonance frequency, the mode, the size, and etc" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001147_pesc.1985.7071009-Figure14-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001147_pesc.1985.7071009-Figure14-1.png", + "caption": "Fig . 14 : Illustration of the flow chart", + "texts": [], + "surrounding_texts": [ + "The shape of all lower odd harmonics up to the ordinal number 69 is shown in Fig. 7 as func tion of the fundamental stator voltage U10 . As result from the phase relations between U10 ' U and U 0 forming a symmetrical three phase s~~tem th~ remaining 1~er 91d har~l1nics ~~ vi\u20ac~ble by three (3 , 9 , lS , 2 1 , 27 , ... ) are compensated in the voltages U1 , U and U3 applied to the stator windings 1f tfte star point connection is left open. So the complete frequency spectrum of the stator voltage is shifted towards higher frequen cies.\nFi'S. 8: Connection between voltage source inverter and induction motor\nTo control the inverter operation a microcom puter is directly connected to the PWM -cir cui t . It enables the independent control of the stator VOltage and frequency .\nThe digital values of the stator voltage and frequency are tra.nsformed into a three phase pulse pattern which is applied to the firing cicui t of the PWM inverter. To generate the stator frequency from a constant reference frequency a programmable rate generator is used . The subsequent counter together with the U - register adresses a PROM from which the a~tual state of the pulse pattern is deter mined. The pulse patterns for a complete pe riod of all three phases are composed by a multiplexer from three different pulse pat terns for a 300 interval . In order to indicate the zero crossing of the stator voltage there are also three signals generated which are connected to the phase detection circuit.\n~ Digital Filte r for Phase Angle Detection\nAlthough the reactance of the induotion motor smoothes the pulsating stator VOltage there still remains a ripple in the stator current. Therefore it is necessary to use a filter for the detection of the zero crossing of the stator current . For the design of a low pass fil tel' it can be supposed that the harmonic oontent of the current is also shifted to wards higher frequencies as a result from the employed PWM-method. The usual analog filter methods mostly have the disadvantage of distorting the phase in formation aocording to a nonlinear function of the frequency . Digi tal fil tel' teChniques offer linear phase FIR filters which are wi dely used already in oommunication systems. The utilization of digital filters is in par ticular advantageous if an adaptation of the fil tel' parameters is demanded and the filter runs in a range of lower frequencies. For the filter design and for the optimization of the filter order efficient algorithms (2) are available. The filter coeffioients can be de rived direotly from the impulse response shown in Fig. 10 for a filter length of nf= 11. From the oorresponding phase response (Fig. 11) the neoessary phase correction can easily be com puted .", + "The impulse response as well as the phase response in the given example result from a Tschebyscheff approximation of a low pass filter with a sampling rate of 100 Hz, a passband cutoff frequency of 10 Hz and a stoppband cutoff frequency of 30 Hz. With the weighting factor ka = 1 between passband and the stopband ripple a stopband attenuation of -40 dB is achieved. Such a filter can be applied in a range of lower stator frequencies up to 10 Hz . The adaptation of the current filter to the operating states of the machine is managed by different sets of offline computed filter co efficients and according to that a variation of the sampling rate. Fig . 12 gives an outline of the microcomputer based current filter and the phase detection circuit. A flow chart (Fig . 13) demonstrates the principles of the filtering process . Its major steps are illustrated in Fig . 14.\nThe detection of the zero crossing of the voltage is supported by a hardware signal ge nerated by the PWM circuit which opens the phase gate. It is closed by the first follow ing sampling pulse which simultaneously ini tializes step 2 of the flow chart. The follo wing steps 3 and 4 are carried out completely under software control . A final phase correc tion enables the compensation of the phase lag caused by the digital filter . In addition to that a phase correction from an analog Bessel prefilter which also provides a linear phase has to be taken into account.\nNeglecting the phase of ini tialization the program is organized in three interrupt le vels. With the highest priority the current samples are taken from the AOC with a pread justed sampling rate and are continuously counted. In the next lower priority level \\30 0 '. ' dBi] 30190Frequency = 19 \\ Main Moe magnitude = 11.6 dBi\\ Main Me direcion = 0.0 deg. Angular,,idth (3 dB] = 22.3 deg. Side lobe level = -1 2.8 dB Fig. 7. Untilted radiation patterns in the E-plane at 17, 18, and 19 GHz. F.,field 'f.,fi.ld ff=l 71 [1 I' Di,.Oi,,ity_Ab.(Th.t.) 90 Phi= 0 30 \u00b0 Phi=l 80 60 F.,field 'f.,field [f=l 91 [11' Di,ediAty_Abs(Thet.j 90 Phi= 0 150J/ 180 0 Phi=l180 120 60 V. CONCLUSIONS A new design for end-wall waveguide-to-microstrip transitions is introduced in this paper. The achieved large bandwidth of about 22% from 16 GHz to 22 GHz is of practical importance. The proposed transition only uses a double slit on the ground plane and a stub matching on the output microstrip line, which simplifies the fabrication process" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003694_j.1559-3584.1955.tb03131.x-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003694_j.1559-3584.1955.tb03131.x-Figure1-1.png", + "caption": "Figure 1. A comparison in size between motors of 1903, 1946, and 1955; all are of %-hp, 1725-rpm. 60-cycle rating.", + "texts": [ + " The initial redesign to a 48-frame size encompasses Westinghouse motors in the four-pole, open, drip-proof, fully guarded ratings from one-third through one-eighth horsepower, continuous duty, 40 degree C temperature rise. Later designs will include the two-pole ratings up through onehalf horsepower and the six-pole ratings up through the one-quarterhorsepowx size. Actually, the present re-rating program is but part of the continuing trend toward smaller motors that are at the same time better electrically and mechanically. Contrast, for example, the one-quarter horsepower motor of 1903 with the new fractionalhorsepower motor of the same rating (Fig. 1). The weight has been reduced from 69 pounds to but 16 for the new motor. The new motor is six pounds lighter than the motor of nine years ago-about a 27-percent reductioli. Obviously, such a drastic cut in size and weight necessitated the solution of a number of engineering design problems. Ventilation-A reduction in motor size without a decrease in power output makes an improvement in ventilation efficiency imperative. S i n c e sound level is an important factor to the user, this increase in ventilation efficiency cannot be accompanied by greater windage noise-preferably there should be less" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000336_1.3074282-Figure10-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000336_1.3074282-Figure10-1.png", + "caption": "Fig. 10 Trajectory of the contact point in the contact plane", + "texts": [ + " The position of the driving mass point in the equatorial plane of the hemisphere, defined by the curve Eqs. 49 and 50 , is shown in Fig. 9. As can be seen, qualitatively, the roselike pattern of human movements is captured. Having specified p t , we then integrated the state dynamic equations 48 . The motion is simulated on the time interval 0, 35 s. Note that the initial position of the mass point is not zero, and in the simulation we start to move the hemisphere from the equilibrium state corresponding to p 0 . This can be considered as a modeling uncertainty. Figure 10 shows the trajectory of the contact point on the contact plane. The trajectory resembles that shown in Fig. 6 top . As can be seen, qualitatively, the basic features of this trajectory the curvature of the trace and the spirallike loops along the trace are similar to those observed in the human movements. While the comparison conducted may look optimistic and speaks in favor of the actuation scheme given by Eqs. 49 and 50 , the real story of human manipulations is, most probably, still Transactions of the ASME /2018 Terms of Use: http://www" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001347_ijabe.v10i6.3142-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001347_ijabe.v10i6.3142-Figure4-1.png", + "caption": "Figure 4 Internal structure", + "texts": [ + " Churning tooth was distributed equally along sucking holes. The diameter of sucking hole is 1.5 mm. But the sucking plate thickness is very thin. The depth of sucking hole is very short. Meanwhile, there are two bristle clearing devices installed in the sucking chamber shell and one bristle clearing devices installed in the seed chamber shell. Positive pressure is used to blow seeds in dropping place. Therefore, problems such as electrostatic interference and blocked sucking hole are prevented, as shown in Figure 4. experiment According to the past study [23] , the optimal negative pressure for a single seed-metering device is 1.6 kPa. In this research, the pipeline was PVC pipe with an inner diameter of 36 mm and an outside diameter of 40 mm. The experiment was carried out at the optimal negative pressure of 1.6 kPa. The air speed of a single seed-metering device was measured by anemometer and the result was 6.6 m/s. As the air flow rate is low, it can be considered as incompressible gas. According to the formula of fluid mechanics [25] , Equation (2), the flow rate of a single seed-metering device was 24" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001093_jrproc.1954.274796-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001093_jrproc.1954.274796-Figure3-1.png", + "caption": "Fig. 3 Sketch of a tapered-depth traveling-wave slot antenna used as the arraying element.", + "texts": [ + " If the mutual impedance between terminals is insignificant the patterns Pn and Qn are the same, apart from a constant, and are independent of the loads connected to the unexcited inputs. However, these patterns is general will differ from the pattern of a single element with all other elements removed. An array of four parallel traveling-wave slot antennas was constructed to illustrate the order of magnitude of the difference between the approximate and exact methods. It is convenient to use traveling-wave slots because the mutual impedance between input terminals is insignificant. Fig. 3 shows a typical element of the array: 1264 A ugust Hines, Rumsey, and Tice: On the Design of A.rrays it was chosen because it has a good end-fire radiation pattern over a 2:1 frequency band and is easily fed from a conventional waveguide. Fig. 4 shows the principal patterns measured in the co-ordinate system shown in Fig. 3. These are the patterns which would be used in the approximate method. Fig. 5 shows the patterns obtained from the same antenna as an element of a twoelement array. Observe that the Eo(6) at 4 = 0 degree pattern is practically the same as that in Fig. 4 whereas the E,0(G) at X = 90 degree pattern is significantly different. Thus the approximate method is evidently adequate for predicting the 4 =0 degree array pattern but palpably inadequate for predicting the 4) = 90 degree array pattern. Fig. 6 shows the same patterns for a three-element array" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001573_j.apm.2014.11.058-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001573_j.apm.2014.11.058-Figure3-1.png", + "caption": "Fig. 3. 2-D equivalent magnetic circuit network of the whole structure of a coaxial magnetic gear mechanism.", + "texts": [ + " 2, the mechanism can be divided into seven layers in the radial direction due to different materials, including: the yoke of the low-speed outer rotor (layer 1), permanent magnets on the low-speed outer rotor (layer 2), the outer air gap (layer 3), the stationary frame and steel pole-pieces (layer 4), the inner air gap (layer 5), the permanent magnets on the high-speed inner rotor (layer 6) and the yoke of the high-speed inner rotor (layer 7), respectively. Then, the coaxial magnetic gear mechanism is divided into N parts in the circumferential direction, i.e. there are 7N nodes in this coaxial magnetic gear mechanism, as shown in Fig. 2. The accuracy of the analytical results is directly proportional to the number of nodes. Analogous to Ohm\u2019s Law in electricity, the 2-D equivalent magnetic circuit network of the whole structure of the coaxial magnetic gear mechanism is shown in Fig. 3. In the analysis, the geometric dimensions and characteristics of the magnetic materials of the magnetic gear mechanism should be specified in order to get the permeance and magnetomotive force between each node. Basically, the inverse of the reluctance R is the permeance P. The reluctance R of a magnetic material is expressed as: Please equiva j.apm. R \u00bc l lA ; \u00f01\u00de where l is the permeability of the magnetic material, l is the length of the magnetic material and A is the cross-sectional area of the magnetic material" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002448_2011-01-0862-Figure10-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002448_2011-01-0862-Figure10-1.png", + "caption": "Figure 10. Crankshaft arrangement, showing central location of flywheel.", + "texts": [ + " This solution also has cost benefits; as well as there being no coupling to procure, no additional bearings are required for the rotor. Although some extra tooling would be required for production to ensure the rotor is assembled to the crankshaft in a safe and controlled manner. The cranktrain cost was reduced further by the use of a two bearing crankshaft. The rotor of the generator has a very low rotating inertia, and was too low to act as the flywheel for the engine. To further reduce the length of the engine, the flywheel was positioned between the two cylinders, as shown in Figure 10, the ability to locate the flywheel in this position was facilitated by the omission of a centre main bearing. The crankshafts for the prototype engines will be made from steel, however series production items would be manufactured from cast iron to reduce material costs. For the early prototype engines the flywheel mass is interchangeable to enable optmisation. The main bearings are based on MAHLE series production bearings to reduce costs in the early stages of the program. Removing the need for a balancer shaft was viewed as being fundamental to creating a cranktrain with low series production cost" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000587_0029-5493(65)90101-9-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000587_0029-5493(65)90101-9-Figure3-1.png", + "caption": "Fig. 3. Yield hexagonal prism.", + "texts": [ + " she l l s by e l imina t ing rn 8 f rom the Hodge \" twomoment l i m i t e d - i n t e r a c t i o n \" su r face [18]; i t i s r e f e r r e d to a s a \"one -momen t l i m i t e d - i n t e r a c - t ion\" y ie ld sur face . This hexagonal p r i s m is much s i m p l e r but is by no means as good a f i t at some points . 1 R e - ducing a l l i t s v e r t i c e s by the f ac to r \u00bd(5~,- 1) p r o d u c e s an i n s c r i b e d y ie ld sur face . A t h r e e - q u a r t e r s i ze p r i s m , however , l i e s within the a c - Another approx ima t ion to the y ie ld su r face i s shown in fig. 3. Th is y ie ld condi t ion i s defined by the fo l lowing eight p l anes : E4 AXISYMMETRIC INTERSECTING SHELI.~ OF REVOLUTION 89 tual y ie ld su r f ace ove r an extended range of v a l - ues of p r a c t i c a l i n t e r e s t for p r e s s u r e v e s s e l s . 3. APPLICATION OF THE THEORIES TO INTERSECTING SHELLS In o r d e r to be able to make use of the y ie ld loci for i n t e r s ec t i ng she l l s as r e p r e s e n t e d in sec t ion 2, some approx ima t ions to the shel l conf igura t ion mus t be in t roduced", + " The results thus obtained are restr icted to those shell configurations where the assumed s t ress profile, NO= const, and conditions (14) or (15) can be imposed. By choosing his boundary conditions to satisfy face I of the yield surface of fig. 1 and the set of eq. (16), Lind [19] has been able to establish such a collapse pressure. The complexity of the equations obtained makes it necessary to use a trial and e r ro r procedure. Gill [20], on the other hand, by taking face I of the hexagonal pr ism yield locus of fig. 3, in association with the boundary conditions (12) and (18) and the set of eq. (17), arr ives at a fairly simple expression for the collapse pressure. The solution by Cloud [21] may also be derived as a particular case of Gill 's expression for the collapse load. An upper bound on the collapse pressure can be found by equating the external rate of doing work to the internal rate of energy dissipation for a kinematically admissible pattern of three hinge circles. A velocity field of the form U= c[1-cos(E-q)] , W= -csin(~-\u00a2) satisfies these hinge conditions" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002775_mnl.2015.0198-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002775_mnl.2015.0198-Figure2-1.png", + "caption": "Fig. 2 BCP assembly procedure. The sequence starts with a structure that lies flat on top of the substrate. Then a microprobe tip or wire-bonder head is used to push the front edge of the structure backwards until the lateral cantilever beams buckle out-of-plane causing the structure\u2019s plate to rise. Finally, a set of anchors placed in the sliding path of the edge prevent the structure to collapse back once the microprobe is removed. Depending on the position of the anchors and the attachment links between the buckled cantilevers and the plate, a range of angles from 0\u00b0 to 90\u00b0+ can be achieved", + "texts": [ + " Moreover, the hinged structures require a higher number of physical layers in the fabrication process. In the second case, hingeless structures such as buckled cantilever platforms (BCP, Fig. 1) [1, 3, 10] and Tsang suspensions [11, 13, 14] require a simpler procedure. They are essentially compliant mechanisms that are deformed and make use of the elastic properties of the composing materials to maintain this deformation using stoppers or a clever structural design (i.e. using their own reaction forces) [1, 11, 13\u201315]. The procedure to assemble a BCP is depicted in Fig. 2. A microprobe station needle or a wire-bonder head is used to push the front edge of the cantilever towards the structure\u2019s main anchors until the cantilever beams are buckled out of plane causing the attached plate to rise. Then a set of four structural anchors, referred to as \u2018stoppers\u2019, is used to contain the restoring forces of these deformed beams and to secure the assembly position. The plate\u2019s angular orientation depends on the position of these stoppers and on the plate\u2019s point of attachment to the buckled beams [1]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003810_851861-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003810_851861-Figure7-1.png", + "caption": "FIGURE 7- ORBITAL ELEMENTS", + "texts": [ + " 3) Initial Time1ine Analysis and Deployment Order Scenario. 4) Trajectory Profile Generation. 5) Attitude Profile Generation. 6) Payload Separation Sequence Analysis. 7) Crew Activity Planning. Upper Stage/Spacecraft Performance Assess ment - The first step of the mixing process in Flight Planning is to determine the amount of performance available from each payload's upper stage motors based on the spacecraft wei ght to be placed into its operati ona 1 orbit. \"Donut\" plots (Figure 6), desc\"ibe the Ri ght Ascensi on of the Ascendi ng Node (RAAN) (Figure 7) change capability as a function of the parking orbit inclination. The plots are generated usi ng a vel oci ty matchi ng algorithm which calculates a family of transfer orbits that match the parking orbit and final orbit. The RAAN change capability and the desired final RAAN directly affect the duration and opening/closing times of the launch window. The parking orbit altitude and the desired final mission orbit altitude, as well as the upper stage rocket motor delta-V's, affect the shape and amount of shading on the \"donut\" plot" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003841_southc.1994.498115-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003841_southc.1994.498115-Figure1-1.png", + "caption": "Fig. 1. Edge Slot Geometry", + "texts": [ + " Experimental results of the measured slot conductances and a formula for the slot depth, including the waveguide wall thickness, for any slot orientation less than 46.6\" is presented. This data was used to determine the amplitude distribution and configuration for each element of the array. Measured radiation patterns are presented and characteristics of the antenna examined. elements. If one element in a large array is slightly different than the other slots, there is a minimal variation in its electrical characteristics. If one element in the eight element array is different from the original design, the electrical characteristics vary sigruficantly. Figure 1 shows the geometry and nomenclature of an inclined edge slot in the narrow wall of a rectangular waveguide. SLOT CONDUCTANCES INTRODUCTION System requirements dictated a need for a low sidelobe, X-band antenna in order to achieve the missions desired goals. Due to the location of the antenna, it's size was limited to less than seven inches in length and two inches in width. To fulfill the performance and size requirements, an eight element edge slot waveguide array antenna was deemed the best choice for an antenna to meet the stringent requirements" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003199_j.matdes.2013.03.066-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003199_j.matdes.2013.03.066-Figure7-1.png", + "caption": "Fig. 7. von Mises stress contours taken from Fig. 6 in (a) 3rd, (b) 6th, (c) 10th and (d) 13th steps.", + "texts": [ + " The plunger and the other assembly features have been removed for better clarification. It is apparent that the maximum stress occurs in corners, because these regions are in contact with the die interior surfaces [20]. Also the heat transfer was higher at the die edges, since the specimen was hot extruded. When the heated billet is inserted into the die, the temperature of the billet will be transferred to the die interior surfaces. Hence sudden cooling of the specimen takes place. To avoid this die is maintained at the required temperature. Fig. 7a\u2013d shows the enlarged view of von Mises stress contours of four different stages of (3), (6), (10) and (13) selected from Fig. 6, respectively. Fig. 7a shows stage (3) of the deformed billet during the first twist extrusion pass in which it was found that the maximum stress is 116 MPa. The maximum stress values for stage 6, 10 and 13 during the first twist extrusion pass were found to be 139, 141 and 158 respectively as shown in Fig. 7b\u2013d. It was found that the stress increase noticed in the above case is not the same on further twist extrusion passes at different temperatures which is evident from Figs. 8\u201310. Figs. 8\u201310 depict the variations of von Mises stress versus time for the process temperatures 350 C, 400 C and 450 C respectively. The billet experiences more stress at the beginning of twist extrusion process and the magnitude of stress goes on decreasing when the billet reaches the end as shown in Fig. 8. The same trend is noticed for the simulation conducted at temperatures 400 C and 450 C and at different passes as shown in Figs" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003054_imece2007-41938-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003054_imece2007-41938-Figure1-1.png", + "caption": "Figure 1. Definition of vehicle dynamics quantities.", + "texts": [ + "org/about-asme/terms-of-use Kr, KV = gains of linearized model with respect to yaw rate and lateral velocity outputs l = wheelbase Lr = lateral relaxation length of tire M = mass r = yaw rate R = effective tire radius s = operator of Laplace transform t = track T = wheel torque Tr, TV = lead time constants of linearized model with respect to yaw rate and lateral velocity outputs Tin = power plant torque \u2206Tc = torque shift of active central differential \u2206Tr = torque shift of active rear differential U = vehicle longitudinal velocity V = vehicle lateral velocity \u03b1 = tire slip angle \u03b4 = road-wheel steering angle \u03b7 = tire longitudinal slip \u03c90 = natural frequency of linearized model \u03b6 = damping ratio of linearized model Subscripts i = tire number (i = 1,...4) j = actuator number (j = 1,...5) NONLINEAR MODEL This section presents a 6 DOF nonlinear vehicle dynamics model which is used as a basis for linearized model derivation. Fig. 1 defines the model quantities (see also Nomenclature). The model is extracted from the more general 10 DOF model [4]; the longitudinal velocity, roll, pitch, and heave DOFs are assumed to be of secondary importance for GCC-handling control, and they are neglected. State-space subsystem The chassis dynamics are described by the following lateral velocity and yaw rate state equations: M \u2212= , (1) yryf FFMUrV ++& )( 2 )( 2 4231 xxxxyryfzz FFtFFtcFbFrI +++\u2212\u2212=& , (2) where Fyf and Fyr are the total lateral forces of the front and rear axles, respectively: , 21 yyyf FFF += . 43 yyyr FFF += Longitudinal slip The rotational dynamics of the i-th wheel in Fig. 1, i = 1,...,4, are described by I RFT xtiwiiwi \u2212=\u03c9& . (3) The wheel rotational speeds \u03c9i are used to calculate the longitudinal slip 2 Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 10/10/2017 T i ii U UR \u2212\u03c9 i =\u03b7 , (4) where Ui is the longitudinal velocity of the vehicle body at i-th corner (Fig. 1): rtU i i 2 )1(\u2212+U = . (5) Lateral slip The tire slip angle is calculated as i i ii U Varctan\u2212= \u03b4\u03b1 , (6) where the lateral velocity at the i-th corner reads V , brV +=2,1 V . (7) crV \u2212=4,3 The tire relaxation lag dynamics are described by [5] Copyright 2007 by ASME erms of Use: http://www.asme.org/about-asme/terms-of-use iii ti ri U L \u03b1\u03c5\u03c5 tan=+& , (8a) , (8b) ii \u03c5\u03b1 arctan, = with U iiiiti VU \u03b4\u03b4 sincos += . (9) The angle \u03b1'i is fed to the static tire model given below. Tire load In the absence of roll, pitch, and heave dynamics, the normal load to each tire may be calculated as [2]: \u2212=\u2206\u2212\u2206= \u03a3\u03a3 x g y g zxzyzz F l h F t h l cMgFFFF mm 2 1 20,12,1 (10a) +=\u2206+\u2206= \u03a3\u03a3 x g y g zxzyzz F l h F t h l bMgFFFF mm 2 1 40,34,3 (10b) with , ", + " \u2211 = \u03a3 = 4 1i xix FF \u2211 = \u03a3 = 4 1i yiy FF Tire The tire model is based on the Magic formula model [6]. The tire forces are calculated as static functions of the longitudinal slip, lateral slip, and the normal tire load: ),,) , ziiiytixti FF \u03b1(,( fF \u03b7= . (11) The calculated tire forces are then transformed to the chassis coordinate system: iytiixtixi FFF \u03b4\u03b4 sincos \u2212= , . (12) iytiixtiyi FFF \u03b4\u03b4 cossin += Torque management The central and rear differential distribute the power plant torque Tin to each wheel according to equations (Fig. 1): 2421 cin TTT \u2206 \u2212==T , r cin T TT \u2206\u2212 \u2206 += 243T , r cin T TT \u2206+ \u2206 += 244T . (13) LINEARIZED TIRE MODEL The tire shows an emphasized nonlinear behavior, as illustrated by the tire static curves in Fig. 2 (cf. [6]). The lateral tire force Fyt is directly controlled through the tire slip angle input \u03b1. The indirect influence on the lateral force comes from the longitudinal force Fxt (the combined slip effect) and the normal force Fz. Hence, the overall linearized tire model can be described by1 1 Note that the longitudinal force input Fxt is used instead of the longitudinal slip \u03b7, because the simplified linearized vehicle model 3 Downloaded From: http://proceedings" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002482_9783527631506-Figure5.1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002482_9783527631506-Figure5.1-1.png", + "caption": "Figure 5.1 Magnitude response of (a) an LP, (b) an HP, (c) a BP, and (d) a notch filter. (e) Magnitude and phase responses of an AP filter.", + "texts": [ + "6) Let us find the frequencies \u03c91 and \u03c92, where |HBP( j\u03c9)| is (1/ \u221a 2) times that at the peak value, namely, |H0|; then, |HBP( j\u03c9)|2 = |H0(\u03c9p/Q p)\u03c9|2 (\u03c92 p \u2212 \u03c92)2 + (\u03c9p/Q p)2\u03c92 = 1 2 |H0|2 (5.7) From the above, we get \u03c92 \u2212 \u03c92 p = \u00b1(\u03c9p/Q p)\u03c9 (5.8) Solving Eq. (5.8) and taking the positive roots \u03c91 and \u03c92 for \u03c9, we get \u03c91\u03c92 = \u03c92 p and \u03c92 \u2212 \u03c91 = (\u03c9p/Q p) (5.9) If \u03c92 \u2212 \u03c91 is defined as the bandwidth (BW) of the BP filter, we then have Q p = \u03c9p BW = \u03c9p \u03c92 \u2212 \u03c91 (5.10) Thus, Q p and BW are inversely related and hence, the higher the Q p, the narrower the BW of the filter. The nature of the magnitude response of the BP filter is shown in Figure 5.1. Even though we cannot relate the BW in the same way as in a BP filter, the peak values for the LP and HP filters also increase with increasing value of Q p. As for the notch filter, depending on whether \u03c9z >\u03c9p, \u03c9z < \u03c9p, or \u03c9z = \u03c9p, it is called an LP notch, an HP notch, or a symmetric notch filter. Figure 5.1 shows the magnitude response for typical biquad LP, HP, BP, and notch filters as well as the magnitude and phase response for the AP filter. 5.3 Realization of Single-Amplifier Biquadratic Filters As mentioned earlier, a passive network consisting of only resistors and capacitors has all its poles on the negative real axis, and hence cannot have complex poles. Therefore, a passive-RC network cannot give rise to a frequency-selective transfer function. Imbedding an active device such as a voltage amplifier with a gain K in a passive-RC network, however, opens up the possibility of realizing a transfer function with complex poles" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000956_amm.271-272.1362-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000956_amm.271-272.1362-Figure2-1.png", + "caption": "Fig. 2 Different shapes of valve plug", + "texts": [ + " No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of Trans Tech Publications, www.ttp.net. (ID: 142.103.160.110, University of British Columbia, Kelowna, Canada-10/07/15,22:00:11) The internal channel structure of pressure reducing valve had important influence for the flow field characteristics. The shape of valve plug was essential to the structure. The influence of three different shapes of valve plug for the pressure reducing valve flow field characteristics was analyzed. The specific structures were shown in Figure. 2: (a) A-type plug; (b) B-type plug; (c) Improved C-type plug. Computational domain and inner grid were completed by the ICEM mesh generation software. In order to improve the quality of grid and ensure calculation precision, inlet and outlet extended passages were divided by Hexa meshing method, internal complex area was divided by Tetra Meshing method. The number of grid node was 195940 and grid number was 1091489. Control equations of numerical simulation. Do Flow was dominated by physical conservation laws, that was flow required to meet the mass conservation equation, momentum conservation equation, energy conservation equation etc [2, 3]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002482_9783527631506-Figure7.20-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002482_9783527631506-Figure7.20-1.png", + "caption": "Figure 7.20 (a) The admittance function of a series RLC-circuit, and (b) impedance function of a shunt RLC-circuit.", + "texts": [], + "surrounding_texts": [ + "3) Design the inner elements (i.e., excluding the source and load ends) of the filter using inverting and noninverting integrators and a normalized value of the scaling resistance R equal to unity. 4) Design the terminal elements using the networks in Figure 7.17a or 7.17b. 5) Perform the necessary frequency and impedance denormalization.\nExample 7.4. It is required to realize a third-order Butterworth LP filter with a cutoff frequency of 1 k rad s\u22121. The input source has zero resistance.\nThis case belongs to the class of singly terminated ladder filters. The prototype ladder is shown in Figure 7.19a. The nominal leapfrog filter structure is shown in Figure 7.19b. A frequency scaling by 103 and an impedance scaling by 104 will make all the resistances to be equal to 10 k and all the capacitances to get multiplied by 10\u22127.", + "7.2.4 Leapfrog Band-Pass Filters\nThe leapfrog technique discussed above is also applicable to BP filters with zeros at the origin and at infinity. This includes, for example, a series resonance network in the series arm and/or a parallel resonance network in the shunt arm of the ladder filter. Figures 7.20a and 7.20b show these two cases with the corresponding Y and Z functions. In the intermediate locations of the ladder, we have Ri \u2192 0 and Rj \u2192 \u221e. Thus, instead of first-order RC transfer function networks, as was in the case of\nan LP filter, we have to use second-order RC-active filter sections in this case to implement the normalized admittance and impedance functions. The second-order RC network must have the capability to produce Q p \u2192 \u221e as will be required for R i \u2192 0 or R j \u2192 \u221e. A Tow\u2013Thomas network with the summing capability at the input, which will also afford this special condition (namely, Q p \u2192 \u221e), is shown in Figure 7.21.\nThe condition Ri \u2192 0 in Yi(s), and Rj \u2192 \u221e in Zj(s) can be realized from this network by setting R1 = \u221e, that is, an open circuit. The leapfrog realization of a prototype BP filter (Figure 7.22a), using the above biquad as a building block, is illustrated in Figure 7.22b.", + "7.2.5 Operational Simulation of a General Ladder Structure\nIn Section 7.2, we considered the cases of all-pole LP and BP filters. The zeros of the transfer function were either at DC (zero frequency) or at infinity. Finite-transmission zeros occur for elliptic filters, stopband filters, and so on. This implies the presence of a parallel resonant network in the series arm and/or that of a series resonant network in the shunt arm of the ladder network. In general, then, we have to consider a series arm as shown in Figure 7.23a and a shunt arm as in Figure 7.23b." + ] + }, + { + "image_filename": "designv6_24_0003027_imece2014-38788-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003027_imece2014-38788-Figure5-1.png", + "caption": "Figure 5: Vector schematic of four-bar mechanism showing vector dyads in both its and position", + "texts": [ + " A vector schematic of a planar linkage, that is, using the complex number technique is proved to be the simplest, yet the most versatile method for synthesis of rigid-body mechanisms [30]. Most of the planar linkages may be thought of as a combination of vector pairs known as dyads [30]. In function generation, the vector loop closure \u0305 \u0305 \u0305 \u0305 \u0305 \u0305 (Figure 4) produces the following equation: \u0305 ( ) \u0305 ( ) \u0305 ( ) (4) where j is the precision-position. For path generation, motion generation (rigid-body guidance), and path generation with prescribed timing, loops \u0305 \u0305 \u0305 \u0305 \u0305 and \u0305 \u0305 \u0305 \u0305 \u0305 (Figure 5), formed by dyads \u0305 \u0305 and \u0305 \u0305 , respectively, produce the following equations: \u0305 ( ) \u0305 ( ) \u0305 (5) \u0305 ( ) \u0305 ( ) \u0305 (6) where j is the precision-position. Equations (4) through (6) can be expanded for each precisionposition to synthesize a rigid-body equivalent mechanism for function, path and motion generation, and path generation with prescribed timing. Energy Equations: Energy stored in the complaint mechanism, during its structural deformation, in the precision-position, is estimated by the potential energy stored in the torsional springs of the pseudo-rigid-body model [13, 14]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002448_2011-01-0862-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002448_2011-01-0862-Figure7-1.png", + "caption": "Figure 7. Comparison of direct acting valvetrain system with rocker actuated system: (a) Direct acting concept design, (b) Rocker actuated concept design.", + "texts": [ + " Reducing the distance between the camshaft axis and the cylinder head fire-face directly reduces the total height of the RE unit and this was prioritized to achieve minimum total package volume. The total height of a reciprocating piston engine is largely defined by the base engine geometry, namely crank throw and connecting rod length, which are normally difficult to significantly reduce. Concept CAD models of the valvetrain gave a distance between cylinder head fire face and the outside diameter of the cam sprocket of 135 mm. Replacing the conventional bucket tappet and shim with a graded bucket gave some height reduction (\u223c10 mm). Figure 7a shows a tappet supported valve stem, which enables a shorter valve guide to be used, however, with an increase in production complexity and cost. Further reductions in valvetrain height required a different layout; a two-valve-per-cylinder, wedge shaped combustion chamber model was created. CAD models of a camshaft-in-head or camshaft-in-block with rocker gave a total valvetrain height of 108 mm. The camshaft-in-block solution was rejected due to increased part count and shorter service interval due to the use of push-rods. The layout of the valvetrain, and efforts to reduce the overall height pushed the camshaft to the inlet side as far as the package space target allowed; a distance 150 mm from the centreline of the cylinders. This meant the camshaft occupied space that would normally be used for the intake manifold. The resulting valvetrain layout is shown in Figure 7b, where it can be seen that an overall height saving of \u223c20 mm was achieved over the direct acting layout with graded buckets. To keep the intake assembly within the target package space, it was decided to locate the inlet plenum alongside the cylinders (beneath the cylinder head overhang). It was a logical step to create an inlet gas path that entered the cylinder head on the fire face rather than a machined face on the side (a manifold face on the side of the cylinder head and intake pipe bend radii would significantly increase the width of the RE engine), as shown in Figure 8" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002224_3527608117-Figure5.15-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002224_3527608117-Figure5.15-1.png", + "caption": "Fig. 5.15 Principle of resistance electro slag welding (ISAF, TU Clausthal).", + "texts": [ + " This requirement causes relatively high costs for expensive welding additives. Regarding component geometry the application of the process is limited to rotationally symmetric bodies with minimum diameters of 150 mm. Substantial process advantages are high deposition rates up to 40 kg h\u20131 and the high surface performance up to 9000 cm2 h\u20131. With approximately 175 J mm\u20132 the surface energy is less than for electro slag cladding. 5.3.3.4 Resistance Electro Slag Welding (RES) Resistance electro slag welding (Fig. 5.15) is mechanized and assigned primarily for coating components with larger (wall) thickness, because the surface energy of about 190 J mm\u20132 is relatively high. Only the use of band electrodes, whose width is up to 180 mm, is of economic importance. The melting electrode is continuously supplied. Comparing the arc process with this resistance fusion welding the resistance heating of the molten slag pool serves as the heat source. The welding current can exceed 3000 A. The welding powder must have a sufficiently high resistance for the molten slag pool to produce the necessary heat but the resistance must not be allowed to cause inadmissibly high current flows over the weld pool, to avoid short-circuits or an ignition of the arc" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000974_2002-01-1347-Figure16-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000974_2002-01-1347-Figure16-1.png", + "caption": "Fig. 16. 3D solid model views are created full-screen, instantly, in Agito\u2019s, TechnADraw\u2122 without 3rd party viewers and can be rotated in any direction.", + "texts": [ + " Developed by TechnASales\u2122 a 3rd party service provider in 1996 and offered as an on-line solution in the spring of 2000. This system links the GUI configuration module, Fig 10. to the TechnADraw\u2122, CAD module Fig 15. The drawing is parametrically produced and displayed in the TechnADraw\u2122, associative CAD, module. The user can control views and interact with the full screen drawing. The 2D drawing can be printed for a record on the manufacturers plot sheet. One click can change the views from 2D to 3D wire, 3D solid, 3D solid model (Fig. 16.), and 3D ray traced, photo quality. TechnADraw\u2122 produces the highest quality sold model views of any system tested. All 3D views can be rotated, with or without dimensions. Dimensions can be turned on or off and displayed in imperial or metric. Many unique features are offered with this solution. Product/part drawings are layered. Users can turn on or off layers to view things like the mountings only. Fig 17. The active layer view can be instantly viewed in 2D or any of the solid or sold model views, with or with out dimensions" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000456_tec.2011.2163718-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000456_tec.2011.2163718-Figure1-1.png", + "caption": "Fig. 1. Wind conversion system with the DFIG topology.", + "texts": [ + "2163718 veloped grid codes with specific requirements for regulating the connection of wind power plants to the electrical network [1]. One of these is the system low-voltage ride-through (LVRT) capability, i.e., the ability of the power plant to remain connected to the grid during voltage sags [2], [3]. In most of the cases, power plants are also required to supply reactive power to the grid in order to guarantee voltage support during voltage sags [3]. WECSs using DFIG drive topologies, as depicted in Fig. 1, are the most commercialized ones over the world. This technology consists of a wound rotor induction generator with the stator terminals directly connected to the grid and the rotor supplied by a back-to-back converter, allowing a broader slip frequency range and, thus, variable speed. Its main advantage is the use of converters dimensioned for a small parcel of the generator rated power (normally 30%), thereby reducing the equipment cost. Nevertheless the advantages of the DFIG, due to the direct connection of the stator to the grid, it is more susceptible to grid disturbances than WECSs using full-scale power converters [4]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002820_primeasia.2015.7450474-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002820_primeasia.2015.7450474-Figure1-1.png", + "caption": "Fig. 1. Top view of H-shaped microstrip line fed patch antenna", + "texts": [ + " The oxygen absorption in this band is at its maximum value (10-15 dB/km) so we only have the option of short-range communications (<2 km). The operating bandwidth designs in [4], [5] cover the entire unlicensed 60 GHz frequency band. Here authors have tried to design a compact antenna. In the first part of this paper authors have included a microstrip line fed H-shaped patch antenna which has high bandwidth and is also highly efficient. For comparison, this paper also includes an antenna with superstrate with improved bandwidth and gain. II. H-SHAPED ANTENNA DESIGN Fig. 1 shows the top view of the microstrip line fed H-shaped patch antenna. The antenna consists of thick 0.635 mm substrate Arlon DiClad 880 of permittivity of 2.2 and loss tangent tan\u03b4=0.0009. The dimensions of the used substrate are 8x8 mm\u00b2. The dimensions of the main radiating element attached to the microstrip line are 4.4 x 0.6 mm\u00b2 whereas the dimensions for the two extended arms are 3.2 x 0.2 mm\u00b2.The length and width of microstrip line are 4mm and 0.8 mm. respectively. The material used for superstrate in our design is Arlon TC600 of permittivity of 6" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003458_s12239-019-0096-6-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003458_s12239-019-0096-6-Figure2-1.png", + "caption": "Figure 2. (a) 3-D Structure of SLB-EMR; (b) Schematic diagram of SLB-EMR.", + "texts": [ + " On the basis of introducing the structure and principle of SLBEMR, a mathematical analytical model of braking torque considering magnetic flux leakage and end effect is established based on the MEC. The eddy current braking characteristics, power generation characteristics, and natural characteristics of the SLB-EMR were simulated by the FEA. Finally, the design method of SLB-EMR was verified by bench test. The proposed SLB-EMR becomes a part of the trailer bridge, as shown in Figure 1. The SLB-EMR is shown in Figure 2, which consists of the braking system, and the power generation system. The braking system is consists of a rotor with teeth, a stator with water channels, and two independent coils. The power generation system is consists of a permanent magnet synchronous motor (PMSM) and a control system. When the SLB-EMR works, the signals are sent to the control module through the retarder gear switch, the control system enables the power generation system to supply the retarder coils, and a magnetic circuit is generated around the coils, as shown in Figure 2 (b). Meanwhile, the eddy currents is generated on the inner surface of the stator, which forms the braking moment that forces the trailer to reduce speed, due to the stator cuts the magnetic field lines emitted by the rotor rotating with the half shaft. At the same time, the rotor of the generator rotates together with the rotor of retarder, and the power generation system produces a high-speed rotating magnetic field, converts kinetic energy into electrical energy by armature reaction under the action of the rotating magnetic field, and supply power for the retarder coils" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001939_2015-01-0088-Figure11-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001939_2015-01-0088-Figure11-1.png", + "caption": "Figure 11. Gear Shift Fork - Contact Pattern - Stress Results", + "texts": [ + " Contact starts at fork legs that will have a hinge effect before it touches third pad meeting the required deflection. A minimum nominal gap is provided for 3rd pad to share the load when max/abuse load appears on the load shifting Jaw. Gap could be adjusted if the FOS in the legs goes below 1.0 for the Max/Abuse Load as shown in the Fig. 9 and Fig. 10. Loads and Boundary conditions as shown in the Fig. 8. Once the third pad come into contact stress pattern shifts from fork legs to the middle of the fork as shown in the Fig. 11. Stress induced should be below the Yield strength of the material (FOS > 1.0). If not, the web thickness will be increased to meet the requirement. Experimental Verification As per VE Commercial Vehicles Ltd, standard durability duty cycle, rig has been setup as shown in the Fig. 12 and tested the transmission assembly in which Aluminum Gear Shift Fork's (1st & Reverse Fork and 4th & 5th Fork) are the test components. Transmission assembly for operating load condition has been evaluated with defined duty cycles (For up-shifting 2400 and down-shifting 1500 RPM) as shown in the below table 1" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000674_iwem.2019.8887881-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000674_iwem.2019.8887881-Figure4-1.png", + "caption": "Fig. 4. Reflection coefficient including multiple reflection.", + "texts": [ + " The Sparameters of FSSs between two dielectric materials with characteristic impedances and can be calculated as = + + \u2212 1 2+ +2+ + + + \u2212 1 (1) From the Eq. (1), + 1 and + 1 are the constant multiple of . Therefore, the phases of these four ingredients are functions of the transmission phase of FSSs. By 978-1-7281-5006-2 / 19 / $31.00 \u00a9 2019 IEEE eliminating and converting it to of FSSs, the Sparameters are represented as = 2 cos+ \u2212 1 2 cos+2 cos+ 2 cos+ \u2212 1 (2) The reflection coefficient of a multilayer dielectric can be derived by summation of multiple reflection from the interface. Fig. 4 shows how to calculate the reflection coefficient taking into account multiple reflections that occur in each layer. , and are the reflection coefficient at the boundaries 1, 2 and 3 in consideration of multiple reflections. The total reflection | | as was calculated from the lower layer sequentially [3]. The reflection coefficient | | of the whole structure can be computed as functions of the transmission phase of FSS, frequency, and incident angle of the electromagnetic waves. Thus, | | is characterized as the contour map shown in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001662_pcicon.1991.162925-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001662_pcicon.1991.162925-Figure1-1.png", + "caption": "Fig. 1 Typical rotor slot and bar shapes.", + "texts": [ + " The assumption is that all stall times are calculated in the same manner and are equally conservative. Similar machines in the same application but designed by different manufacturers have been found to perform differently and to have different rates of failure or success. Recent developments in calculating rotor bar life analysis [ l , 21 confirm that the stall time calculation is not as simple as just shown but a rather complex calculation affected by the slot design, the stator slot-pitch, and the material properties of the rotor bar [ 8 ] . Figure 1 shows a number of the more popular shapes developed for the bars used in the cages of squirrel cage induction machines. The shapes are designed to produce the different torque and inrush (starting current) characteristics which are requested for the many applications in which induction motors are found. Figure 2 shows the corresponding shape o f the speed-torque characteristic that might be expected for each bar. A general description of each bar shape and the corresponding speed-torque curve will help to understand some of the significant points that will emerge during the course of the paper" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001320_861355-Figure18-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001320_861355-Figure18-1.png", + "caption": "Figure 18 - Toric CVT for DOE hybrid passenger", + "texts": [ + " Tracor Corporation and the Wright Aero nautics Division have separately demonstrated uniqu~ traction drive CVT concepts in several vehicles, but they appear to have abandoned the effort in the late 1970s. Airesearch Corporation, under contract with the U.S. Department of Energy, developed a design for a toric CVT for use in a hybrid pas senger car. Traction Propulsion, Inc., also participated in the project until 1984, when the Department of Energy terminated the contract before the design was fully implemented. A cross section of a transmission representative of that effort is shown in Figure 18 (Reference 20). Vadetec Corporation demonstrated several designs of an ingenious nutating cone-type traction drive CVT designed for automobiles and tractors [21], Figure 19. The exact status of this project is not fully known, but recent comments in the press indicate that the tractor CVT development is continuing at another company under license from Vadetec) while at Vadetec a new kinematic arrangement) more suitable for passenger cars, is being developed [22]. In the last few years) frequent reports of considerable activity on toroidal CVTs have been forthcoming from Japan" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003121_12.826223-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003121_12.826223-Figure4-1.png", + "caption": "Figure 4: Example of a straight Fluted Drill used for Gun Drilling1", + "texts": [ + " Unfortunately, the material delivery lead time was nearly 20 weeks and the cost was high. Relatively, Ti 3.0Al-2.5V (grade 9) cost less, had a slightly better delivery, and an acceptable but significantly lower strength (see table 3). Fortunately, the longest tube needed for the structure was short enough and the outside diameter was small enough to consider even another alternative. Titanium 6.0Al-4V, Grade 5, in rod form with the correct outside diameter and length was readily available. Also, it was determined that gun drilling (see figure 4), an affordable machining process that uses straight fluted drills which allow coolant (either compressed air or a suitable liquid) to be directed through the drill's body, directly to the cutting face1 would be used to hollow out the Titanium rod material; thereby creating a seamless tube. By using gun drilling, the lead time was reduced to 4 weeks and the cost was minimized. 3 DESIGN AND ANALYSIS In addition to designing for performance, the NIRCam strut assembly was designed for manufacturing, assembly, and test" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003923_ecce.2017.8096460-Figure18-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003923_ecce.2017.8096460-Figure18-1.png", + "caption": "Fig. 18 The prototype of the 24s/10p RPM-FS machine.", + "texts": [ + " However, for the machines where Pr=10, 11 and 13, the maximum speed is 4500r/min, 5000r/min and 7800r/min, respectively, which is significantly lower than that of the machine with 14-rotor-poles under the same DC-bus voltage (600V) as shown in Figs. 16 and 17. In addition, the constant power range of the RPM-FS machine with 14-rotorpoles is from 1500r/min to 7500r/min as shown in Fig. 17, which is significantly wider than the other three RPM-FS machines. IV. EXPERIMENTS VALIDATION In order to validate the previous analysis, a prototype of the 24s/10p RPM-FS machine is manufactured and the experiments on back EMF and torque performances are carried out in this section. Fig. 18 shows the prototype machine, and the main design parameters are in accordance with those listed in Table II. Fig. 19 gives the details of the experimental test bench. An induction motor-based dynamometer is used as the variable load, which is controlled by a variable-frequency drive (ABB-ACS800). The toque transducer (HBM-T40) is utilized to measured transient torque, and then the value is displayed on a monitor in the control panel. Fig. 20 shows the open-circuit phase back-EMF waveforms of the 24s/10p RPM-FS machine at 1500r/min" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000005_lawp.2007.914115-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000005_lawp.2007.914115-Figure1-1.png", + "caption": "Fig. 1. Geometry and dimensions of the slot PICA: , , , , , , (50- characteristic impedance), .", + "texts": [ + " It is etched or machined from two metal layers of the PCB. The integration leads to a very low-cost solution because no parts other than the PCB, and hence no assembly, is necessary. The presented antenna is called \"slot PICA\" because its structure resembles the PICA. The slot PICA comprises a leaf-shaped metal structure in one plane, and a larger leaf-shaped slot in a second metal plane. A microstrip feed line connects to the metal structure in the first plane. The second plane acts as ground for the microstrip line, cf. Fig. 1. The leaf-shaped slot in the ground plane results in strong electromagnetic coupling to the feeding structure. The antenna impedance can thereby be controlled by adjusting the slot and feed. The area of the feeding structure is approximately a quarter of the slot area. The distance between the bottom edges of the leaf-shaped slot and the feed is crucial for the impedance matching, in particular at the lower end the frequency range and at the highest frequencies. For characterization, an antenna according to the following is built: The leaf-shaped slot and feed line are etched on a 60 mm 60 mm RO3003 substrate with a thickness , a relative permittivity and loss factor . Dimensions are given in Fig. 1. An SMA connector soldered to the 50- microstrip line serves as antenna port. No external matching circuit is included in the presented design. Simulations and optimization are performed using CST Microwave Studio (MWS) [13], which is based on the finite 1536-1225/$25.00 \u00a9 2008 IEEE integration technique (FIT). Port impedance and radiation patterns are measured in a semi-anechoic room, and radiation efficiency is measured in a reverberation chamber. Port reflection coefficient is measured and simulated up to 30 GHz. The results are seen in Fig. 2. The curves show a marked lowest frequency of operation at 2.2 GHz, and a good match all the way to 30 GHz. The 3.1\u201310.6 GHz frequency band is currently allocated by the Federal Communications Commission (FCC) for UWB radio applications. Therefore, radiation patterns are measured and simulated, at a number of relevant frequencies. Figs. 3\u20138 show results at 2.7, 5.2 , and 11.8 GHz, respectively, relative to the coordinate system in Fig. 1. The presented antenna is symmetrical around the plane, which results in zero -field along y axis in the plane. Although the simulated zero cross polarizations in the plane are not included, as shown in Figs. 3, 5, and 7, it is difficult to achieve the same results in measurements due to the slightly asymmetrical setup and some uncertain reflections. Furthermore, since the antenna is planar, it will by necessity have a pure linear polarization in the plane of the antenna, i.e., the plane. Highly omnidirectional coverage can be seen at all measured frequencies" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002597_j.ijthermalsci.2007.06.020-Figure15-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002597_j.ijthermalsci.2007.06.020-Figure15-1.png", + "caption": "Fig. 15. Oil well cross-sectional view for concentric case including the electric submersible pump flat power cable (dimensions are in m).", + "texts": [ + " Compared to the balanced condition, the peak value of the eddy-current density increases for the 10% and 20%, and 100% imbalance (single-phase) operations by 6%, 12% and 73%, respectively. In addition, it is worth mentioning that single-phase operation generates a pulsating magnetic field according to Eq. (9) and hence higher localized corrosion on the motor casing, see Fig. 1. OPERA-2D is also used to simulate an oil well cross-section with an electric submersible pump cable touching its tubing. This cross-section consists of several layers with different materials which are shown in Fig. 15 and the dimensions are summarized in Table 3. The used electric submersible pump flat power cable is a 3-phase flat cable with the shown configurations in Fig. 16. The different motor operating conditions are shown in Table 4 with Table 4 Simulated line currents in A (rms) Conductor No. Balanced case Single phasing 1 100 0\u25e6 0 2 100 120\u25e6 173.2 0 3 100 240\u25e6 \u2212173.2 0 Fig. 17. Configuration of the oil well with power flat cable. their operating current flowing in the cable for each condition. These results describe the distribution of the magnetic-flux density (B) and the current-density (J ) at the well casing and its tubing" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002010_jae-2008-999-Figure21-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002010_jae-2008-999-Figure21-1.png", + "caption": "Fig. 21. Prototyped mechanical system of the 8.5[mm] diameter RUSM.", + "texts": [], + "surrounding_texts": [ + "Many researchers have been proposed the various kinds of analysis method for the piezoelectric motor considering the contact mechanism. However, the exact analysis has been impossible up to now. Also, there was no arranged design method for the piezoelectric motor. The analysis and design of the piezoelectric motor was using the trial-error-method or using the analysis of vibrator without the consideration of the contact condition. The former is inefficient in the aspect of time and cost and the latter cannot estimate motor characteristic at all. To address these problems, we suggested the analysis and design method of the USM using analytic method combined with numerical method, which was named as the CM in this research. By the proposed analysis and design method, the USM can be designed systematically considering the speed of motor. To verify the suggested analysis and design method and for the demand of small actuators in industry, the 8.5[mm] outer diameter RUSM was developed in this research. Using the prototyped RUSM, it was verified that the proposed analysis and design method was correct by comparing its outcomes with the experimental data. But, the nonlinearity resulting from the heat generation and from the high mechanical force which is applied to the motor was not considered in the proposed analysis method. Consequently, further investigation is required to identify and solve the problems related to the nonlinearity of the USM. For the vibrator of the USM, various kinds of comb-tooth structures were prototyped. Hence, we found out the relation between the speed of the motor and the number of teeth for the first time. It has the significant meaning in the sense that the tooth design method has not been suggested yet. Also, the experiment data showing the relation between the speed and the number of teeth validated the suggested CM. To sum up, the exact analysis of the USM has been impossible and the arranged design method for the USM has not been suggested till now. Hence, it is remarkable that the USM can be analyzed and designed systematically by the proposed method, while considering the contact condition in case of no mechanical force is applied to the motor ideally. It is also important to note that the analysis and design of many other kinds of machines, which use similar mechanism, is possible with the suggested method. Appendix The piezoelectric material is a z-axis poling. The material coefficients are: Mechanical stiffness matrix for constant electric field E: cE = 13.25 6.94 6.46 0.0 0.0 0.0 6.94 13.25 6.46 0.0 0.0 0.0 6.46 6.46 10.52 0.0 0.0 0.0 0.0 0.0 0.0 2.68 0.0 0.0 0.0 0.0 0.0 0.0 2.68 0.0 0.0 0.0 0.0 0.0 0.0 3.16 \u00d7 1010[N/m2] Piezoelectric matrix: e = 0.0 0.0 0.0 0.0 12.82 0.0 0.0 0.0 0.0 12.82 0.0 0.0 \u22126.61 \u22126.61 13.5 0.0 0.0 0.0 [C/m2] Permittivity matrix for constant mechanical strain S: \u03b5S = 7.32 0.0 0.0 0.0 7.32 0.0 0.0 0.0 6.28 \u00d7 10\u22129[F/m2] Density: \u03c1 = 500[kg/m3] Mechanical quality factor: Q = 900. The elastic body of the vibrator is made of the phosphor-bronze. The material coefficients are: Mechanical stiffness matrix: cm = 179.75 96.79 96.79 0.0 0.0 0.0 96.79 179.75 96.79 0.0 0.0 0.0 96.79 96.79 179.75 0.0 0.0 0.0 0.0 0.0 0.0 41.481 0.0 0.0 0.0 0.0 0.0 0.0 41.481 0.0 0.0 0.0 0.0 0.0 0.0 41.481 \u00d7 109[N/m2] Density: \u03c1m = 780[kg/m3] Mechanical quality factor: Qm = 3000." + ] + }, + { + "image_filename": "designv6_24_0002702_4.868039-Figure13-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002702_4.868039-Figure13-1.png", + "caption": "Fig. 13. Compact stabilization technique using a Mesa resistor patch.", + "texts": [ + " 2) The capacitor was treated as part of the input matching network, hence there was no impact on in-band gain. 3) A resistively loaded quarter-wave short-circuited stub was used to resistively terminate out-of-band signals, without reducing in-band gain. 4) The resistive matching at the input stage stabilizes the gate at both in-band and out-of-band frequencies. 5) Series RC stabilization circuits were used at all the bias terminals to suppress low-frequency oscillations. 6) A new method of above-band stabilization that occupied virtually no chip area, as is shown in Fig. 13, was used. It consists simply of an extremely compact (75 m long) open-circuited mesa resistor patch, which has a resistivity of 200 per square. As frequency increases above V-band, the resistive loading also increases, while remaining lumped and therefore wideband to beyond of the device. In addition, potential odd-mode oscillations between the two output HEMTs were analyzed based on the circuit partitioning techniques described in [10]. Parallel clamping resistors were placed at the input and output matching network to suppress any odd-modeoscillations" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000880_0921-5093(93)90209-w-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000880_0921-5093(93)90209-w-Figure3-1.png", + "caption": "Fig. 3. Evolution of casting with total d.C. power input (2214 aluminium alloy billet of diameter 320 ram): (a) conventional casting, P= 0; (b) electromagnetic process, I'= 2.2 kW. Key: a, liquid metal; b, pasty zone; c, solidified metal; d, segregation zone: e, exudations; f, estimated heat flux profile; g, smooth surface; h, height of solidifying metal in contact with ingot mould; i, inductors; j, iron core; n, nuclei.", + "texts": [ + " The effect of the induction electromagnetic field here is identical to that observed in the CREM (casting, refining, electromagnetic) process [7]. The t ime-mean electromagnetic body force ( J x B ) may be resolved into a radial component (principally irrotational) and a vertical component (primarily rotational). The potential forces, balanced by a pressure gradient, result in the formation of a convex surface meniscus, while the rotational forces are responsible for an electromagnetic stirring similar to that encountered in a coreless induction furnace (Fig. 3). As in the CREM process, the action of such vigorous forced convection results in refinement of the grains and a uniformly distributed structure. An example of the penetration of electromagnetic vibrating forces of frequencies 50 and 100 Hz is shown in Fig. 4. It appears that the magnitude of the 50 Hz vibrations is largely predominant, particularly outside the electromagnetic skin depth, i.e. in the bulk liquid. Moreover, methodical measurements allow one to obtain a rough estimate of the peak of the oscillating electromagnetic pressure, which is of the order of 0", + " However, the superior efficiency of the vibrational method is clearly apparent here, where the mean grain size of the equiaxed grains is of the ordder of 150 ~m. This superiority is confirmed by methodical experiments carried out on billets and slabs of 2214 aluminium alloy, which is characterized by a wide freezing range. Moreover, the thickness of the peripheral segregation zone is practically reduced to zero and the surface aspect significantly improved. These effects are explained by the designed modification of the heat flow distribution (Fig. 3(b)). The macrostructures of the materials investigated have been revealed in order to provide information on variations in structure, such as grain size and columnar and equiaxed crystals. To this end, the ingots were sectioned, mechanically polished and then immersed in specific etching solutions. Figure 5 displays macrographs obtained from 1085 aluminium alloy billets of diameter 320 mm. These ingots were produced using a graphite mould and without inoculation of grain-refining master alloys. Because of its relatively high degree of purity (99" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000761_j.jmatprotec.2016.09.004-Figure16-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000761_j.jmatprotec.2016.09.004-Figure16-1.png", + "caption": "Fig. 16. Effective stress distribution of valve forging process.", + "texts": [ + " When mold pressing occurred, the equivalent strain on the plate was maximized at the bottom and at the die to a value of approximately 2 at the point of contact. Indentation forgings were easily generated. Once the billet formation and forging processes were completed, the die corner below the point of contact deformation was small. None of the deformed parts reached a certain level of deformation in the CWR process to ensure that the valve of the microstructure was uniform and that streamline distribution was rational. Fig. 16 exhibits the effective stress distribution in the valve blank forging process. Forging deformation mainly occurred in the initial stage when the metal material was in contact with the mold parts and its adjacent area. Equivalent stress was maximized in the top contact and punch parts of the blank end. Moreover, the strain 10 H. Ji et al. / Journal of Materials Process v u i r c e t d t d d t c w s is 4 mm alue was roughly 267 MPa. The other parts of the body were almost ndeformed. As the top die moved downward, the blank that was n contact with the die surface increased in size" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003039_nt69-a28479-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003039_nt69-a28479-Figure4-1.png", + "caption": "Fig. 4. Fuel-capsule configuration.", + "texts": [ + "\" The term initial implies that the magnitude of this uncertainty will decrease throughout the development design phase of the generator to some finite percentage of the original or initial error. Figure 3 represents a typical peak-capsule temperature as a function of per centage of initial error. This difference in reentry temperature between that value corre sponding to n-standard deviations from the mean or nominal value may be referred to as 100% ini tial error. STRUCTLo\"RAL EVALUATION A typical fuel capsule, illustrated in detail by Fig. 4, consists of three concentric cylindrical tubes with end closures surrounding the nucleus of fuel distributed in a metallic matrix. The inner NUCLEAR APPLICATIONS & TECHNOLOGY VOL. 7 OCTOBER 1969 369 u... 3000 0_ UJ a::: ~ 2800 I-\u00ab a::: UJ 0- rE 2600 I>-a::: I-z UJ 2400 ...., a::: ~ \u00ab UJ 0- 2200 0 20 40 60 80 100 liner immediately surrounding the fuel matrix is primarily a decontamination vessel that allows a clean weld environment for making the final clo sure of the next liner. The next liner is desig nated as the structural member and is assigned to resist the internal pressure resulting from gen erated helium at elevated temperature" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002266_lmwc.2006.879484-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002266_lmwc.2006.879484-Figure1-1.png", + "caption": "Fig. 1. Schematic of the new cross-slot-coupled DRA. Inset: centered design in [6].", + "texts": [ + " In these cases, the complicated shapes of the DRAs, the metallic strips or even the arrays of simple-shaped DRAs increased their manufacturing difficulty and thus their cost and in some cases also their size and weight. In [6], another design was proposed, in which CP operation was accomplished by using two crossed slots of unequal lengths to couple energy from a microstrip line to a simple cylindrical DRA. Both slots were angled at 45 with respect to the feeding microstrip line and their centers were at the same position, centered underneath the dielectric resonator (DR), as depicted in the inset of Fig. 1. Their lengths were unequal, so that two near-degenerate orthogonal modes of equal amplitude and 90 phase difference were excited at frequencies close to that of the fundamental [8] mode of the DRA. The resulting CP bandwidth was 3.91%. This letter proposes a new design of a cross-slot coupled DRA, taking into consideration that the slot and the DRA resonances are two partially independent mechanisms [9]. The new design is illustrated in Fig. 1. Each slot is resonant at a particular Manuscript received February 6, 2006; revised April 26, 2006. This work was supported by ETH Research Grant TH-38/04-1. The authors are with the Laboratory for Electromagnetic Fields and Microwave Electronics, ETH Zurich, Switzerland (e-mail: almpanis@ifh.ee.ethz.ch). Digital Object Identifier 10.1109/LMWC.2006.879484 frequency, which depends on its length as well as the permittivity of sub- and superstrate (DRA). The 90 angle crossing of the slots prevents the coupling between the two excited orthogonal modes and therefore, the investigation for each slot resonance may even be conducted separately. CP operation for the DRA is obtained when the two modes have equal amplitudes and 90 phase difference (for right-handed or left-handed circular polarization). The amplitude of the orthogonal modes is affected by the matching of each slot and consequently also by the stub length (distance between the center of the slot and the open-end of the microstrip line as depicted by in Fig. 1). Therefore, excitation by two crossed slots may provide CP operation over a significantly expanded bandwidth, provided the following requirements are satisfied. First, the length of one slot must be adequately larger than that of the other in order to excite modes at different frequencies. Second, a different stub length of the feeding microstrip line must be chosen for every slot so that better matching can be ensured and thus near-degenerate modes will be excited. As discussed before, the slot lengths and the permittivity of the sub- and superstrate (DRA) determine the frequencies of the slot resonances, while the DRA modes depend on the DR dimensions, permittivity, as well as the feeding mechanism", + " However, in order to obtain optimum CP operation for the DRA, its radius might need to be altered slightly. Further fine-tuning might still be necessary, in order to enhance the CP bandwidth without deteriorating the return loss or the radiation patterns. This optimization involves mainly the stub lengths and the position of the disc center relative to the center line of the feeding microstrip. For a CP operation of the DRA at 5.7 GHz, the aforementioned procedure results in an antenna geometry, as illustrated in Fig. 1. The dielectric disc is made from Rogers TMM 10i laminate, with dielectric permittivity 9.8, height 5.1 mm and radius 15.2 mm. The DR lies on top of the crossed slots which are etched in the finite ground plane of a carrier substrate ( 2.2 and thickness 0.7874 mm) with dimensions 120 mm 120 mm. The center of the DR is exactly on top of the center of the slot with length . This substrate carries on the opposite side the 50 microstrip line with 2.4 mm. The dimensions of the crossed slots are 11.4 mm, 8" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000092_8.247750-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000092_8.247750-Figure1-1.png", + "caption": "Fig. 1. Microstrip antennas on different faces of a wedge.", + "texts": [ + " Mutual couplings between a circular MSA and a phased array which has 128 circular MSA elements located on different faces of a polyhedron are measured. The measured coupling level due to beam steering of the array and the level calculated by the theory are in good agreement . In Section V several numerical simulations which describe some properties of the antenna coupling are given. 11. THEORY First of all, we derive an approximate formula for mutual coupling between two linearly polarized circular MSAs separated by a wedge. Figure 1 shows the geometry of two circular MSAs located on different faces of a wedge. The radiated magnetic field H , from the equivalent magnetic current element .Im1 of MSAl is given as [2] H , = VV .A, , / jwq,p, , - jwAml , where A,, is the magnetic vector potential for a current element. A,, is given by the cavity model [2] as .J,1 = U 1 ( k 1 ~ Z ) K m ~ = A71(k,,Z)cos (41 - 410)i+~, (3) where K is a constant, +1 is the angle of a current element, and +lo is the location angle of the feeding pin of the MSA" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001203_cencon.2015.7409516-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001203_cencon.2015.7409516-Figure3-1.png", + "caption": "Figure 3. Slotted Radial Configuration", + "texts": [ + " The choice of bigger diameter tends to have better performance due to the flywheel effect of that of the conventional machines. With the choice on the radial flux flow it takes longer flux path and if the flux flow is shorten and somehow the yoke flux could be better utilized the utilization factor in the axial is better than that of the radial configurations. Axial derive a lower weight and higher utilization factor. Ideally the short magnetic flux brings more torque density with respect to the amount of electrical and magnetic loading in the circuit. Figure. 3 and Figure 4 shows the radial and the axial configurations topology. The output power (P) for any electrical machine is expressed as in Eq. (1). (1) where : efficiency, : number of phases T : Torque exerted by the machine is the torque constant. The peak value of the air gap phase EMF for the axial flux machine is as in Eq. (2). (2) where effective stack length, : is the ratio of air-gap diameter to the outer diameter , f: frequency of rotation, p is the number of magnetic poles. The peak value of the air gap phase EMF for the axial flux machine is as in Eq" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000108_s1672-6529(11)60006-1-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000108_s1672-6529(11)60006-1-Figure7-1.png", + "caption": "Fig. 7 Concept of relative rotation angle.", + "texts": [ + " Here, the non-dimensional vertebral joint length is defined as the percentage of absolute length to total body length. As compared with Fig. 5, the difference of non-dimensional vertebral joint length between two groups is not significant, which implies that the individuals have little effect on non-dimensional vertebral joint length. The overall direction of forward travel (parallel to the flow and long axis of the tank) is indicated by x, while \u03b8 represent the angle between two adjacent midline segments Seg(i) and Seg(i+1) (solid lines in Fig. 7). Note that the carangiform swimming is a mode of BCF propulsion, in which the large amplitude of the undula- tions is mostly restricted to the one-half or even one-third posterior part of the body and increases sharply in the caudal area. In addition, as the length of each vertebral joint is rather short, it is difficult to accurately measure the kinematic parameters of each adjacent joint. Hereby, to meet the requirements of bionics, we divided ten markers on the dorsal of crucian into five consecutive segments (Seg(i), i = 1 to 5), where Seg(i) is the segment from markers (i\u22121) to i (refer to Table 1)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001410_tmag.2013.2238616-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001410_tmag.2013.2238616-Figure4-1.png", + "caption": "Fig. 4. The definition of position angle.", + "texts": [ + " It will be described in detail in the next section. Fig. 3 shows the assembled spherical motor. Also Table I shows the specification of the spherical motor. B. Definition of Motion Angles The spherical motor has 3-DOF and the position of rotor is variable on 3D space. Therefore, before we define the current functions and do experiment on the spherical motor position control, the position angle of rotor should be defined first. In this paper, the position of rotor is defined by the rotating and tilting angles. Fig. 4 shows the definition of the rotating and tilting angles as rotor position. Fig. 4(a) and (b) describe the concept of the rotor and stator. Especially, the stator coils are described in detail. The stator of the spherical motor consists of 6 segments as in Fig. 2, and each segment has a rotating coil and a tilting coil. The coils in each segment are named A to F respectively. Also the rotating coil is represented by subscript \u2018r\u2019, and the tilting coil is represented by subscript \u2018t\u2019. For example, coil \u2018 \u2019 refers to the tilting coil of segment B. The upper coil on x-y plane is named by subscript \u2018u\u2019, and the lower one is by subscript \u2018l\u2019, thus, the tilting coil \u2018 \u2019 consists of coils \u2018 \u2019 and \u2018 \u2019. These two coils are wound in a opposite direction and are connected in series. Therefore, the current of tilting coil generates the magneto-motive force in the opposite direction in each coil. The tilting movement is realized by current in the tilting coils. Fig. 4(c) and (d) show the rotor and stator relation when the rotor is tilted. The position relation can be described by using the tilting angles and . In Fig. 4(c), is the angle between the shaft and z-axis. It is related to the magnitude of the tilting current. In Fig. 4(d), is the angle between the x-axis and the projection of shaft to the x-y plane. The resulting magneto-motive force of the tilting coils is related with . Fig. 4(c) shows that the tilting coils that are mounted on the other side of the sphere such as \u201c \u2013 \u201d are excited by an opposite direction current which either generates a magnetic flux headed outward or inward form the surface of the sphere. In Fig. 4(e), is the rotation angle. The rotor is rotated by a 3 phases current in the rotating coils, which is identical to the motion of 3 phases synchronous motor. To comprehend the spherical motor motion, the angles in spherical motor should be defined. Moreover, in the control of 3-DOF system, the definitions of angles are essential. There are many methods to control the spherical motor. In this paper the reference current method is chosen and the current for 3-DOF motion is calculated. The currents in the rotating coils are identical with that of 3-phase, 4-pole rotary motor" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000761_j.jmatprotec.2016.09.004-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000761_j.jmatprotec.2016.09.004-Figure6-1.png", + "caption": "Fig. 6(a) shows the fractured workpiece in the experimental condition when the forming angle ( ) was equal to 32\u25e6 and the stretching angle ( ) increased to 5\u25e6 with no mandrel. An enlargement of the inner bore of the wedge occurred in every rolled piece in the experiment. As shown in Fig. 6(b), the bore at which wedging was initiated showed the largest diameter. The diameters of the bores on both sides decreased gradually until the same diameter was reached. In this study, our mold was designed to have different stretching angles. Hence, the stretching angles from the wedged-in section to the stretching section were set with different values. Using a mandrel optimized the hole in the rolling process. The method that effectively solved the enlargement of the inner bore was then selected. The final effect is shown in Fig. 6(c).", + "texts": [], + "surrounding_texts": [ + "H. Ji et al. / Journal of Materials Processing Technology 240 (2017) 1\u201311 3\ne a l i e\n2 h\n2\np T n fl c c r i c u t\n\u03b5\nw r v t e\nm 9 1 o C 1 r e\n\u03b5\np t t l O d t o v t w\nThe present study proposes a new hollow valve process. A 3D lastic-plastic FE model of CWR-forging is developed to evalute this process. The validity of this method is verified through a aboratory experiment. Fundamental formation characteristics are nvestigated on the basis of this reasonable 3D FE model. Finally, the ffect of a critical process parameter is determined numerically.\n. Numerical modeling of the chain process for producing a ollow valve\n.1. Cross wedge rolling process\n4Cr9Si2 is the main material used in valves. It is sensitive to temerature and has a narrow heat deformation temperature range. he chemical compositions (wt.%) are shown in Table 1. The Arrheius equation is widely used to describe the relationship among ow stress, strain rate, and temperature under high temperature onditions (Zhang et al., 2012). Huo et al. (2015) used a new appliation of unified constitutive equations for the CWR of high-speed ailway axle steel. The constitutive equations were implemented n the commercial FE software DEFORM-3D. The CWR process was haracterized by significant deformation. In the present study, we sed the Kumar model to describe the following constitutive relaion (Eq. (1)):\n\u02d9 = A [ sinh (\u02db )n ] exp( \u2212Q RT ) (1)\nhere A and are the material constants, n is a constant closely elated to the strain rate, is the flow stress (MPa), R is the uniersal gas constant (8.31 J mol\u22121 k\u22121), \u03b5\u0307 is the strain rate (s\u22121), T is he deformation temperature in Kelvin (K), and Q is the activation nergy of deformation (J mol\u22121).\nHot compression tests were conducted on a Gleeble-1500 theral mechanical simulator under the forming temperatures of\n50 \u25e6C, 1000 \u25e6C, 1050 \u25e6C, and 1100 \u25e6C. The strain rates of 0.01, 0.1, , and 5 s\u22121 were set. The true stress-strain curve could then be btained. Fig. 2 shows the grain size of 4Cr9Si2 after annealing. ylindrical specimens with a diameter of 8 mm and a height of 5 mm were machined from the wrought billet. The constitutive elation of material was determined through linear regression. The quation is written as follows (Eq. (2)):\n\u02d9 = e31.25[sinh(0.00758 )]4.9535 exp(\u2212382980/RT) (2)\nThe geometry model shown in Fig. 3(a) consists of a roll, workiece, guide plate, and mandrel. The diameter of the mandrel is less han that of the workpiece. The former is also related to the goal of he workpiece, wall thickness, and the reduction in rolling the holow billet mill with respect to the determination of mandrel size. nce the size of the center is symmetrical after rolling, the manrel diameter should correspond to the size of the inner hole of he rolled hollow billet with a corresponding volume. The mandrel ptimizes the hole in the rolling process. Fig. 3(b) shows the control alve of the CWR process for the hollow mandrel formation. Prior o establishing the FE model of CWR, the following assumptions ere made:\n1) The tools and guide plates are rigid. Given the negligible elastic deformation of tools and guide plates, two wedged tools and guide plates were considered to be rigid models for the tool materials. 2) A billet is a plastic material model. During formation, a billet is characterized by significant plastic deformation. The extent of\nelastic deformation is slight; thus, it can be disregarded.\n3) A roll rotates at a fixed step. In practical rolling, roll speed first increases rapidly, stabilizes to a constant, and then decelerates.", + "4 H. Ji et al. / Journal of Materials Processing Technology 240 (2017) 1\u201311\n4\n5\nf\nw t\ne s\n2\nw n T s d o a ( f d\ni\n) Only half of the model was simulated to maintain asymmetrical CWR and to reduce CPU processing time. ) The friction coefficient was assumed to be constant throughout the CWR process regardless of whether cooling water was used. The frictional force in the shear friction model is defined by (Eq. (3)):\ns = m \u00d7 k (3)\nhere fs is the frictional stress, k is the shear yield stress, and m is he friction factor.\nTable 2 provides a summary of the adopted simulation paramters of the CWR process. The main process parameters for imulation are listed in Table 3.\n.2. Forging process\nThe geometry model illustrated in Fig. 4(a) consists of a top die, orkpiece, and bottom die. The forging die in CWR has been used in\necks for locating datum. Neck size is unchanged during formation. he hollow valve stem that is connected to the neck parts of the urface is the cone. The angle of the valve cone is 4\u25e6. The forging ie is positioned on the conical surface, thereby reducing the size f the workpiece during axial channeling formation. Moreover, the ngle of the cone ( ) is equal to that of the workpiece, die radius R) is greater than the target product neck radius of 1\u20132 mm. These actors enhance the formation of the workpiece in place. Fig. 4(b)\nepicts the control valve employed in hollow valve forging.\nThe summary of the adopted simulation parameters of the forgng process is listed in Table 4.\n3. Experimental tests on the chain process for producing hollow valves\nExperimental tests on the CWR process were conducted using the H500 mill and electric tube furnace at the University of Science and Technology Beijing in Beijing, China. This equipment is displayed in Fig. 5(a). Fig. 5(b) shows the forging mill. The forging experiment was conducted at Huaiji Dengyun Auto-Parts Co., Ltd. in Guangzhou.", + "H. Ji et al. / Journal of Materials Processing Technology 240 (2017) 1\u201311 5\nF m\nT d i 4 f s r w i p p\nf s r f d i o p r p a s s p F t g\nig. 5. Machine for CWR-forging used in the experiments, (a) H500 mill (b) forging achine.\nhe basic capture cross section was between 4.5 and 4.6 mm in iameter, and the size distribution of this section was uniform. The\nnner hole at the center of the symmetry measured approximately .7 mm. The symmetrical center was cut open for forging; thereore, subsequent processing was not affected when the hole of the ymmetrical center was large. In the deformed 3D simulation of two olls, the gap was adjusted by 10.5 mm, and the mandrel diameter as 4 mm. Hence, the experimental and FE results differed slightly n terms of diameter, although their inner holes were similar. This henomenon is depicted in Fig. 7(b). The experimental results are resented in Table 5.\nTo analyze metal flow, four rolling cross sections were derived rom neutral planes that measured 5, 10, 20, and 30 mm in this tudy. Eighteen points were selected for analysis; the sequential olling surfaces of these points were inclined toward another surace (Fig. 8(a)). Fig. 8(b) shows that the 18 points on the axial isplacement of the cross section were not equal. The characterstics of cross sections a, b, and c were detected near the minimum f the center area. When the radius direction increased gradually to oints 4 and 15, the displacement decreased until the surface was eached. The friction induced by the die on the surface of the rolled iece induced the axial displacement of the outermost layers. The xial displacement in Section d was basically the same because this ection is located at the end of the rolled piece. Different sections howed varied axial displacements; therefore, the farther the rolled iece-up wedge was positioned, the greater the axial displacement. ig. 8(c) shows the axial position of 18 points. Fig. 8(d) shows he initial longitudinal single grid, and Fig. 8(e) shows the sinle longitudinal grid after deformation. The parallel element mesh\n(d = 4 mm).\ndeformation also occurs. The axial flow of the material increases from the inner wall to the outer wall, and the difference increases as rolling length increases. Therefore, axial inhomogeneous deformation is significant in the CWR of hollow parts. Uneven deformation and local deformation are mainly attributed to surface friction.\nFig. 9 exhibits a CWR-stretching strain distribution at 1.5 s. The stretching stage was a stable rolling stage in which the strain changed smoothly and the radial metal was compressed. The radial strain was compressive and was maximized in the entrance side as the tensile strain state. The tangential strain was dominated by compressive strain, that is, the deformation zone located near the deformation of the metal pull produced tensile strain. The axial strain was tensile. The deformation zone generated by the squeeze pressure strain of the die wedge indicated that when the inner metal experienced maximum tangential and axial strains, the outer layer gradually decreased in size. This finding suggests that in workpiece deformation, metal flow is hindered when the penetration capability of the stretching segment is enhanced relative to the inner surface of the outer metal. Metal flows freely when the stretching strain of a circular uniform distribution decreases gradually from the surface to the inner surface of the deformed workpiece. The surface is maximized.\nFig. 10 depicts CWR-stretching stress distribution at 1.5 s. The change in cross-section stress was more stable than the deformation force of the stretching of the knifing zone given the high permeability. Stress anisotropy was maximized in the workpieces that came into contact with the die parts. The workpiece deformation region of a cross section was elliptical. The radial stress ran along the radial direction period because contact with the mold surface gradually dropped to the inner layer at this time. The presence of an ellipse and of surrounding metal reduced the tensile stress at the left and right sides of the local metal. The tangential stress at the outer layer was compressive, and that on the metal in the inner layer was tensile. Thus, tensile stress decreased slightly in the direction of the mold. When the stress on the inner and outer metal layers was compressive, the stress on the outer metal was lower than that on the metal plate in the inner layer. Furthermore, the tangential stress distribution was small in a wedge section of a uniformly elliptic degree. As depicted in Fig. 10(c), the mold under the direction of the workpiece and die, the stress near the contact zone and metal mold wedge, and the inner surface of the workpiece were not fully compressed. The axial flow of the outer metal in the metal surface was under tensile stress. Horizontally, the tensile stress on the outer metal in the vicinity was subjected to elliptical movement, and the inner layer metal was subjected to limited compressive stress." + ] + }, + { + "image_filename": "designv6_24_0001203_cencon.2015.7409516-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001203_cencon.2015.7409516-Figure8-1.png", + "caption": "Figure 8(a) shows the rotor surface design and the quarter section of magnet placing is shown. Figure 8(b) shows the coil surface base and the coil groupings. Figure 8(c) shows the assembled sandwiched model of the stator and the rotor. Figure 8(d) brings the sandwiched configuration with end capping on both the sides. End turns are taken from the coil leads for connecting to the external electrical network. The developed model is based on the detailing as shown in Table III.", + "texts": [], + "surrounding_texts": [ + "Based on the design equations and the magnetic circuit a design approach to the proposed structure is as shown in Figure 7." + ] + }, + { + "image_filename": "designv6_24_0000280_9781119078104.ch3-Figure3.5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000280_9781119078104.ch3-Figure3.5-1.png", + "caption": "Figure 3.5 Illustration of N-slot rectangular patch", + "texts": [ + " As a number of slot resonators can be accommodated in a small footprint, they are the most space-efficient [10]. For a slot resonator with a physical length L0, the resonant frequency or the frequency whereby the signature will be produced can be predicted using Equation 3.3 [11]: fr = c 2L0 \u221a 2 1 + \ud835\udf00r (3.3) where c denotes the speed of light and \ud835\udf00r denotes the relative permittivity of the substrate. A rectangular patch with N number of concentric U-shaped slots has been investigated in this research for data encoding. Figure 3.5 shows a generic illustration of a rectangular patch with N slots. The slots S1, S2, S3, \u2026 , SN have a total physical length of L1, L2, L3, \u2026 , LN, and from Equation 3.3, each slot will resonate at distinct frequency. As the patch is illuminated by a plane wave, it shows a corresponding resonance in the backscattered RCS spectrum. Each slot will produce a magnitude \u201cdip\u201d resulting in N number of resonances in the frequency spectrum (Figure 3.5). Depending on the tag design, this multislot rectangular patch can encode data in two configurations: (i) binary encoding and (ii) frequency shifting. Binary Encoding In this configuration, each slot resonator represents a single data bit. The number of slots has 1:1 correspondence with the number of encoded bits. The presence or absence of a magnitude \u201cdip\u201d or \u201cnull\u201d is represented by \u201c0\u201d or \u201c1\u201d. This means that an N-slot patch can encode N bits of data. Binary encoding has been studied comprehensively in Ref" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001572_tmag.1981.1061290-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001572_tmag.1981.1061290-Figure5-1.png", + "caption": "Fig. 5 Coil Geometry", + "texts": [], + "surrounding_texts": [ + "1942 IEEE TRANSACTIONS ON MAGNETICS, VOL. MAG-17, NO. 5, SEPTEMBER 1981\nBOW-SHAPED TOROIDAL FIELD COILS\nPeter Bonanos\nIXTRODUCTION The maximum f i e l d which can be produced i n a\nt o r o i d a l c o i l is usually determined by a l imi t ing t e n s i l e s t r e s s r e a c h e d a t h e i n n e r major adius R , . In Ref. 1 a structural rrangement, termed a \"bow c o i l , \" was described which per imi ts a reduction in the tension a t the inner radius. A f u l l e r and more general descr ip t ion is presented here and compared with other forms of toroids. Performance limits are estimated and useful engineering parameters are tabulated. A simple s t r u c t u r a l a n a l y s i s is also included to demonstrate the mechanical behavior o\u20ac p r a c t i c a l c o i l s .\nGENERAL MECHANICS The t o t a l v e r t i c a l f o r c e on each alf-coil of a\nt o ro ida l a r r ay is given by\nwhere I is t h e c o i l c u r r e n t and B,R, the f ie ld- rad ius product a t some the co i l shape . in te rna l po in t . &\"is independent of\nThe sum of the reac t ion forces P I and P2 must of course, equal P. Their respective values however a r e funct ions of t h e c o i l a s p e c t r a t i o A equal to (R1 + R 2 ) / ( R 2 - R1) and the s t ruc tura l Suppor t arrangement. Since the cross-sectional area vailable t o support tension is l e s s a t R1 t h a n a t R2, methods of The author is w j t h Plasma Physics Laboratory, Princeton University, Princeton, NJ 08544, USA\nreducing f, E P,/P a re sought . For c i rcu lar co i l s having usefu l aspec t ra t ios th i s f rac t ion has a range 0 .6 ? f 7 0 .7 . For cons tan t t ens ion co i l s fp = 0.5.\nThe s t ructural arrangement shown in Fig. 2 can P\nfurther reduce f P'\nA t ens ion l ink is a t t ached to each c o i l a t g with inc l ina t ion angle 0 and tangent to the co i l . The l i n k s a re a t tached to a cent ra l cy l inder which t ransmits a compressive force C directed towards the midplane. The compression C reduces the co i l t ens ion a t R, and, through t e l i n k s , increases the t ens ion a t R2. I\u20ac viewed i n a l i t e r a l ly g loba l s ense , t he equa to r i a l p ressure in the to ro id i s used t o produce compression from pole to pole .\nBOW C O I L SHAPES A coi l shape can be generated using connected,\nconstant ension, arc segments if\nwhere p is the rad ius of curvature , T the t nsion and 13 the load per unit length of a r c ds . Using the Chin-shel l approximation f rtheoidal f ie ld , B = BoRo/R, the uni t load i s p = I BoRo/2R. For the outer arc oE t h e c o i l a t R2 where the t ens ion i s P? = ( 1 - fp)P the curve, using Eqs. ( 1 ) and ( 2 ) , is glven by\nds/da = (1 - f p ) R An (R /R 1 . 2 1\n( 3 )\n0018-9464/81/0900-1942$00.75 0 1981 IEEE", + "1943\nThe tension within the inner, curved portion.of the coil (between g and R1) is T.. The tension in the portion of the cod at k, is then link is then (P2, - T. ) . The %orce in the straight\np1 = f P = T . - (p2 - T ~ ) sine . P 1\n( 4 )\nThe shapes for constant 8 and varying f are shown Fig. 4. P\nCombining ( l ) , (2), and ( 4 ) the inner curve of the coil is given by\nand the link tension is\nSince coil stresses are constant for a given field strength and aspect ratio it is convenient to normalize the geometry using a unit average radius R = (R1 + R2)/2 = 1 and the inverse aspect ratio ( ? . e . , toroidicity) E = (R2 - R1)/(Rl + R2) = 1/A. The shapes for constant fp and varying 8 are shown in Fig. 3 .\nin", + "1944\nThe inner cross section Ai is partitioned such that 1 / 3 Ai is devoted to insulation and coolant. The remainder is available to support forces and the sum for all coils is equal to 8 n ~ ( 1 - & ) / 3 where % is the opacity defined as nA0/27cR2h is taken to be 1 / 3 (the radial half-thickness. At the outer radius R2 an figure is drawn to scale). The area for insulation and coolant is assumed constant hroughout he coil length. The total net area at R2 is then Using BR = p0n I/2x (MKS) and converting ( 1 ) to n(A, - Ai/3) = 8% E &h/3* the present notation, the vertical coil force sum is\nA tensile force corresponding to 100 MF\u2019a (14.5 ksi) on the net area is now allowed. If the field strength is Limited by the inner cross section (at R1) then, from (7), the maximum value at Ro is\nIf limited by the net cross section at R2 then\nThe total field energy W and inductance L are related by\nIf a thin shell is again assumed then, for Ro = l m ,\nand the single-turn inductance (MKS) is given by\nL = - \\ ZR henries. Po I + &\ni-& R\nFour coil shapes are now compared: ( 1 ) Circular Coils, ( 2 ) Constant Tension Coils, ( 3 ) Constant Stress Bow Coils, and ( 4 ) Demountable BOW Coils. For the circular coils a support scheme is assumed such that the in-plane moments are negligibly small and the point of zero moment is at R2 (or nearby) so that f may be estimated. A low moment is admittedly d5fficult to achieve but toroids such as TFTR attempt this ~bjective.~ For the constant stress bow coil the tensile stresses at R1 and R2 are set equal (but the effective Stresses u = 2 -cmqx are not). The necessary fraction f is then a functson of the net cross sections and the tgridcity. For the assumed packing fraction and transparency f = 1 - E. For the demountable toroid fp is set equal to - 0.1. The coil joints are assumed to be in the straight portion of the coil at R1 and 1 0 % of the total vertical force is applied compressively. P\nThe results are shown in Fig. 6 (next page) for toroidicities 0.5 < < 0.8. and radial halfthicknesses ch = 0.1, 0.2 and 0 . 3 . For the bow coil examples R is always 4 5 O . Note that the \u201cD\u201c coil and constant stress bow coil are identical for = 0.5. Representative values are also listed in Table I.\nSTRUCTURAL ANALYSIS Two practical problems are now addressed ( 1 ) The\neffect of deformations (in both the coil and structural supports) on the internal force distribution, and ( 2 ) A realizable structure quivalent o the schematic arrangement shown in Fig. 2 The cylinder and pinned link configuration is difficult to construct due to the congestion where the links converge. The inverted arrangement shown in Fig. 7 is simpler. The coil structure is extended axially in the shape of a triangular blade ( B ) and a massive ring (A) restrains the radial motion. The blade then pushes towards the midplane at R1. A coil shape was geperated for = 0.7 ch = 0.2 and f = - .2. A structural model is shown in Fig. 8 . &e analysis, using $ structural code,6 indicates that distortions reduce the comprewive force at R1 to\n= - 0.17. All forces are within - 5% of the predscted values. fP ." + ] + }, + { + "image_filename": "designv6_24_0002311_cpe.2016.7544196-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002311_cpe.2016.7544196-Figure1-1.png", + "caption": "Figure 1: Stator slot geometry of an in-wheel motor with four conductors", + "texts": [ + " Knowing Cws in the early stage of the motor development is essential for predicting the common-mode current, which causes the electromagnetic interference (EMI). An influence of temperature is also discussed, since it is important parameter for the evaluation of Cws. At the end, the simulation results of common-mode current at two different temperatures, using simplified RLC circuit, is presented. II. FEM SIMULATION OF THE SINGLE STATOR SLOT The in-wheel motor has 168 slots in the stator, with four wires (solid conductors) in each slot. The geometry of a single stator slot is presented in Fig. 1. The slot height and width are 11.5 mm and 2.6 mm, respectively. There are four parallel solid conductors (wires) with a nominal diameter of 2.12 mm, which are insulated with 0.07 mm wire insulation. Between the stator lamination pack and conductors there is an insulation sheet with a thickness of 0.15 mm. Empty space in the slot is filled with epoxy resin. A 2D electrostatic finite-element method (FEM) simulation has been made in Cedrat Flux software in order to calculate the capacitance per unit length cws [6]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000761_j.jmatprotec.2016.09.004-Figure15-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000761_j.jmatprotec.2016.09.004-Figure15-1.png", + "caption": "Fig. 15. Effective strain distribution of valve forging process.", + "texts": [ + " When the overall temperature ranged from 1000 \u25e6C to 1180 \u25e6C, the dynamic recrystallization microstructure could be determined easily. Valves were formed in the final stage to ensure full disk size, metal flow, and angular clearance. A punch effect was generated under the pressure of a small amount of metal to axial flow. The flow was resisted, and the deformation was complicated. Deformation energy was converted into heat, and additional blank disks were generated. H. Ji et al. / Journal of Materials Processing Technology 240 (2017) 1\u201311 9 f s a m v f t t T r T w Fig. 15 displays the effective strain distribution in the valve blank orging process. Forging deformation mainly occurred in the initial tage when the metal material was in contact with the mold parts nd its adjacent area and when the equivalent strain was maxiized in the top contact and punch parts of the blank end. The strain alue was approximately 0.6. The other parts were almost undeormed. As the punch moved, the blank that was in contact with he die surface increased along with the axial deformation resisance" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001924_sp.j.2096-5796.2019.0016-Figure15-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001924_sp.j.2096-5796.2019.0016-Figure15-1.png", + "caption": "Figure 15 (a) Self-decoupled mechanical structure of the 4 DOF F/T sensor; (b) Prototype of the 4 DOF F/T sensor.", + "texts": [ + " The inter-dimensional coupling problem is main factor limiting measuring accuracy of multi-dimensional F/T sensors, so we must decouple the output signals to reduce or eliminate coupling errors. Generally, there are two decoupling methods for multi-dimensional F/T sensors. One is hardware decoupling method, which starts with structural design and mechanical manufacturing process to eliminate the coupling errors. But this method is difficult to implement and greatly increases the cost of sensor manufacturing. Song et al. developed a novel four-dimensional F/T sensor with a novel self-decoupled mechanical structure for human-computer interaction[39], seen in Figure 15. The experimental results 2019\u5e74 \u7b2c1\u5377 \u7b2c2\u671f\uff1a121\u2014135\u865a\u62df\u73b0\u5b9e\u4e0e\u667a\u80fd\u786c\u4ef6 Virtual Reality & Intelligent Hardware demonstrated that the four-axis F/T sensor with self-decoupled structure have the maximum measurement error of 1.5% Zhao et al. proposed a mechanical decoupling method for parallel three-dimensional force sensors, which uses rolling friction instead of sliding friction to reduce coupling[40]. Wu et al. designed a sixdimensional force sensor with self-decoupling structure[41], but precise slip clearance and groove are needed, and the contact force between elastomer and groove side wall will produce additional coupling" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002224_3527608117-Figure1.9-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002224_3527608117-Figure1.9-1.png", + "caption": "Fig. 1.9 Direct and indirect squeeze casting.", + "texts": [ + " In gas pressure infiltration the response times are clearly longer than in squeeze casting, so that the materials must be carefully selected and coordinated, in order to be able to produce the appropriate composite material for the appropriate requirements. Squeeze casting or pressure casting are the most common manufacturing variants for MMCs. After a slow mold filling the melt solidifies under very high pressure, which leads to a fine-grained structure. In comparison with die-casted parts the squeeze-casted parts do not contain gas inclusions, which permits thermal treatment of the produced parts. One can differentiate between direct and indirect squeeze casting (Fig. 1.9). With direct squeeze casting the pressure for the infiltration of the prefabricated preforms is applied directly to the melt. The die is thereby part of the mold, which simplifies the structure of the tools substantially. However, with the direct procedure there is a disadvantage in that the volume of the melt must be determined exactly, since no gate is present and thus the quantity of the melt determines the size of the cast construction unit. A further disadvantage is the appearance of oxidation products, formed in the cast part during dosage" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003553_iciap.2003.1234118-Figure2.1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003553_iciap.2003.1234118-Figure2.1-1.png", + "caption": "Figure 2.1: Mantis head camera model. The surface, which cross-section, is given by ( )z g r= , is viewed by a camera with focal point moving along the X-axis. \u03c1", + "texts": [ + " In this paper we develop a mathematical model of the Proceedings of the 12th International Conference on Image Analysis and Processing (ICIAP\u201903) 0-7695-1948-2/03 $17.00 \u00a9 2003 IEEE biologically motivated visual-motor system for depth estimation; we describe an implementation of the system and experimental environment; we present and discuss the performance of the system and provide experimental results and error analysis of the algorithm; we present the conclusions and propose potential usage of the system in mobile robot environment. Figure 2.1 illustrates the process. The camera moves left and right (pure translation) along the X-axis according to the function )(tcc = , where we set )0(cc = . Typically this motion is with constant speed (and changing direction at the edges of the platform) such as 0( ) * *c s V\u03c4 \u03c4= , where s is 1 or \u20131 depending on the peering direction. denotes the displacement along the X-axis from the CCD center on the image plane at which a feature is projected, and 0 ( )r \u03c1 and ( )tr \u03c1 are the displacements where points observed at the displacement \u03c1 on the image plane are located on the surface of the object" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000343_isitia.2018.8710966-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000343_isitia.2018.8710966-Figure4-1.png", + "caption": "Figure 4 The realized UAV", + "texts": [ + "01, and derivative gain of 4. IV. EXPERIMETAL RESULT Several tests have been conducted to test whether the design has worked well. Airplane stability testing has been performed by testing pitch and roll angles of stability control, PWM control signals, altitude control, and GPS navigation control. Flying tests have also been conducted to test aircraft stability and automated testing of the navigation system with GPS waypoints entered. All components have been integrated and the aircraft results are shown in figure 4. This test is intended to determine the response of aircraft stability control based on pitch and roll angles. The resulting pitching motion is the response of an angular error to the 0 \u00b0 pitch angle setpoint of the IMU. While the rolling motion is the response of angle error to setpoint 0 \u00b0 roll. The control response to the aircraft stability control is the result of accelerometer angle combined with the angular velocity of the gyroscope by using Complementary filter. The test was done by looking at the response of Complementary Filter output angle to the change of predetermined position, including pitching and rolling position" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003345_amm.813-814.964-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003345_amm.813-814.964-Figure3-1.png", + "caption": "Figure 3. Idea Developed for Brake Pedal.", + "texts": [ + " The design of automatic braking control systems is clearly dependent on the braking system characteristics - Material property. Earlier Brake Pedal Figure 1. Originals Brake Pedal. The pedal shown in figure is used by the driver of a vehicle to operate the brakes. This pedal is hinged about base plate by a hinged pin our foot apply force on this pedal. As shown figure 2, the 3D cad model of brake pedal is drawn in creo2.0 software as per dimension specification of old brake pedal.As existing pedal has more mass,best cad models are developed with consideration of different aspect as shown in figure 3. The material used for the brake pedal is is cold rolled steel sheet(FePo3 En10130 series) series which has following chemical and mechanical properties as shown in Table 1. As shown in Figure 4, with the help of ANSYS software, the geometric model was divided into tetrahedral finite elements of higher order (solid187). Each element is defined by 10 nodes \u2013 at corners and midside of edges of the tetrahedron. Each node of this element has three degrees of freedom allowing translational movement in x, y and z directions" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000429_tmag.2013.2276092-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000429_tmag.2013.2276092-Figure1-1.png", + "caption": "Fig. 1. Initial design of the motor topology. (a) Three-slot stator. (b) Two-pole rotor.", + "texts": [ + " The initial design of the motor consists of a stator having three slots which carry nonoverlapping windings, and a Manuscript received April 22, 2013; revised July 25, 2013; accepted July 26, 2013. Date of current version December 23, 2013. Corresponding author: C.-C. Hwang (e-mail: cchwang@fcu.edu.tw). Color versions of one or more of the figures in this paper are available online at http://ieeexplore.ieee.org. Digital Object Identifier 10.1109/TMAG.2013.2276092 2-pole diametrically magnetized sintered NdFeB magnet rotor with a carbon fiber sleeve as shown in Fig. 1 and Table I. The 2-pole rotor consists of a unique cylindrical piece of magnet and a shaft. The magnet has a remanence of 1.2 T and a relative recoil permeability of 1.05. Since the ratio of the slot opening to slot pitch is low, the losses induced from the stator slotting permeance harmonics are generally quite small [17]. Therefore, rather than using laminations, the rotor yoke uses a magnet solid steel, such as alloy ID 35 CD 4 [18]. The losses of the motor include iron losses in the stator, copper losses in the winding, eddy current losses in the rotor, and rotational losses due to friction and windage" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001764_tia.2010.2057398-Figure15-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001764_tia.2010.2057398-Figure15-1.png", + "caption": "Fig. 15. Thermocouple detector location in the cage.", + "texts": [ + " 14 presents a test to calculate the comparison where a good correlation between these values can be noticed. The small differences can be caused by parameter variations: material properties, losses, assembly variations, etc. The minor changes in the machine geometry can lead to significant ventilation changes. It is not uncommon to test \u201csister\u201d machines and notice variations within the range of the errors presented. Under the stall condition, the temperatures were measured using thermocouples distributed on the bars, the core, and the end ring. Their locations are indicated in Fig. 15. Thermocouple 1 was placed at the top of the bar and thermo- couple 3 at the bottom, both at the middle of the bar length. In the same way, thermocouples 2 and 4 were positioned on a tooth at the top and the bottom, respectively. As the model considers the heat transferred from the rotor bars to the rotor lamination, the temperature rise of the tooth is critical for the model validation process. Thermocouples 5 and 6 were placed on one of the end rings. The radial end ring expansion was measured to check the calculated value" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001905_2010-01-0530-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001905_2010-01-0530-Figure1-1.png", + "caption": "Figure 1. 6 DOF vehicle rollover model.", + "texts": [ + " The roll stability was also defined by its static rollover threshold, SRT [8,9,10,11,12]. El-Gindy [8] suggested a slightly different definition in which the SRT is \u201cThe maximum lateral acceleration beyond which static rollover of the vehicle occurs\u201d. This new definition was proposed to distinguish between SRT measurements made using tilt table and those made using full-scale, quasi-steady turning tests. Much work was devoted to study roll stability to control vehicle rollover [12,13,14]. The mathematical modeling of vehicle rollover is shown in Figure 1, where the following assumptions are considered: 1. The model is symmetric about the vertical axis (z). 2. The unsprungmass mu type is solid axle. 3. Both the sprungmass ms and the unsprungmass mu are free to move in roll \u03a6, bounce z, and lateral y directions only. The theoretical analysis is made to investigate the body roll, bounce and lateral displacement due to impulse force. Applying Lagrange's approach, the equations of motion in matrix form are obtained, (see Appendix A). Solving the equations of motion using Matlab software, the frequency dependent bounce, roll, and lateral displacements of the sprungmass and the unsprungmass are obtained (Figures 2, 3, and 4)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002744_1.5122082-Figure10-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002744_1.5122082-Figure10-1.png", + "caption": "FIGURE 10. Meridional shape and shape of the scapula apparatus 2D IMP-0015-0683-0331. The project by the method of universal modeling", + "texts": [ + " The following method for selecting input sizes for low-flow impellers is proposed: for given parameters and constructive constraints according to the formula (3) and other known relations from [4] 0 1 1 b1 s h D ,D , b ,\u03b2 ,R ,R the sizes are calculated for the values 1DA , FK = 0.90, DK = 1.03; the input angle of the blades is assumed to be 300; the value of the height of the vanes at the inlet is chosen, at which the ratio of the flow coefficients is /sf des 1.3, determined by calculation according to the 8th version of the Universal Modeling Method. In accordance with these recommendations, the impeller 2D IMP-0015-0683-0331 was designed. Fig. 10 shows the meridional shape and shape of the 2D IMP-0015-0683-0331 blading apparatus, and in Fig. 11 - velocity diagrams in the design mode. 030032-7 The speed diagrams of the impeller 2d IMP-0015-0683-0331 for the corrected recommendations of the primary design at the design mode des = 0.015 on the steel character are similar to the 2D XXX3-Q velocity diagrams (Fig. 6 above), but differ in the lower input speeds and lower velocity peaks at input edges. According to the calculation of the velocity diagrams in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003283_870182-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003283_870182-Figure3-1.png", + "caption": "Figure 3", + "texts": [ + " Th Is paper compares these two methods through a structura I analysis (FEA), laboratory tests, and manufacturing cons iderations. SpecifIcally, the \"HSK\" method mlnJmlzes the structural dependence of the joint on the header materlaj's_ wall thickness through an appropriate tank to header jolnt design that reduces materlal stresses and detlectJon. HSK Plastic Tank Attachment Method It the lnherent detlciencles of the tabbed header jo lot des 1gn cou I d be m1 t I gated then 1ncreased durabi 1ltv could be expected. Modlne's development of the paten ted \"HSK\" pi dst 1c ta ok attachment method focused on thls objective (Figure 3) (2, 3)*. 2. ANALYSIS AND TEST METHOD 2. 1 PLASTIC TANK ATTJlCfMEKT SYSTEMS When assembl ing a plastic tank to the radiator core, several components are combined at the tank to header joint to develop a water tight seal. Figure 4 illustrates how these components interact to provide the des ired seal In the \"tabbed header\" and 11 HSKlI plastic tank attachment methods. Note that both methods uti I Ize the same plastic tank and gasket des igns. The thl rd component, a metal header, primarily serves to hold the tank and gasket together" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002748_j.ymssp.2015.11.019-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002748_j.ymssp.2015.11.019-Figure1-1.png", + "caption": "Fig. 1. Coordinate systems.", + "texts": [ + " A wide and detailed description of vehicle characteristics, such as drag (Cx) and downforce (Cz) aerodynamic coefficients, tires rolling resistance parameters and ackermann steering coefficients, gives the possibility to model dynamic effects commonly neglected because of their intrinsic complexity, but essential in an activity focused on the characterization of tire interactions. To describe the vehicle motions two coordinate systems have been introduced: one earth-fixed (X0; Y0), the other (x; y) integral to the vehicle as shown in Fig. 1. With reference to the same figure, v is absolute velocity of the center of gravity (CG) referred to the earth-fixed axis system and U (longitudinal velocity) and V (lateral velocity) are its components in the vehicle axis system; r is the yaw rate evaluated in the earth fixed system, \u03b2 is the vehicle sideslip angle, Fxi and Fyi are the longitudinal and the lateral components of the tire\u2013road interaction forces respectively. The front and rear wheel tracks are indicated with tF and tR, while the distances from front and rear axle to the center of gravity are represented by a and b, respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003395_1.1778720-Figure11-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003395_1.1778720-Figure11-1.png", + "caption": "Fig. 11 Best-fit displacement field calculated using MoorePenrose analysis for target field wd\u00c4A0z2ez", + "texts": [ + ", while the displacements are clearly not periodic. The first target displacement field is described by wd5A0z2ez . (19) Here, the displacement corresponds to the vertical displacement of the nodes of the solid face sheet corresponding to a state of constant curvature of kzz52A0 with A0 as the amplitude factor and ez the unit vector perpendicular to the plate ~aligned with the z-axis!. The Moore-Penrose best-fit actuations for this displacement field are calculated as described above. The achievable displacement field is shown in Fig. 11\u2014only the achievable field is shown, as it is indistinguishable from the target field. There are 384 members that are actuated in this simulation and only 209 target nodal displacements. However, the rank of B is only 194, so it is interesting that the achievable field is so close to the target field. Actuation energy and strains will be discussed in Section 5. The second target displacement field is 658 \u00d5 Vol. 71, SEPTEMBER 2004 rom: http://appliedmechanics.asmedigitalcollection.asme.org/ on 01/28/20 wd5A0ez sinS 2pz Lz D (20) where Lz is the length of the unit cell in the z-direction" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000661_aero.2005.1559404-Figure12-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000661_aero.2005.1559404-Figure12-1.png", + "caption": "Figure 12. Exploded view of H-pol feed.", + "texts": [ + " Each 4x2 subarray consists of four 1x2 longitudinal shunt slot subarrays. The eight 4x2 subarrays are fed from another layer consisting of eight rows of series-shunt angled coupling slots \u2013 four slots per row. An eight-way power divider on a third layer provides the correct amplitude and phase taper to each of the eight angled coupling slot rows. The final H-pol configuration consists of four machined aluminum layers, with couplers transferring the power between the layers. An exploded view of the H-pol feed showing the different layers is shown in Figure 12. Silver paint was initially used to provide electrical contact between the layers; however, this proved to be much more difficult in the H-pol feed due to the short drying time of the paint. Initial measurements showed an asymmetry in the pattern, which was inconsistent with the symmetry of the feed and was subsequently traced back to poor electrical contact between layers in some areas. The parts were re-bonded with conductive silver epoxy and re-tested. The resulting patterns are shown in Figure 13" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000083_ir-03-2018-0042-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000083_ir-03-2018-0042-Figure5-1.png", + "caption": "Figure 5 Knee joint description", + "texts": [ + " This shaft is attached to the motor to transmit the angular movement to the shank link. The motor is fastened on the other side by using four bolts and spacer, whereas a locker pin is used on the other side to prevent axial motion of the shaft. The wearer is attached to the wearable vehicle through a fixable tie. Thigh and shank links can be adjusted by a telescopic bar that can be fastened at different positions by two screws in both sides to accommodate all the wearer\u2019s thigh and shank links. The detailed design is shown in Figure 5. The wearable vehicle ankle has three DOFs; the flexion/ extension axis coincides with the human ankle flexion/ extension axis, and the ankle joint connects the shank link with the feet. For design simplification, the abduction/adduction of the wearable vehicle ankle does not pass through the human foot\u2019s rotation. It also has ankle rotation. Flexion/extension movement is actuated by motor which is fastened on the bearing mount by four bolts and spacer. Two ball bearings are mounted in the bearing mount in the two sides, which is attached to themotor by themotor shaft" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002761_s11071-016-2794-1-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002761_s11071-016-2794-1-Figure1-1.png", + "caption": "Fig. 1 Tethered satellite system of two degree of freedom", + "texts": [ + " This paper studies the NNMs and their stabilities for a TSS, in which the mother satellite is treated as a rigid body. The study begins with the modeling of the inplane motion of the TSS during station-keeping phase in Sect. 2. Themethod ofmultiple scales inmatrix form is employed to compute the approximate solutions for the NNMs, and their stabilities are analyzed in Sect. 3. The case studies are given in Sect. 4. 2 Modeling of a tethered satellite system Consider an in-plane tethered satellite system moved in an unperturbed Kepler circular orbit, as shown in Fig. 1. The mother satellite is treated as a rigid body of mass M , and the subsatellite is envisioned to be a point of mass m that is attached to the mother satellite through an inelastic massless tether of length l at a joint point of distance, \u03c1, to the mass center of the mother satellite. It is assumed that themass of themother satellite is much greater than that of the subsatellite, and the center of mass of the system coincides with that of the mother satellite, where the mother satellite moves in an unperturbed Kepler circle orbit of radius R and true anomaly \u03bd. The Earth-centered inertial frame is denoted byO-XYZ, the origin of which is located at the center of the Earth. The origin of orbital frame C-xyz is put at the mass center of the system, with the x-axis pointing toward the center of the Earth, the y-axis following the tangent of orbit and the z-axis completing the right-handed coordinate system. The body frame C-xb ybzb is established along with the principal axes of the spacecraft. According toFig. 1, it is easy towrite out the position vector of the mother satellite and the subsatellite in inertial frame as r1 = [ R cos v R sin v ] , (1) and r2 = [ (R \u2212 l cos \u03b8 \u2212 \u03c1 cos\u03b1) cos v + (\u03c1 sin \u03b1 + l sin \u03b8) sin v (R \u2212 l cos \u03b8 \u2212 \u03c1 cos\u03b1) sin v \u2212 (\u03c1 sin \u03b1 + l sin \u03b8) cos v ] . (2) The potential energy of the system is expressed in the following form [31] V = \u2212\u03bceM |r1| \u2212 \u03bcem |r2| \u2212 1 2 \u03bce |r1|3 (Ix + Iy + Iz) + 3 2 \u03bce |r1|3 ( Ix cos 2 \u03b1 + Iy sin 2 \u03b1 + Iz ) , (3) where \u03bce = 3.986 \u00d7 1014 m3/s2 is the gravitational parameter of Earth, Ix , Iy and Iz are the principal moments of inertia of the mother satellite expressed in the body frame" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003865_tia.2019.2923717-Figure15-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003865_tia.2019.2923717-Figure15-1.png", + "caption": "Fig. 15. Manufactured stator and rotor.", + "texts": [ + " The effective value of the voltage is 133.2V, which shows good agreement with the design target of 220/\u221a3 (=127V). It also means that even if the armature current is applied, the estimated values of the reactance and the back EMF are still valid with little saturation effects. In table III, the simulated values of back EMF, synchronous reactance, the phase voltage and average torque are compared with the predicted values in design procedure, showing very good agreements. actually manufactured. Fig. 15 shows the stator and the rotor assembled respectively, and Fig. 16 shows the set-up for experiment. First, the back EMF was measured at the base speed of 214 rpm and the phase voltage waveforms are represented in Fig. 17. The rms value of the back EMF measured is about 71 V, which is very similar to the design value of 66.3 V as well as the finite element analysis results 0093-9994 (c) 2019 IEEE. Personal use is permitted, but republication/redistribution requires IEEE permission. See http://www" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001251_cdc.1995.478947-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001251_cdc.1995.478947-Figure4-1.png", + "caption": "Figure 4: Cart and Pendulum Hovering Motions for a = 0, t,\ufffd. The dotted lines indicate the amplitude of the response to the forcing. These are exagger ated for the purpose of illustration. In the case of the stable responses, the dynamic effects of forcing dominate gravity.", + "texts": [ + "pha= 0 For Qo = a - %, a + % ' (15) is positive implying V is 3\ufffd r==========l (b) alpha = pil4 4,-\ufffd\ufffd\ufffd\ufffd====\ufffd 3.;\ufffd-concavc up at Qo and hence (13) is stable in the sense of 2.5 unsIlIIJIe 2.5 unstable stable Lyapunov. For Qo = {a,a + 1r}, (15) is negative, hence 1.; V is concave down at Qo and (13) is unstable. 0 1 \ufffd stable 2 Remark 3 The stable equilibria of the averaged system give rise to the hovering motions which appear as ( be comes small (w becomes large), and which are the aim of our study. Qualitatively, these appear as in Figure 4, where the stable motions of the system have the pendulum executing small amplitude oscillations in a neighborhood of the averaged system equilibrium. Laboratory exper iments indicate that this hovering phenomenon is quite stable. Note that the unforced (13 = 0) pendulum has two equi libria - a hanging equilibrium (Q, P) = ( 0,0), and an in verted equilibrium (Q, P) = (1r,0). The other two equi libria arise as the result of bifurcations of (13). We view ('; as the bifurcation parameter, and a plays a role in de termining the type of bifurcation" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000881_pesc.2008.4592618-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000881_pesc.2008.4592618-Figure8-1.png", + "caption": "Figure 8.Lateral N-MOS hole current density under drain and gate at 15V polarization (eox=100nm)", + "texts": [ + " It can be seen that N-MOS operating under reduced nominal voltage levels will have a better chance to not enter in avalanche mode. A deeper analysis within the power device has been carried out thanks to numerical simulation results. As expected, the avalanche takes place around the drain\u2013 channel region N+P junction. It is due to the contributions of a lateral electric field coming from drain\u2013source polarization and a vertical electric field coming from gatechannel region polarization as it can be seen figure 8. The analysis of this situation shows that a reduction of the operating voltage level of both vertical and lateral devices brings to a satisfactory tradeoff. The oxide thickness is reduced as well as the nominal gate voltage level. In such way, the vertical power device channel is inverted in similar conditions. On the other way, the lateral device is able to withstand a reasonable voltage level. Figure 9 shows a cross section of a lateral device operated under 7V and having an oxide thickness of 30nm" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000947_icmmt.2008.4540409-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000947_icmmt.2008.4540409-Figure4-1.png", + "caption": "Figure 4 Mechanical drawing of the antenna array", + "texts": [], + "surrounding_texts": [ + "As shown in Figure 5, the array consists of four L shape elements instead of monopole antennas to decrease the overall height. The distance between adjacent elements is 13mm, which is around a quarter of wavelength at 5.9GHz. Each element has a 6mm vertical segment and a 6mm horizontal segment. The ground plane is assumed to be infinite in all simulations. The Figure 5 shows the drawing of the feeding network. The four antennas are numbered from #1 to #4. Two isolated feeding ports are numbered as #5 and #6. As shown in Table 1, when the port 1 is excited alone, the signal power at all elements is same and the phase descends 90\u00b0 between adjacent elements from #1 to #4. Considering the 90\u00b0 space separation between adjacent elements, the phase distribution excited by port 1 generates an endfire radiation pattern towards the element #4. When the port 2 is excited alone, the phase relationship between elements flips, thus generates an endfire radiation pattern towards the element #1. Table 1 Normalized power and phase of antenna elements when the array is excited from different ports Ant. 1 Ant. 2 Ant. 3 Ant. 4 Port 1 o13525.0 \u2220 o4525.0 \u2220 o4525.0 \u2212\u2220 o13525.0 \u2212\u2220 Port 2 o13525.0 \u2212\u2220 o4525.0 \u2212\u2220 o4525.0 \u2220 o13525.0 \u2220 III. SIMULATION RESULTS Figure 6 shows the simulated transmission loss of the feeding network when feeding from the port 1. The transmission loss of port 2 is similar to port 1. In the working band of WAVE (5.850-5.925GHz), the transmission loss to all four antenna elements is well balanced and close to -6dB. Figure 7 shows the simulated transmission phase of the feeding network when feeding from the port 1. In the working band of WAVE, the transmission phase difference between adjacent elements is close to 90\u00b0. Figure 8 shows the simulated return loss of both port 1 and port 2. In the working band of WAVE, the return loss is well below -20 dB. Figure 9 shows the isolation between port 1 and port 2. In the working band of WAVE, the isolation is below -10dB. In a WAVE application the two radiation patterns do not need to work at the same time, thus the isolation is not a critical specification. It only has a small impact on the gain of array. Figure 9 Simulated isolation between two feeding ports of the feeding network, terminated by antenna impedances Figure 10 and Figure 11 show the radiation patterns of both port 1 and port 2 in the elevation plane and the azimuth plane respectively. The simulated gain of the array is 11.5dBi. The 3dB beamwidth in the azimuth plane is 82\u00b0. Due the assumption of infinite ground, in the elevation plane the radiation pattern only exists on the top hemisphere. The 3dB beamwidth in the elevation plane is 40\u00b0. -10 0 10 0 30 60 90 120 150 180 210 240 270 300 330 -10 0 10 G ai n (d B i) Port 1 Port 2 Figure 10 Simulated radiation patterns of port 1 and port 2 in the elevation plane" + ] + }, + { + "image_filename": "designv6_24_0002733_6.2012-5330-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002733_6.2012-5330-Figure6-1.png", + "caption": "Figure 6. Conceptual image of Cosmic Mariner rendezvous with nuclear fuel depot at EML-1.", + "texts": [ + " By keeping most of the on-board electronics in an unpowered state during the transfer, the impact of single-event effects (SEE) caused by interaction with energetic particles is reduced. Also, keeping Cosmic Mariner unmanned during this D ow nl oa de d by P U R D U E U N IV E R SI T Y o n Se pt em be r 16 , 2 01 6 | h ttp :// ar c. ai aa .o rg | D O I: 1 0. 25 14 /6 .2 01 2- 53 30 American Institute of Aeronautics and Astronautics 5 transfer reduces the health hazard to the eventual inhabitants. Once Cosmic Mariner reaches EML-1, it will rendezvous and dock with a nuclear fuel depot in order to activate its nuclear reactor (Fig. 6). It will then return to a normal power state prior to the arrival of its first inhabitants via a commercial crew transport vehicle (Fig. 7). An initial crew of up to 12 can be comfortably accommodated. IV. Transfer to Lunar Orbit and Initial Operations After the arrival of the first inhabitants at Cosmic Mariner, the vehicle will transfer to lunar polar orbit (LPO) to commence initial operations. A low-thrust trajectory from EML-1 to LPO will require a \u0394V of approximately 1.5 km/s. With the nuclear power available to the electric engine, this transfer can be completed in a relatively short timeframe" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002600_icems.2011.6073669-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002600_icems.2011.6073669-Figure2-1.png", + "caption": "Fig. 2. Structure of AFFSPM generator.", + "texts": [ + " The structure is economically feasible for the AFFSPM wind power generator. This is because, first, the AFFSPM generator features the high power density, high torque density and the good heat dissipation, and the additional copper loss fever caused by Id could be overcame. Moreover, the four quadrant operation and the power bi-directional flow are not necessary in direct drive wind power grid-connected system. Combined with the factor of reducing cost, the configuration of non-controllable AC/DC and controllable DC/AC is chosen to realize this new strategy. Fig.2 shows the structure of the AFFSPM generator, the 12 coils on each stator is divided into three phases, 4 coils are in series in each phase. Assuming that the AFFSPMG permanent magnetic flux is sine, and the magnetic path saturation and higher harmonic of the back-EMF are ignored. The equations of voltage and the electromagnetic torque are given by \u23aa \u23a9 \u23aa \u23a8 \u23a7 \u2212+= +++\u22c5= +\u2212\u22c5= ])([5.1 / / sqsdsqsdsqme mesqsdsdsqsqsq sdsqsqsdsdsd iiLLipT RiiLdtdiLU RiiLdtdiLU \u03c8 \u03c8\u03c9\u03c9 \u03c9 (1) Where Lsd, Lsq are the direct-axis and the quadrature-axis inductance respectively, e\u03c9 is the electricity angular velocity, m\u03c8 is the amplitude of the stator flux-linkage, and P is the pole-pair number" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001599_j.eaef.2016.01.002-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001599_j.eaef.2016.01.002-Figure1-1.png", + "caption": "Fig. 1. Isometric view of the moringa seed shelling machine.", + "texts": [], + "surrounding_texts": [ + "This device controls the quantity of seeds that is being introduced into the shelling chamber. It was designed in such away as to allow for variation in the amount of seeds that is being introduced into shelling chamber." + ] + }, + { + "image_filename": "designv6_24_0003027_imece2014-38788-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003027_imece2014-38788-Figure6-1.png", + "caption": "Figure 6: Solid model of the compliant mechanism designed in example 1a", + "texts": [], + "surrounding_texts": [ + "Specifying appropriate energy/torque for a mechanism at various precision positions is cumbersome. For simplicity, a heuristic judgment may be made between the energy/torque specifications and the rotation of the pseudo-rigid-body links of the compliant mechanism, to ensure the specifications are appropriate. On various occasions, however, a designer may still need assistance with providing appropriate specifications. The above presented optimization formulation guides the designer with this. The function value at the end of the optimization process is an excellent indicator of the energy or force/torque specifications. If the function value is not close to zero, then some iteration must be conducted by changing the initial estimates drastically. This will ensure a search for the global minimum. If the function value at these various starting positions is still not close to zero, then an unrealistic problem definition may exist. In this instance, the following steps should be conducted to better understand the change in direction: 1. Determine whether or not the energy/torque at various positions is in agreement with the rotation of the pseudo-rigid-body links. 2. If the result from Step 1 is deemed satisfactory, then the user should either increase or decrease the energy/torque specifications. 3. Examine the function value at the end of Step 2. If the function value is approaching zero, then continue in the same direction until the desired function value is achieved. In case the function value is diverging further, change the direction and repeat Step 3. The above process is illustrated in examples presented in the following section." + ] + }, + { + "image_filename": "designv6_24_0001288_icnn.1994.374789-FigureI-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001288_icnn.1994.374789-FigureI-1.png", + "caption": "Fig. I Kinematics Principle and Structural Layout of a Joint Module", + "texts": [], + "surrounding_texts": [ + "Positioning/tracking performance of a modular robot system relies very on the kinematics accuracy of its joint modules. In order to satisfy diverse task requirements, various types of joint modules of one, two and even three degrees of fieedom, different dimensions and performance specifications have been designed[l3]. In this study a type of modular joint with two rotary DOFs is dealt with, whose kinematics principle and structural layout are i l l d in Fig. 1 [l]. There are two stages of movement transmissions in this modular joint: two harmonic reducers (Reducer Module) amplifying torque and a differential mechanism which consists of three bevel gears changing two onedimensional movements aM1 and aMz into two two-dimensional movements ( P D and (Ps (ioint angles of a robot). The basic transmission relations can be described by where (PO and (Ps represent rotary and swing angles of the module; aM, and CtMz represent rotator angles of the motors; and iG and iK are ratios of the harmonic reducer and of a pair of bevel gears, respectively. Kinematic errors in the joint angles caused by mechanical inaccuracies have been modelled as 272 1 where AqD and Aqs represent the errors in joint angles; Aqi, and Aqi2 the emrs of harmonic reducers; and the barckla~hes betwemi two pairs of engaging bevel gears; e,,,e,, and e,, the effective cccenthities of axial, radial and angular fluctuations in the gear-baing system of bevel gear i; Ufl,U,, and U,, the Sensitivity Coefficients comsponding to e,,,e,, and e,, which are functions of the joint angles and directions of eccentricities; and rlo and r30 the pitch radii of bevel gear 1 and 3, respectively[4]. Additionally, C, and C, (=l or +1) represent the contact coefficients of backlash. As the mechanism always trausfers moment, gears tend to keep in contact. Thus, C,(C,)=+l indicates that two gears are in contact in the direction of movement and C1(C2)==1 does that two gears are in contact in reverse direction of movement." + ] + }, + { + "image_filename": "designv6_24_0000644_apmc.2007.4554952-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000644_apmc.2007.4554952-Figure2-1.png", + "caption": "Fig. 2. (a) Single-slot endwall transition. (b) Double-slot endwall transition.", + "texts": [ + " An antenna array example is introduced in Section IV. Finally, some concluding results are presented in Section V. II. PROPOSED TRANSITION Because of the large number of parameters and the fairly sensitive nature of the coupling mechanism for this coupler, the design procedure requires an accurate analysis tool. Thus, we used an accurate numerical package which employs the Finite Integral Technique (FIT). The power splitter is designed on a 31 mil RT5880 substrate in order to feed a two-element array. Thus the structure of the Fig. 2(a) has been employed. This structure includes a WR-55 waveguide ended in the microstrip substrate. The attempt to design a power splitter using a single slot in the ground failed due to weak coupling and poor return loss. The maximum coupling obtained was about 4.5 dB, but it turned worse when two stubs were used near the slot. Two matching stubs are employed to improve the coupling and the matching ofthe waveguide to microstrip line. Therefore, another strategy to be utilized is to use double slots and double matching stubs on the microstrip circuit. To implement this idea, two coupling slots have been etched on the ground plane of the microstrip circuit and consequently on the top of the waveguide. In the optimization process, the larger separation between the two slots, the better matching was achieved. To match the waveguide to the microstrip line, both stubs were placed on the top of the slots. The structure in Fig 2(b) shows the final configuration in which either slots is placed close to the side walls of the waveguide. The magnitude of current on the ground plane and the microstrip lines are depicted in Fig. 3 and Fig. 4, respectively. It is obvious that the current and field vectors on the slots are in opposite directions. III. SIMULATION RESULTS After optimization of the transition, the simulations show over more than 22% of relative bandwidth from 16 to 22.2 GHz. Fig. 5 shows the return loss and Fig. 6 presents the coupling between two ports of the power splitter" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000661_aero.2005.1559404-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000661_aero.2005.1559404-Figure2-1.png", + "caption": "Figure 2. Illustration of symmetric-fed reflectarray", + "texts": [ + " The initial concept was a symmetric-fed antenna design, which is briefly described in the following section. Although this design was proven in breadboard testing, the placement of electronics and cabling could not meet system requirements. To overcome this problem, a novel offset-fed reflectarray design was developed. Although this design significantly increased the technical risk, the design software developed for the symmetric design enabled rapid development of the offset design. The initial WSOA interferometer concept uses a symmetricfed reflectarray antenna design [4] as illustrated in Figure 2. Each reflectarray is comprised of five panels and a two feeds, one feed for vertical polarization and one for horizontal polarization. A mechanical deployment mechanism is used to fold the flat panels into a stowage configuration for launch. The panels use variable-sized square patch reflectarray elements printed on Rogers RO4003 substrate material. A piecewise parabolic reflectarray design is used to minimize the total phase shift that must be supplied by the reflectarray elements in order to collimate the beam" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000202_icra.2012.6225306-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000202_icra.2012.6225306-Figure2-1.png", + "caption": "Fig. 2. Shows a ghosted view of the force-amplifying device. (1) is the tool/probe with distal-mounted force sensor, (2) shows the slider feature that contacts the user\u2019s fingertip, (3) are the bearing rollers seated within the rail (4). Load is transmitted from the slider feature and on to the tool/probe clamp (5) through the force sensor that measures Ff (6). The linear motor (7) is coupled to the tool/probe clamp, bushings (8) along the through-bore support the tool/probe and (9) is the electrical cable outlet point.", + "texts": [ + " The linear motor can provide high forces (1.03N continuous, 2.72N peak) relative to its size (8\u00d78\u00d730mm) and weight (15g), making it suitable for a compact hand-held device. Another reason for the selection of the linear motor is that it has just one single moving component: a magnetic shaft which translates within the bore of the motor housing, making the motor completely free of backlash and with minimal friction. The linear motor has integrated Hall-effect sensors to provide position feedback and permit motion control. Fig. 2 illustrates the engineering features of the force-amplifying device. The vectors of the forces acting on the sliding assembly; Ft, Fm and Ff are not collinear (this is best illustrated in Fig. 1) which leads to a moment being generated as the device is loaded during operation. In order to avoid jamming of the sliding assembly due to this moment, the slider assembly translates on bearing rollers that are seated within a rail so as to minimize friction. The rail is mechanically clamped to the casing shell that is held by the user\u2019s grip" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002320_mwsym.1992.188038-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002320_mwsym.1992.188038-Figure1-1.png", + "caption": "Figure 1.0 Four MESFET combiner", + "texts": [ + " In this structure, varactorless frequency tuning via bias adjustment is studied. Results of a periodic MESFET based combiner are presented here. The structure contains four MESFET oscillator cells which corporately feed a uniform linear array of microstrip patch antennas. Frequency tuning as well as power combining is investigated. These structures have promising application in doppler motion sensors, noninvasive medical imaging, and other array radiating applications. 545 CH3 141-9/92/0000-0545$01 .OO 0 1992 IEEE Figure 1. shows the circuit structure. A transmission line is loaded with oscillator \u201ccells\u201d which are spaced hg, the guide wavelength of the oscillation frequency, apart. This spacing allows for the geometry of the antenna array as well as the proper phasing of the feeds of the array. The individual \u201ccells\u201d are designed using a small signal iterative procedure, utilizing a commercially available microwave CAD package, to optimize the reflection coefficient at the output of each device. A short circuited series inductive stub is used as a feedback element with the E T to ensure instability" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000459_jae-2010-1266-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000459_jae-2010-1266-Figure4-1.png", + "caption": "Fig. 4. Transmission torque experiment device.", + "texts": [ + "58 mm is smaller than that of adjacence magnets on the same gear, the distance between the gears is considered as sufficient gap in order to set a magnetic circuit. In the experiment, we study the relationship between the distance of magnetic gears and the transmission torque. The motor current and the output torque are measured. The input torque can be obtained from product of the motor current and the motor\u2018s torque coefficient. The output torque is measured by a torque detector that is attached with the gear 2. The outline of measurement experiment device is shown in a Fig. 4. First, the magnetic gear 1 is placed at the maximum distance and the output shaft, that is connected directly to the torque detect device, is fixed. The input torque is enlarging until both the magnetic gears slip. When the magnetic gear is slipped, the input current and the output torque are recorded. This step is repeated 10 times and the average is calculated. Secondly, the magnetic gears come close about 0.25 mm by distance adjustment screw of slider. The input current and the output torque are recorded when the magnetic gear was slipped like above step, and the average is obtained by 10 time\u2019s trial" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002453_978-3-030-20131-9_174-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002453_978-3-030-20131-9_174-Figure1-1.png", + "caption": "Fig. 1. Representation of the variable stiffness mechanism mounted on a robot arm (driving a differential mechanism); A: driving joint (actuator), B: driven joint, C: freely rotating idler, D: beam spring, E: belt, F: lead screw.", + "texts": [ + " In this paper, a simple variable-stiffness transmission is presented which draws on some of the principles from prior literature. Stiffness is adjusted mechanically, using a compliant beam as the stiffness element. Adjustment of the mechanism simultaneously changes the load on the spring element, as well as the orientation of the load with respect to the spring (effectively changing the transmission ratio). The result is a wide range of achievable stiffness properties. The variable stiffness mechanism (see Figure 1) consists of a primary actuator, a driven rotational joint, a belt coupling, a spring-loaded idler (freely rotating pulley mounted on an elastic beam), and a lead screw with secondary actuator. The lead screw is used to adjust the position of the base of the beam with respect to the primary actuator and driven joint. The belt couples the primary actuator to the driven joint under a no-slip condition, but is also acted upon by the idler mounted on the distal extremity of the beam. Actuation of the lead screw results in variation of the stiffness response of the driven joint with respect to the actuator input" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002779_s11044-016-9540-9-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002779_s11044-016-9540-9-Figure1-1.png", + "caption": "Fig. 1 Flexible multibody system", + "texts": [ + " 3, the adjoint equations are derived for the given dynamic system. The procedure is tested by means of a flexible slider\u2013crank mechanism in Sect. 4. Finally, Sect. 5 concludes with a brief summary and discussion. The method of flexible multibody systems is a well-established approach to model and analyze mechanisms in which the single bodies undergo large rigid body motions and deformations. Next to rigid and flexible bodies, these systems are assembled from spring and damper elements, actuators, and ideal joints; see Fig. 1. If the deformations are comparatively small, then the floating frame of reference formulation can be used to efficiently incorporate flexible bodies into the multibody system; see [18, 19]. In the following, the basic equations of the floating frame of reference formulation are briefly reviewed, and the dynamic problem is formulated in minimal coordinates using a coordinate partitioning. Thereby, we assume that the flexible bodies are parameterized by the independent design variables x \u2208 R h, whose influence on a scalar objective function \u03c8 \u2208R will be identified in the course of this work" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000403_ijhfms.2006.011683-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000403_ijhfms.2006.011683-Figure3-1.png", + "caption": "Figure 3 Joint-link system with external loads", + "texts": [ + " c ncW W W (10) The conservative work can be expressed as the negative of the potential energy of the force system, i.e. cW V . Thus we can write ( ) nc ncW T T V W L W (11) where V is the total potential energy of the system, Wnc is the virtual work done by nonconservative forces and torques and L T V is called the Lagrangian function. Then equation (9) can be written as follows: nc ncW Wd L L d dt dt 0 q q q q (12) We will derive a more explicit form of the equation for a general serial manipulator and force system, as illustrated in Figure 3. The Cartesian coordinate of the point fixed in the ith local frame in terms of the global coordinate system was given earlier as 0 0 1( ,..., ) i i i i iq qr T r (13) The velocity of a point with differential mass dm in link i can be derived as follows: 0 0 0 1 ( ) ( ) ( ) i i ii i i i i j i jj d d q dt dt q T qv r T r r (14) The kinetic energy Ti of link i is calculated from the kinetic energy of a differential mass as 1 ( ) 2 T i i i m T Tr dmv v (15) where Tr is the trace of a matrix, i.e. the sum of the diagonal elements of a matrix", + " The consequent final equations of motion are compiled as follows in the vector-matrix form as coupled, non-linear, second-order ordinary differential equations: kT T N i i k ki k m F = M(q)q + V(q,q) + J g + J K q - q M (28) where Jk is the Jacobian matrix (for any point in any local coordinate) of the point at k kr (4 1) location vector with respect to the kth local coordinate frame. 0 0 0 1 0 1 1 6 ( ) ( ) ( ) ... ... ( ) ( ) ... ( ) ... ( ) k k kk k k k k k k i k i k k q q q T q T q T q r r r J q Z q Z q Z q (29) Here, we only take the first three elements of the (4 1) vectors 0( ( ) ) k k i kqT q r . 1iZ represents the 3 1 local z-axis vector of joint i expressed in terms of the global coordinate system. Recall from Figure 3 that rotations always occur about the z-axis, based on the D-H method. Note that in equation (28), all of the variables are functions of time, i.e., t , k k tF F , k k tM M , and tq q , with k = 1,2,\u2026,n. To apply the above equations of motion to the human joint-link system, some physical characteristics of anatomical segments are needed to determine the coefficients in the equations of motion. These physical properties are the mass of each link, centre of mass for each link, moments of inertia for each link and joint stiffness" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003634_powereng.2015.7266300-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003634_powereng.2015.7266300-Figure1-1.png", + "caption": "Fig. 1. Basic dimensioning of outer-runner rotor BLDC motors.", + "texts": [ + " The main limiting parameters (herein indicated as initial design parameters) for the specific application are presented in Table 3. It should not be forgotten that this application is a reverse engineering application where stator and rotor main dimensions were assigned according to the wheel rim and the drum brake housing inner and outer diameter dimensions. According to basic dimensioning procedures [13] of outer rotor BLDC motors, and with reference to the graphical representation given in Fig.1, stator outer diameter can be calculated from; 2 2r c mD D l \u03b4= \u2212 \u2212 (1) where D is the stator outer diameter, Drc is the rotor inner diameter, lm is the magnet thickness and is the air gap length of the motor. bss1 is the upper slot opening parameter and is calculated by; 1 2 sw s s t s s D h b b Q \u03c0 \u2212 = \u2212 (2) where hsw indicates the tooth thickness of the stator core, bts is the tooth width of the core and Qs expresses the stator slot number. And bss2 gives the lower slot opening parameter and can be calculated as; 2 2 ss s s t s s D h b b Q \u03c0 \u2212 = \u2212 (3) The parameter hss given in the last equation denotes the slot length" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000398_iecon.2010.5675174-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000398_iecon.2010.5675174-Figure2-1.png", + "caption": "Fig. 2 \u2013 Cross-section of a regular four-phase 8/6 SRM.", + "texts": [ + " For these reasons, we started with the experimental tests in order to determine the real machine characteristic and afterwards we design the analytical model. The purpose of this paper is to present accurate models based on analytical functions for the magnetic characteristics of regular SRM structures. A regular switched reluctance is one in which the rotor and stator poles are symmetrical about their centre-lines and equally spaced around the rotor and stator respectively. The four\u2013phase 8/6 SRM is an example of this type of machines, as shown in Fig. 2. The SRM construction is relatively simple as compared to other electrical machines. On the other hand, the SRM characteristics depend on numerous factors [9] namely machine structure (number of phases, number of stator and rotor poles, stator and rotor poles arcs), magnetic properties of the laminations, converter topology and control strategy [10]- [13]. Due to the non-linear magnetization of the iron and to the variable air gap, the flux-linkage of phase j, \u03a8j is a nonlinear function of the stator phase current ij and rotor angular position \u03b8 : ),( \u03b8\u03c8 jijj\u03a8 = (1) This complex function can be obtained by finite-element field computation using the machine geometry and properties or experimental measurement on the real machine" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002010_jae-2008-999-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002010_jae-2008-999-Figure3-1.png", + "caption": "Fig. 3. Proposed pattern and size of the piezoelectric ceramic for the 8.5[mm] outer diameter RUSM.", + "texts": [ + " (63) Finally, the 3D-FDM software for the transient analysis of a piezoelectric system is constructed by using the above mentioned FDM formulations. Briefly, we developed the analysis software, named as the SNU-PIEZO-3D for the steady state analysis and the transient analysis of the USM. The tetrahedral element with 4-node is used in this research. 5. The characteristic analysis and design of a 8.5[mm] outer diameter RUSM The first stage for the design of the RUSM is the design of a piezoelectric ceramic using an analytic method. In this research, we proposed a geometric for the piezoelectric ceramic as shown in Fig. 3 for an 8.5[mm] outer diameter RUSM taking the size and effective generation of traveling wave into consideration. In this section, we suggest the analytic method to verify the generation of a traveling wave for the RUSM and verified the generation of a travelling wave for the proposed RUSM. Analytic method, derived in Section 2, must be modified for the RUSM because the coordinate system and the piezoelectric pattern are changed as shown in Figs 2\u20134. Hence, we derived the modified analytic method for the RUSM and verified the generation of a travelling wave of the proposed RUSM, as follow", + " If the amplitude of two electrical sources is the same then the summation of waves at each piezoelectric ceramic is expressed by: 3\u2211 i=0 A sin {\u03c9t\u2212 k(x + ai) + \u03d5i} + A sin {\u03c9t + k(x + ai) + \u03d5i} (64) where, A: the amplitude of the wave k: the wave number k [rad/m]. ai: the space difference between the ith-piezoelectric ceramic and the 0th-piezoelectric ceramic [m] \u03d5i: the time difference between the ith-piezoelectric ceramic and the 0th-piezoelectric ceramic [rad]. As shown in Fig. 4, the space difference between two piezoelectric ceramics a is: a = \u03c0r 2 (65) We designed the piezoelectric ceramic pattern as shown in Fig. 3 for the generation of three waves. Hence the wave length \u03bb should be (2\u03c0r/3). Using this value and from Eq. (15), we derived the wave length \u03bb: \u03bb = 4a (n\u2212m) = 2\u03c0r 3 . (66) From Eqs (65) and (66), (n\u2212m) should be 3. Using this value and from Eq. (15), we derived the wave number k: k = 2\u03c0 a (n\u2212m) 4 = 3 r . (67) From Fig. 4 and by applying Eqs (65)\u2013(67) into Eq. (64), Eq. (64) is summarized into: 4A sin(\u03c9t\u2212 kx) (68) From this result, it was verified that the proposed Fig. 3 model can generate a traveling wave and can be used for the vibrator of an 8.5[mm] outer diameter RUSM. A mechanical characteristic of a piezoelectric system, such as the displacement, is maximized about the input electrical voltage at resonance. Hence, the mechanical displacement can be maximized by matching the input frequency with the resonance frequency of the piezoelectric system. This demonstrates the importance of the impedance analysis for a piezoelectric system. There exist many resonance modes in a piezoelectric system", + " Figure 6 indicates the impedances of the piezoelectric transducer obtained from the 3-D FEM and the experimental data. The piezoelectric material coefficient, which was the VIBRIT 420 and the experimental impedance result were from [37]. From the results, it was verified that the FEM routine used in this research was accurate. The RUSM in this research was operated at the suitable 6th-mode wave. To find the resonance frequency, which generated the desired mode, the impedance and the mode were analyzed by using 3D-FEM at this design stage. To calculate the impedance of a RUSM, which is shown in Fig. 3, the impedance Eq. (77) had to be modified because the paralleled electrodes of the RUSM: 1 Ztot(w) = n\u2211 i=1 1 Zi(w) = Z2Z3 . . . Zn\u22121Zn + Z1Z3 . . . Zn\u22121Zn + . . . + Z1Z2 . . . Zn\u22122Zn\u22121 Z1Z2 . . . Zn\u22121Zn (78) where, Ztot: the total impedance of the RUSM Zi: the impedance which is calculated by Eq. (71) n: the total number of the paralleled input part (in this case n = 2). Hence, the impedance of the RUSM can be expressed by: Ztot(w) = n\u220f k=1 Zk n\u2211 i=1 n\u220f j=1,j =i Zj (79) where, n\u220f j=1,j =i Zj : the accumulated multiplication from j is 1 to n except when j = i" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001691_12.538682-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001691_12.538682-Figure1-1.png", + "caption": "Figure 1: Schematic representation of the Kagome-structure. The face-sheet is shown in blue, the core in green and the Kagome backplane is red. Actuators are placed in-lieu of the Kagome-members. The control-points are used to define the target deformation.", + "texts": [ + "ne of the goals in shape morphing technology is to cause surfaces to displace even when resisted by large pressure loads (or heavy weights). The challenges become especially demanding when minimum weight requirements and power budgets are imposed. This challenge can be addressed by seeking structures that are simultaneously staticallydeterminate, yet stiff. A two-dimensional manifestation consists of a corrugated structure that can bend and hinge with much higher authority than bimorphs and other competing approaches (Lu et al., 2001). Another manifestation is the Kagome structure depicted on figure 1 (Hutchinson et al, 2003; Hyun and Torquato, 2002), having the attribute that it can be actuated into intricate surface shapes, ranging from bending to twisting to undulating. The intent of the study is to provide an experimental assessment of the concept with associated analysis. For this initial demonstration, hinging and twisting will be explored. The design process selects the geometry and the preferred materials. It ascertains the stresses, relative to the failure envelope, and ascertains the actuator authority needed to maximize the load capacity as a function of the designated displacements. Smart Structures and Materials 2004: Smart Structures and Integrated Systems, edited by Alison B. Flatau, Proc. of SPIE Vol. 5390 (SPIE, Bellingham, WA, 2004) 0277-786X/04/$15 \u00b7 doi: 10.1117/12.538682 175 Downloaded From: http://proceedings.spiedigitallibrary.org/ on 06/26/2016 Terms of Use: http://spiedigitallibrary.org/ss/TermsOfUse.aspx The basic structural design is depicted on figure 1. It consists of a solid face-sheet with a Kagome back-plane and a tetrahedral core. The length of the panel is chosen to include six hexagonal units of the Kagome plate, while the width incorporates four units with member length, L=5.1cm. Along the sides, to avert degradation of the buckling resistance, patch trusses are used. For initial demonstration and to facilitate fabrication, the back-plane and core members have the same length and cross section. The configuration is rigidly supported at one end", + " The same calculations are used to evaluate the forces on the actuators as a function of the displacements. At this stage, the dimensions are adjusted and the load capacity re-evaluated to assure that there is no cyclic yielding (indicative of fatigue) or buckling. A detailed analysis is available elsewhere (dos Santos e Lucato et al, 2004) The magnitudes of the passive loads that can be supported without failure are derived for a cantilever plate with a line load, P, (per unit width) imposed at the free-end (figure 1). Emphasis is placed on analytic formula that can be used to select materials and to assess the scaling. A complete analysis, presented elsewhere (Wicks and Hutchinson, 2001) (figure 2), reveals that when optimized, this design is as good as the best available truss structures and competitive with honeycomb core panels. In loadings that cause the solid face to experience compression, it fails by local buckling, requiring that this face be relatively thick. The consequence is a minimum weight design comparable to that for the octet truss (Deshpande and Fleck, 2001)", + "74 df dc 2 Lc s (7) At the optimum, \u03c7c \u2248 1 (Wicks, 2003). The present design (s/Lc =5.3), with back-plane and core members having the same cross section, is sub-optimal (\u03c7c \u2248 0.32): whereupon the loads are limited by yielding of the back-plane. Accordingly, (7) can be used to estimate the passive load capacity as, PM \u2248 1.2kN/m (equivalent to a load of 20kg). Note that, replacing stainless steel with a Ti alloy (\u03c3Y = 800 MPa) should increase the load capacity by a factor four: PM \u2248 4.8kN/m. Because of the bonded nature of the structure (figure 1), forces are induced upon imposing an actuation strain that can cause the system to be actuator-limited, rather than structure-limited. Before embarking on a shape morphing demonstration, it is essential to evaluate these forces and compare them with the operating characteristics of the actuator. Results are presented for actuators placed along the mid-section, half way between the support and the free end of the cantilever (figure 1), for the two scenarios outlined above. For hinging, all of the actuators are imparted the same extension (figure 3a). For twisting, different extensions are imposed on each actuator in the sequence described below (figure 3b). Resistance of the Structure. Preliminary results for hinging displacements provide an indication of the resistance of the structure to actuation, expressed through the magnitudes of the forces on the actuators, FA, relative to those associated with the passive loads. The calculations also reveal that FA is largely governed by the bending deformations induced in the core members and the face-plate immediately above the actuators" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000977_8.999616-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000977_8.999616-Figure1-1.png", + "caption": "Fig. 1. Axially symmetric parabolic reflector antenna and field components.", + "texts": [ + " In Section II, a brief summary of the theory of CAPs is given, and the two differential equations for the CAPs are discussed. Section III presents the discretization of these differential equations. The meshes for electric and magnetic azimuthal fields are offset to facilitate the discretization. The occurrence of singularities of the FD equations are discussed. In Section IV, illustrative results of induced current densities on the surfaces of reflector antennas are presented. An axially symmetric parabolic reflector is shown in Fig. 1. Morgan et al. [13] studied the electromagnetic field problems involving continuously or discretely inhomogeneous axially symmetric media and showed that the CAP is an economical 0018-926X/02$17.00 \u00a9 2002 IEEE way to deal with such type of problems. Using the CAPs, the 3-D problem can be transformed into a series of 2-D problems, hence, an axially symmetric parabolic reflector antenna with very large dimensions can be analyzed efficiently by the FD method. In the application of the CAPs, the fields need be expressed in azimuthal Fourier series in ", + " Letting and be the electric field and magnetic field in free space, respectively, we have (1a) where is the intrinsic impedance of free space. is the cylindrical coordinates in space. and can be expressed as (1b) Substituting (1b) into the Maxwell\u2019s equations, we have (2) From (2), for a given number , the corresponding differential equations for the azimuthal components are derived as (3) in which , is the wave number of free space. It can be seen that the azimuthal components and are coupled with each other and are independent of variable . Referring to Fig. 1, the components of and in the and directions can be expressed as functions of their azimuthal components (4a) (4b) where . There exist two cases for which (3) can be simplified: 1) , and 2) ( ). In [14], the case with was discussed and the differential equations reduced to for for (5a) For or , (3) can be simplified as (5b) B. Boundary Condition and Induced Current Density For scattering problems of metallic bodies of revolution, the boundary conditions for azimuthal components and on the metallic surfaces are (6a) (6b) where the superscripts and denote, respectively, scattered and incident components" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000587_0029-5493(65)90101-9-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000587_0029-5493(65)90101-9-Figure5-1.png", + "caption": "Fig. 5. Notation for cylindrical part of shell", + "texts": [], + "surrounding_texts": [ + "and is bounded by the i n t e r s e c t i ons of th i s cy l i n - de r with the plane m x = 0 and the p l anes\n2 n 0 - n x = ~ l . (7)\nThe faces IV and V l ie on the p a r a b o l o i d s\nm x = :~\u00bd [2 - ( 2 n 0 - 1) 2 - ( 2 n 0 - 2 n x - 1) 2] . (8)\nThis y ie ld locus has the o r ig in of coo rd ina t e s as a c en t e r of s y m m e t r y . I t has uniquely d e t e r - mined suppor t ing p lanes at a l l po in ts except those on the p a r a b o l i c a r c s (4) and (5), the s e g - ment CD, and the poin ts obtained f rom these by s y m m e t r y with r e s p e c t to the or igin .\nAn exact y ie ld locus for a sandwich she l l , if M 0 and N O a r e t aken to r e p r e s e n t the y ie ld m o - ment and the y ie ld fo rce of the sandwich she l l , was d e s c r i b e d by Hodge [9] and is shown in fig. 2. If, however , M 0 and N O a r e given the va lues c o r r e spond ing to a so l id she l l , th is po lyhedron r e p r e s e n t s an approx ima t ion to the exact y ie ld locus of fig. 1. The c o r r e spond ing f a c e s of the po lyhedron l ie in the fol lowing p lanes :\nface I n O = 1\nII n O - n x = 1\nH I n x - m x = - 1\nI V 2 n 0 - n x + m x = 2\nV 2 n 0 . n x - m x = 9\n(9)\nface I nO = 1\nII n O - n ~ o = 1\nIII m ~ o = l\ni v n \u00a2 = 1\nOther p l anes can be obtained by s y m m e t r y with r e s p e c t to the or ig in .\nTh is y i e l d su r f ace was f i r s t p r o p o s e d by D ruc ke r and Shield [11] for ro ta t iona l ly s y m m e t - r i c she l l s . I t can a lso be obtained for cy l i nd r i ca ! she l l s by e l imina t ing rn 8 f rom the Hodge \" twomoment l i m i t e d - i n t e r a c t i o n \" su r face [18]; i t i s r e f e r r e d to a s a \"one -momen t l i m i t e d - i n t e r a c - t ion\" y ie ld sur face .\nThis hexagonal p r i s m is much s i m p l e r but is by no means as good a f i t at some points . 1 R e - ducing a l l i t s v e r t i c e s by the f ac to r \u00bd(5~,- 1) p r o d u c e s an i n s c r i b e d y ie ld sur face . A t h r e e - q u a r t e r s i ze p r i s m , however , l i e s within the a c -\nAnother approx ima t ion to the y ie ld su r face i s shown in fig. 3. Th is y ie ld condi t ion i s defined by the fo l lowing eight p l anes :", + "E4 AXISYMMETRIC INTERSECTING SHELI.~ OF REVOLUTION 89\ntual y ie ld su r f ace ove r an extended range of v a l - ues of p r a c t i c a l i n t e r e s t for p r e s s u r e v e s s e l s .\n3. APPLICATION OF THE THEORIES TO INTERSECTING SHELLS\nIn o r d e r to be able to make use of the y ie ld loci for i n t e r s ec t i ng she l l s as r e p r e s e n t e d in sec t ion 2, some approx ima t ions to the shel l conf igura t ion mus t be in t roduced. Although the fo l - lowing a rgumen t holds t rue for any she l l of r e v o - lution, we r e s t r i c t o u r s e l v e s to the spec ia l c a se of a r a d i a l out let f rom a s p h e r i c a l she l l subjec t to in te rna l p r e s s u r e as shown in fig. 4.\nF o r the c y l i n d r i c a l p a r t of the she l l , the equat ions of equ i l ib r ium, with the notat ion of f ig. 5, a r e given by:\ndQ NO dMx + - - r = P ' dx - Q ' N x = \u00bdpr \" (10)\nF o r N o = const the in teg ra t ion of eq. (10) y i e ld s to\n(11) M x = \u00bd( , - ~ - ) x ' 2 + A x , B .\nThe constants A and B can be evaluated using any suitable stress distribution, provided this assumed stress field nowhere exceeds any of the chosen yield surfaces of section 2, e.g.\n/\na t x = l , Q = 0 , M x = M c , N o = N c , (12) a t x : 0 , Q : Q ' , M x : M ' c , NO : N c . T where Mc, M c, N c a r e the a p p r o p r i a t e va lues of the m o m e ~ s and t h ru s t on the y ie ld sur face .\nS i m i l a r l y for the s p h e r i c a l p a r t of the she l l , equ i l ib r ium equat ions with the notat ion of fig. 6 become\"\nN 0 sinq~ + Q cosq~= \u00bdPR sinq) ,\ns i n e + N 0 sin~0 + d~(Q s i n g ) = p R s i n ~ , (13) N, dM\u00a2 de sin(p + (Mq) -Mo)cosq ) - Q R sinq) = 0 .\nIn o r d e r to mee t the r e q u i r e m e n t s govern ing the use of the y ie ld su r face , MO must be e l i m i - nated f rom the t h i rd equation. Two poss ib l e ways of achieving th is may be dev i sed by set t ing:\na) M O = 0 , (14)\nb) MO = M~o \u2022 (15)\nIn t roduct ion of the condi t ion (14) in the equ i - l i b r i um equat ions (13) and in tegra t ing l eads to:\nQ = (\u00bdPR -No)cp + C ,\n=R(\u00bdPR .-N0)(1 - \u00a2 cot q)) - CR cot ~o + ~ (16) Mq)\nNq~ = \u00bdPR - (\u00bdPR - NO)(p cot \u00a2 - C cot ~0 .\nS imi l a r l y , by in se r t i ng condit ion (15) in eq. (13) and in tegra t ing , we obtain:", + "Q = (\u00bdiPR -No)~o + C ,\nMq~ = [\u00bd(~R -No)q~2 +Cq)]R + D , (17)\nN~o = ~OR - (~PRI \"No) s~ + sin~C\nHere too, the integration is performed under the condition that circumferential membrane forces are constant. By invoking a suitable stress field, the constants C and D may be evaluated thus:\na t q ~ = / 3 , Q=O , Mq~=Ms, N O = N s , (18) at~o = a , Q = Q\" Mq~ ' , = M s , N O = N s \u2022\nwhere Ms, Ms, N s lie on the yield surface. Satisfying force and moment equilibrium at the sphere-cylinder junction leads to an expression for the collapse pressure in terms of the shell geometry. As described in section 1, however, this will establish only a \"lower bound\" on the collapse pressure, provided approprmte inequalities on M~o(Mx) , Nq)(Nx) and NO, depending upon the yield surfaces of section 2, are satisfied. The results thus obtained are restr icted to those shell configurations where the assumed s t ress profile, NO= const, and conditions (14) or (15) can be imposed.\nBy choosing his boundary conditions to satisfy face I of the yield surface of fig. 1 and the set of eq. (16), Lind [19] has been able to establish such a collapse pressure. The complexity of the equations obtained makes it necessary to use a trial and e r ro r procedure. Gill [20], on the other hand, by taking face I of the hexagonal pr ism yield locus of fig. 3, in association with the boundary conditions (12) and (18) and the set of eq. (17), arr ives at a fairly simple expression for the collapse pressure. The solution by Cloud [21] may also be derived as a particular case of Gill 's expression for the collapse load.\nAn upper bound on the collapse pressure can be found by equating the external rate of doing work to the internal rate of energy dissipation for a kinematically admissible pattern of three hinge circles. A velocity field of the form\nU= c[1-cos(E-q)] , W= -csin(~-\u00a2)\nsatisfies these hinge conditions. Here too, the contribution of M 0 to the work equations, in spite of the change of the circumferential curvature, must be neglected. Gill has attempted such a procedure, but instead of minimizing the parameters which locate the positions of the hinge c i r -\ncles in the work equation, has taken them from the so-called lower bound solution.\n4. DISCUSSION\nIn the previously described approximate theories of rotationally symmetric shells, based upon neglecting entirely the circumferential moment M 8 in both the yield condition and equilibrium equations, the solutions are reliable only for those regions of the shell some distance from the axis of revolution. Moreover, the assumption that there is no interaction between meridianal moment M~ and membrane forces NO and N~, makes the approximations more unrealistic. For instance, in the case of the yield surface comprising a circumscribed hexagonal prism, the upper bound (or kinematic t h e o r e m ) c l e a r l y gives an upper bound on the collapse pressure. On the other hand, merely satisfying equilibrium with such a circumscribed yield surface gives an approximate result which cannot be identified either as a lower or an upper bound. Fur thermore, the elimination of M 0 from the equations representing the yield surface, makes it dubious whether the flow rule, following from the identity of plastic potential and yield surface, could be applied to geometrical entities obtained as a r e - sult of operations performed.\nSo long as large factors of safety are used, the approximate solutions described above are satisfactory for design purposes. As the r e - quirements of high pressures and economy of design become more stringent, however, the needs for more exact analyses are pressing. In order to achieve such solutions, consideration must be given to certain .aspects of the problem. It was shown that the yield surface for a cylindrical shell is, in general, non-linear, even though the yield condition may be piecewise linear (Tresca). The same applies to the shell configuration. The ambiguities of the solutions arise from this non-linear characterist ic and one method of recovering a piecewise linear problem is to approximate the uniform shell to an idealized sandwich shell. Another method is by a piecewise linear approximation to the yield surface.\nThe current investigations are concerned with establishing proper bounds to the intersecting shell problem by considering the approaches described above. Comparisons between existing approximate theories should be made in order to guide the designer in choosing appropriate solutions by delineating their ranges of validity." + ] + }, + { + "image_filename": "designv6_24_0000231_s00170-017-0346-6-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000231_s00170-017-0346-6-Figure3-1.png", + "caption": "Fig. 3 Changing of the loading scheme for conventional (a, b) and large radius bending (c, d)", + "texts": [ + " For conventional air bending, one can observe that during forming, the radius of the inner surface of the formed plate is larger than the radius of the punch. However, Fig. 2 shows that for large radius bending, the actual radius of the inner surface of the plate is smaller than the punch radius. In the current section, the bending moment diagrams are depicted according to rules presented in [19]. Conventional air bending is described as a central loading scheme, namely three-point bending. In the early phases of the forming process, the shape of the bending moment diagram is triangular (see Fig. 3a). During the actual plastic forming process, the shape of the bending moment diagram looks like a triangle with rounded sides (see Fig. 3b). In contrast with conventional air bending, the loading scheme for large radius bending depends on the bending angle. The contact points move from the center of the plate towards the shoulders of the die. The most suitable description for this behavior is four-point bending. Figure 3c shows the loading scheme in the early stages of loading application, with the corresponding moment distribution. When the punch displacement increases, the bending angle increases, the multi-breakage appears, and the punch contacts the plate in two contact zones (see Fig. 3d). Figure 3d presents the shape of the bending moment distribution for this loading scheme, which has changed from a triangle to a trapezoidal pattern on the sides and a top slightly inclined depending on the friction situation. For forming operations performed with the punch of a large radius, the more pronounced multi-breakage effect is expected due to more considerable change of bendingmoment or contact point position c (see Fig. 3d). The springback angle is the reduction of the angle through elastic unloading after the bending force is removed; it is calculated as [9] \u0394\u03b2 \u00bc M E 0 I dL \u00f01\u00de where M is the internal bending moment, E\u2019 the plane strain modulus (E\u2019 = E / (1 \u2212 \u03bd2)), \u03bd the Poisson ratio, I the second moment of area about the middle axis (I = bt3/12), and L the length of the plate in a curvilinear axis. This section assumes a fully elastic behavior of the plate; i.e. in the further discussion, only the elastic bending moment makes a difference for the springback and the total bending force. For the complete treatment of elastoplastic bending, the reader can refer to [9]. The elastic bending moment patterns for three- and four-point bending schemes are shown in Fig. 3, where P is the bending force, c the distance of the contact points, and d the horizontal shift of the contact point with the die. Indices \u201cc\u201d and \u201cm\u201d refer to conventional and multibreakage bending, respectively. For the elastic case, the springback angle, as the integral of Eq. 1, can be estimated as the area of the hatched region in Fig. 3d; this area is larger than the one in Fig. 3b, so a stronger tendency to springback in comparison with three-point bending is expected. This issue underlines the importance of taking into account the multi-breakage effect while calculating the springback for large radius bending. The correct understanding of how the loading scheme in large radius air bending progresses is essential to describe the large radius bending effectively and correctly. The multibreakage effect is a crucial phenomenon, and it is necessary to take it into account as it affects the bending force and consequently the final bent shape\u2014the springback and the bend allowance", + " From the geometrical description, the length of the straight part of the plate midplane is as follows: ld \u00bc w 2cos\u03b8 \u2212 Rd \u00fe t 2 \u00fe RA tan\u03b8 \u00f02\u00de la \u00bc \u03b8RA \u00f03\u00de where w (see Fig. 4c) can be evaluated in function of the die opening w0: w \u00bc w0 \u00fe 2Rd tan 45 \u2212 \u03b1d 4 \u00f04\u00de Figure 4b shows the normal and tangential reaction forces on the die shoulders and their levers from point C. Since the plate wraps closely around the punch, the radius of curvature and the moment are the same in points A and C. The flat top of the trapezoidal profile (see Fig. 3d) can represent this. From this assumption, the levers can be determined by the following: blev \u2261 ld \u00f05\u00de blev2 \u2261 t . 2 \u00f06\u00de The bending moment is specified as follows: MA \u00bc Fnblev \u00fe Ftblev2 \u00f07\u00de For a Coulomb friction approximation and a friction coefficient \u03bc, the relation between Ft and Fn is given by the following: Ft \u00bc \u03bcFn \u00f08\u00de Taking into consideration Eq. 8, the normal force is equal to Fn \u00bc MA blev \u00fe \u03bc blev2 \u00f09\u00de By definition, the unit moment \u03c3* A is the bending moment per unit of width and per squared unit of thickness" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001044_j.proeng.2017.01.201-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001044_j.proeng.2017.01.201-Figure1-1.png", + "caption": "Fig. 1. Structure of steel channel.", + "texts": [ + " Peer-review under responsibility of the organizing committee of the 13th Global Congress on Manufacturing and Management focused on frame mass and model optimization analysis. In this paper the large-scale software ABAQUS was used to build a parametric finite element model for the frame of SX360 dump trucks, in order to provide structural optimal design for the frame, ensure reasonable frame design and improve its overall operating performance [2-5]. Main structural parameters of the frame: Q235 steel material, 36 /kg1085.7 mm in density, 1500mm in length, 500mm in width, with 5# steel channel used as standard parts to support at 520mm and 980mm..As shown in Fig. 1. At first in the functional module SKETCH in ABAQUS, a two-dimensional section was drawn for the frame and then in the functional module PART, it was stretched in different faces to form a three-dimensional physical model, only a half of which would be extracted for analysis and calculation based on symmetric frame structure and load. The simplified three-dimensional physical model of the frame is shown in Fig. 2. After incorporating with proper materials, section attributes and assembly parts, divide the built threedimensional physical model into grids" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000514_1.c034448-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000514_1.c034448-Figure2-1.png", + "caption": "Fig. 2 NASA generic transport model (GTM).", + "texts": [ + " Additionally, it is of interest to consider flight control requirements with differential thrust control, vertical tail sizing for the engine-out take-off condition, and the overall system architecture and desired benefits. An MDO framework can be developed that includes aircraft geometry, aerodynamics, propulsion, wing structure, aeropropulsive-elasticity, stability, flight control, and system analysis. The baseline aircraft model is the NASA GTM that has a similar planform as the Boeing 757 aircraft equipped with conventional turbofan engines as shown in Fig. 2. The distributed propulsion aircraft is modified from the baseline GTM by replacing the turbofan engines with four propulsors and one generator per wing, referred to as single-generator distributed propulsion aircraft, and two generators per wing, referred to as dual-generator distributed propulsion aircraft. The generators, which are essentially a gas turbine engine core, generate power to drive the propulsors that are modeled as electric fans. For the single-generator configuration, the four propulsors are positioned along the normalized wing stations \u03b7 2y\u2215b 0", + " The flexiblewing configuration iswith the baseline bending stiffness and the reduced torsional stiffness equal to 50% of the baseline value. In Figs. 14 and 15, the spanwise lift distribution at Mach 0.8 and 30,000 ft normalized to the mean aerodynamic chord (MAC) c along the normalized wing station \u03b7 is plotted for the single-generator distributed propulsion aircraft with the stiff wings and flexiblewings, respectively, to illustrate the effect of the thrust distribution across the wing span on the lift distribution, hence L\u2215D. The baseline configuration refers to the GTM with two conventional turbofan engines as shown in Fig. 2. The uniform thrust case corresponds to the equal thrust generated by the propulsors. The 50%\u2215 50%\u2215 \u221250%\u2215\u221250% denotes a 50% increase in thrust of propulsor 1 and Fig. 11 Coupled aero-propulsive-elastic analysis framework for distributed propulsion aircraft. D ow nl oa de d by U N IV E R SI T Y O F M A N C H E ST E R o n M ar ch 1 2, 2 01 8 | h ttp :// ar c. ai aa .o rg | D O I: 1 0. 25 14 /1 .C 03 44 48 propulsor 2 relative to the uniform thrust value and 50% decrease in thrust of propulsor 3 and propulsor 4" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003845_8.467637-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003845_8.467637-Figure5-1.png", + "caption": "Fig. 5. Non-offset antenna configuration, yo^ = 0), showing the subreflector, the main reflector, and representative feed rays. (a) scanned beam. (b) nonscanned beam.", + "texts": [ + " In comparison, a flat subreflector has a simpler geometry and can be rotated without requiring feed repointing. The optimum planar subreflector location for any scan angle is the one that images the virtual source in the focal arc to the stationary feed point. The geometry of Fig. 2(b) is adopted in this paper because a smaller subreflector can be used to intercept the rays propagating from the feed to the main reflector. This subreflector can be tilted to redirect rays for the various scanning directions. Fig. 5(a) shows an example of the extreme scanning operation of a non-offset antenna configuration = 0). The subreflector plane is perpendicular to the line connecting the feed point source A and the focal point F at the midpoint of that line. The limits of the subreflector are determined by the efficiently illuminated reflector portion (between E and I). The center of the subreflector (H,) and its normal determine the rotation center (B), located on the z = 0 plane and the rotation radius segment R = H,B. The subreflector is tilted for scanning the field-of-view about an axis defined by the rotation center B and the axis unit which is parallel to the subreflector and the z = 0 planes. The subreflector location for the nonscanning operation is shown in Fig. 5(b) where the center of the subreflector has moved to H,. The 1026 IEEE TRANSACTIONS ON ANTENNAS AND PROPAGATION, VOL. 43, NO. 10, OCTOBER 1995 focal point V of the Up is taken as the specular image of A about the nonscanned subreflector plane. This choice simplifies the subreflector scanning motion to a rotation with axis normal to the plane containing F, V, F', and A. The feed location must be optimized taking into account the following conditions: the length of the radius R should be minimized to simplify the subreflector motion; the feed-cone angle Of, corresponding to the subreflector edge points (see Fig. 5), should be almost constant to ensure efficient subreflector illumination for scanned beam conditions; and the feed should be near the main reflector surface for easy mechanical support. Once the center and radius of rotation of the subreflector are determined, the motion of the subreflector must be specified to explore the rest of the field-of-view. To model this motion, we define an arbitrary parameter ,8 that has a value ,8 = 0 for the boresight beam and values p = 1 and p = -1 for the two extreme scan directions, respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000855_0954406220936732-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000855_0954406220936732-Figure8-1.png", + "caption": "Figure 8. Deformation under single moment: (a) deformation under Mx; (b) deformation under Mz.", + "texts": [ + " By observing the stiffness matrix, we also found that the x-axis displacement stiffness of the force sensor is larger, compared with that of y- and z-axes, which are identical in value. The rotation stiffness of y- and z-axes is the same and larger compared with that of x-axis. We carried out the finite element analysis in ANSYS Workbench. The moving platform exerts the force of 1000N and the torque of 10N m, respectively. The deformation of the elastomer in the specified direction under the action of each single-dimensional force/moment we obtained, which are shown in Figure 7, Figure 8, and Table 2. The numerical value is compared with the simulation value and the results were concluded in Table 2. As presented in Table 2, the deformations of the moving platform calculated by equation (28) are slightly smaller than the simulation values, and the errors are remained within 10% in which the maximum error is 9.05% in the deformation of y/z axis, and the minimum error is 5.72% in the rotation deformation of y/z axis. From the simulation results, it can be found that the deformation of x-axis is much smaller than that of y- and z-axes, and the rotation deformation of x-axis is larger than that of y-axis and z-axis, which is consistent with the results of the calculated stiffness matrix in equation (28)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002862_detc2007-34856-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002862_detc2007-34856-Figure7-1.png", + "caption": "Figure 7: Fixed-Pitch CVT [9]", + "texts": [], + "surrounding_texts": [ + "The fixed-pitch CVT (see Fig. 8), developed by Kenneth B. Hawthorn [9], operates on principles similar to the pivot-arm CVT. It contains sprockets held by a carrier mechanism that controls the radial position of the sprockets, thus allowing the effective radius of the CVT to be changed, and in so doing the transmission ratio. The design also has only two points of contact per chain, one on the power side and one on the load side. Multiple chains are incorporated to provide multiple contact points which maintain engagement between the power and load sides through their rotations. Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 02/04/2016 Term This embodiment is likewise subject to the non-integer tooth problem, in that the distance between the sprockets can change in a continuous manner. This continuous change would allow the distance between the sprockets, specifically where they would mesh with a chain (assuming the sprockets could not rotate), to assume values not divisible evenly by the pitch of the chain. As in the previous embodiment, this would cause slack to occur in the chain as the effective radius of the CVT decreased, and the chain would skip off the sprockets when the effective radius increased. To overcome the non-integer tooth problem, Hawthorn proposes a different method of allowing reorientation of the sprockets than the one-way clutches proposed by Christensen. He instead proposes a specially designed sprocket (see Fig. 8), called the power sprocket, that is able to freely rotate upon its supporting shaft, thereby allowing the chain to engage properly with the sprocket. Once the chain becomes fully seated on the sprocket, however, it causes the sprocket to lock on its supporting shaft, thereby eliminating the sprocket\u2019s rotation, allowing it to transmit torque. 4 Copyright \u00a9 2007 by ASME s of Use: http://www.asme.org/about-asme/terms-of-use" + ] + }, + { + "image_filename": "designv6_24_0003088_ias.2006.256631-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003088_ias.2006.256631-Figure4-1.png", + "caption": "Fig. 4. Schematic representation of (left) and measured (right) estimation error resulting from inexact decoupling of secondary saliencies in the negative sequence carrier signal current.", + "texts": [ + " The signal processing consists of four steps: 1) coordinate transformation, 2) separation of the desired carrier signal component from the overall signal using bandpass filtering, 3) compensation of any remaining undesired, deterministic content, i.e., secondary saliencies, and 4) estimation of the saliency position. Ideally, the carrier signals entering the \u201csaliency position estimation\u201d block shown in Fig. 3 would only consist of rotor position related components. However, this is never achieved in practice due to both incorrect decoupling of secondary saliencies, as well as, the presence of other \u201cnoise\u201d in the signals. Fig. 4 graphically shows the effect on rotor position estimation using the negative sequence carrier signal current when the component containing the desired information, IcnR, is contaminated by additional components, Icnhi (8). These additional components form a disturbing signal \u03a3Icnhi, which has a maximum potential value of |\u03a3Icnhi|max. i cn qds_cn = IcnR e jR\u03b8rm+ \u03a3Icnhi e jhi \u03b8i (8) A key parameter determining the accuracy and robustness of the estimated position is the magnitude of the rotor position dependent component, IcnR, (signal) relative to the other components, Icnhi, (noise). The total harmonic distortion, THD, caused by the undesired \u201cnoise\u201d was found to be an insightful metric for quantifying this relationship [7]. The THD of the negative sequence carrier signal current is calculated using (9). THD(i cn qds_cn) = \u03a3\u03b9cn2 \u2013 \u03b9cn(R\u00b7\u03c9rm)2 \u03a3\u03b9cn2 (9) where \u03b9cn=FFT(i cn qds_cn) and \u03a3\u03b9cn2=\u03a3+bw n=\u2212bw\u03b9cn(n)2 with \u03b9cn(R\u00b7\u03c9rm) being the desired signal, IcnR, in Fig. 4. Similarly, the THD of the zero sequence carrier signal voltage can be calculated using (10). THD(v c 0qds_c) = \u03a3\u03c5c02 \u2013 \u03c5c0(R\u00b7\u03c9rm)2 \u03a3\u03c5c02 (10) where \u03c5c0 =FFT(v c 0qds_c) and \u03a3\u03c5c0 2 =\u03a3+bw n=\u2212bw\u03c5c0(n)2 It should be noted that calculation of the THD using (9) and (10) is defined using the input signals to the \u201csaliency position estimation\u201d block in Fig. 3, after the decoupling of any secondary saliencies, and assumes the filtering removed all content outside the filter bandwidth. Fig. 5 shows the THD of the negative sequence carrier signal current and the zero sequence carrier signal voltage, as a function of the carrier frequency and voltage magnitude", + " The THD of the zero sequence carrier signal voltage is usually smaller than that of the negative sequence carrier signal current, with the difference increasing as the carrier frequency increases. It can also be noted from Fig. 5 that errors in decoupling cause similar values of THD in both signals. This result is not unexpected since the error in decoupling dominates the \u201cnoise\u201d contribution in the THD calculation and therefore results in almost the same value for both signals. The exact relationship between the THD of the carrier signals and the overall accuracy of the estimated position can be obtained for the case of a single, known disturbing component, Icnhi (see Fig. 4 and (8)). The curve for this exact relationship is shown in Fig. 6 (solid lines). The maximum error with a single secondary saliency is limited to \u224812.9 mechanical degrees, which occurs when Icnhi\u2265IcnR and a THD of 1/\u221a2. This value corresponds to a rotor slot pitch and means, in practice, that a jump in the estimated saliency position to the next rotor bar occurred. This represents instability in the sensorless estimate [13]. In practice, the carrier signal contains multiple disturbing components" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003772_313-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003772_313-Figure8-1.png", + "caption": "Figure 8.", + "texts": [ + " Hence the general effect of applying a large stress to the specimen is to reduce its effective elastic modulus, bring it to a condition in which hysteresis loops can be obtained, and replace the low elastic limit of the unstrained specimen by an elastic-hysteresis limit which is much greater (the limiting stress being five or ten times as great) and shares all the properties of the actual elastic limit. The type of hysteresis loop obtained in the investigations cited above, and discussed by T~ml inson(~) and by Prandtl('), is that shown in figure 8. Such a loop depends upon the deviation of the line of increasing stress ABCD from a straight line at C. It requires, then, that the stress applied, at any rate to part of the specimen, should exceed the elastic limit. The loop is a result of permanent set of the specimen occurring at the extremes of stress, and can only be closed if equal stresses of opposite signs are applied alternately. Since it would presumably not exist if the reversal of the stress were made at the point C, it depends upon the material entering the plastic or overstrained state, and the name plastic hysteresis may be suggested for this type of loop" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002482_9783527631506-Figure7.41-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002482_9783527631506-Figure7.41-1.png", + "caption": "Figure 7.41 Leapfrog realization of the elliptic filter of Example 7.9.", + "texts": [ + "39b will realize Eq. (7.57) provided gZj = gmjZ\u2032 j , that is Z\u2032 j = g gmj Zj for even j (7.60) Thus, the set of I\u2013V relations given by Eq. (7.53), and hence the ladder of Figure 7.38, can be realized by the leapfrog structure utilizing OTAs, as shown in Figure 7.40. It is noted that the various Z\u2032 j ( j odd or even) are due to R, L, and C elements, which can be easily realized using OTA-C structures. We now illustrate the procedure by an example. Example 7.9. Realize the third-order elliptic LP filter shown in Figure 7.41a by the leapfrog structure discussed above. Using the method discussed above, we see that the filter can be realized by the structure of Figure 7.40, where the various impedances Z\u2032 1, Z\u2032 2, Z\u2032 3, and Z\u2032 4 are given by Eqs. (7.59) and (7.60). Hence, Z\u2032 1 = 1 ggm1 G1, Z\u2032 3 = 1 ggm3 (sC3 + 1 sL3 ) Z\u2032 2 = g gm2 1 sC2 , Z\u2032 4 = g gm2 1 (G4 + sC4) Let us choose gm1 = gm2 = gm3 = gm4 = g = gm = G1; then the above expressions become Z\u2032 1 = R1, Z\u2032 2 = 1 sC2 , Z\u2032 3 = s c3 g2 m + 1 s ( 1 g2 mL3 ) , Y \u2032 4 = (G4 + sC4) We see that Z\u2032 1 is a grounded resistor of value R1, while Z\u2032 4 = 1/Y \u2032 4 is a parallel combination of a grounded capacitor C4 and a grounded resistor R4", + " The impedance Z\u2032 3 is a series combination of an inductor of value C3/g2 m and a capacitor of value g2 mL3. We have the option of having the inductor or the capacitor grounded. Even though the former arrangement uses only two OTAs, 7.6 High-Order Filters Using Switched-Capacitor (SC) Networks 245 the latter arrangement is preferable since in an IC implementation, it is better to have grounded capacitors. In such a case, the ungrounded inductor can be realized using three OTAs and a grounded capacitor (see Table 5.10, Row J). The complete structure realizing the ladder network is shown in Figure 7.41b. There are a number of other multiloop structures utilizing OTAs that can be used to realize a general high-order filter, and the reader is referred to Deliyanis, Sun, and Fidler (1999) for more details. 7.6 High-Order Filters Using Switched-Capacitor (SC) Networks In SC filters, we can follow, for high-order filter realization, a scenario similar to that in the case of continuous filters. Thus, we can build a given high-order filter using a cascade of second- and first-order SC networks. We can also use second-order sections in a feedback structure to generate a given z-domain transfer function" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001212_6.2008-5955-Figure18-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001212_6.2008-5955-Figure18-1.png", + "caption": "Figure 18. Medium Range Commercial Jet Parametric Geometry", + "texts": [ + " The vehicle was chosen to perform the same mission and have the same configuration characteristics as the Boeing 747-100 with a mission profile as shown on Figure 16. Thrust loading and wing loading of the vehicle are set to equal 0.25 and 132 lb/ft2 respectively. 17 of 23 As described before, aircraft geometric information is hierarchically modelled with support from the parametric geometry modeller that provides a unified geometric description to all disciplines. The aircraft geometry used in the validation analyses as developed by the parametric modeller is shown on Figure 17, Figure 18 and Figure 19 respectively. Comparison of the primary sizing results from the new design environment and the real aircraft data for the evaluated examples is shown on Table 1, Table 2 and Table 3. The implemented design environment correlated well with key aircraft parameters, not only in terms of weight but also in estimated performance such as takeoff and landing field lengths. 18 of 23 19 of 23 Note as well in Table 3 how the different aerodynamic methods used for drag buildup provide good correlation with published aerodynamic data" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003990_jjap.56.06gg01-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003990_jjap.56.06gg01-Figure1-1.png", + "caption": "Fig. 1. (Color online) Schematic configuration of the minimal-fab machine. (a) Photo-image of the minimal-fab machines, (b) hematic-sealed wafer transfer vehicle named minimal shuttle, and (c) load-lock system named PLAD system.", + "texts": [ + " The minimal-fab concept includes three distinct features as (1) wafers size is a halfinch (12.5mm) and processing single wafer, (2) processing and transferring the wafer with identically clean level of a super cleanroom but without a cleanroom, (3) the machine size has compact dimensions of 144 cm in height, 30 cm in width, and 45 cm in depth.1) To keep a clean wafer without a cleanroom, the hermetic-sealed minimal shuttle, particle lock air-tight docking (PLAD) system and clean process chamber are equipped in the minimal-fab machine, as shown in Fig. 1. After the proposal of minimal-fab concept, many kinds of minimal-fab machines have speedily been developed except for the ion implantation (I=I) minimal-fab machine which is under development. Meanwhile, the gate-last processes for metal\u2013oxide\u2013semiconductor field-effect-transistors (MOSFETs) fabrication by using the developed minimal-fab machines have actively been developed.2\u20135) In such minimal MOSFET fabrications, the solid source diffusion by spin on dopants (SOD) was used to the formation of source\u2013drain (SD) regions because the minimal I=I machine is under development as mentioned before" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001000_s12206-017-0605-3-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001000_s12206-017-0605-3-Figure2-1.png", + "caption": "Fig. 2. The model of IRCSHS.", + "texts": [ + " In addition, the tension in the suspension cables should be properly allocated to meet the safety and performance requirements. Hence, it is necessary to evaluate the design by calculating the dynamic responses of the platform and the cables. Asynchronous motion velocities and different frequencies will cause imbalances in cable tension. Nonlinear dynamic behaviors, such as no-smooth phenomenon, will appear at certain range of asynchronous motion velocities and special frequency. 2. Description of IRCSHS The Incompletely restrained cable-suspended swinging and heaving system (IRCSHS) is designed and shown in Fig. 2, which composed of two drums, sheaves, cables, base frame, suspended platform and heaving system. The cables are winded on the drums respectively, which is shown in Fig. 2. The IRCSHS is equivalent to the suspended platform hang with four suspension cables. The four connected points of the suspension cable are symmetrical distribution around the suspended platform. The cables are fixed on the drums by a pressing plate respectively, and the drums simultaneously control the suspended platform posture by changing the cable length with a periodic motion. The IRCSHS uses suspension cables instead of conventional links, which brings the advantage of the simple structure and low power consumption but at the same time introduces a more complex dynamic behavior" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001247_j.ijleo.2013.09.021-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001247_j.ijleo.2013.09.021-Figure6-1.png", + "caption": "Fig. 6. Optimum structure of Cassegrain antenna.", + "texts": [ + " Therefore, the relation between the system block rate and the antenna gain factor in the case of \u03b5 = 0.1, \u03b5 = 0.2, \u03b5 = 0.3 have been obtained, shown in Fig. 5 3.3. Optimum structure of Cassegrain antenna Though a Cassegrain configuration is widely used for expending the beam in a reflecting mirror system, the loss caused by the secondary mirror is seriously, which would affect the effect of transmission [8]. Therefore, the traditional structure have been substituted by an improved structure made up of two pyramid, which is shown in Fig. 6. In this antenna, the primary mirror is just full of laser beam 2. The Double pyramidal system is shown in Fig. 7. Gaussian beam have been reshaped as hollow beam, it has been researched that the existence of hollow laser beam has been assured under the refraction disciplines. Because laser beam is parallel beam 2 with tiny divergence and the radius of beam waist is near to the antenna, the beam radius L d 2R where p is the laser power, L is the length of Double pyramidal system, is apex angle, 2R is the diameter of the hollow beam" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003759_a:1008935628281-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003759_a:1008935628281-Figure2-1.png", + "caption": "Figure 2. VIRTUE\u2019s coordinates. From a geometric point of view, VIRTUE is located at EC and fixates point EF . Each of the cameras are located at equidistant points on the rim of a circle of radius r . The one remaining degree of freedom, the twist of the head, is defined by 8.", + "texts": [ + " Results showing the application of the algorithm when applied to a number of test objects and an experimental evaluation of the algorithm is provided in Section 5. Finally Section 6 draws some conclusion and discusses possible future research directions. VIRTUE: A VIRtual TrinocUlar stEreo-head, is a trinocular stereo head implemented using a single camera mounted on the end-effector of a robotic manipulator. VIRTUE\u2019s three virtual cameras form the vertices of an equilateral triangle and fixate points which lie on the normal which projects to the triangle\u2019s center. This is sketched in Fig. 2. Ignoring torsion of individual cameras, VIRTUE has eight degrees of freedom. The head is centered at EC , fixates a point EF , and has a variable baseline r . VIRTUE is also free to rotate by 8 about the vector EC \u2212 EF . The details of VIRTUE\u2019s design, and its forward and inverse kinematics are described in Lang and Jenkin (1996a, b). Although VIRTUE could be constructed as a special purpose robotic device, the single eyed VIRtual TrinocUlar stEreo-head (VIRTUE) in Fig. 3 offers a number of advantages over the use of a specially designed stereo-head" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002575_125906-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002575_125906-Figure3-1.png", + "caption": "Figure 3. Pre-tightening structure composed of solid cylindrical GMM. (a) Giant magnetostrictive structure. (b) Pre-tightening structure.", + "texts": [ + " Using the GMM structures for double-nut ball screw\u2019s pre-tightening is still a new area for research and few documents in this field have been found. GMM (Terfenol-D) can be made into solid or hollow cylinder, as shown in figure\u00a0 1, they can both be used for ball screw pre-tightening. The pre-tightening structure composed of the hollow cylindrical GMM is as shown in figure\u00a02, where: 1 & 4 are nuts, 2 is force sensor, 3 is the coil and 5 is the hollow cylindrical GMM. The giant magnetostrictive structure composed of a solid cylindrical GMM is as shown in figure\u00a03(a), Terfenol-D rod is magnetized by the working current in the excitation coil, and it will occur as a magnetostrictive effect. The pre-tightening structure is as shown in figure\u00a03(b), the output force of the GMM structure is transferred to nuts by hinge-levers. Electromagnetic field of giant magnetostrictive structure can be analyzed by finite element method [12\u201315]. GMM is magnetized to generate magnetostriction, its operating principle is as shown in figure\u00a04. The hollow cylindrical GMM is coaxial with the screw, its radial thickness is 7 mm and axial length is 60 mm, the diameter of the screw is 40 mm, the air gap between the screw and the yoke is 7 mm. There\u2019s an aluminum sleeve between the screw and the GMM, its radial thickness is 3 mm", + " Generally, the axial contact rigidity of ball screw only involves the axial elastic deformation \u03b4 between the ball and the roller path, and does not involve the deformation of nut and screw. When the pre-tightening force Fp in the ball screw changes, the axial elastic deformation \u03b4 between the ball and the roller path will change. In this section, the experiments about measurement and adjustment of ball screw pre-tightening force are conducted, as well as the experiment of ball screw axial contact rigidity using the GMM structures. The solid cylindrical GMM structure is as shown in figure\u00a03(a), its parameters are as shown in table\u00a02. According to figure\u00a03(b), the structure\u2019s output force F1 is transferred to ball screw by hinge-levers, as shown in figure\u00a08. According to the principle of moment balance, the relation between acting force F1 of GMM structure and counter-acting force F2 of screw nuts is F2 = (l1/l2)F1,where l1 = 204.6 mm, l2 = 33.95 mm. The GMM structure\u2019s output force F1 is magnified 6 times by hinge-levers, that is, the pre-tightening force F2 will be magnified by the hinge-levers. The multi-field coupling relation in a giant magnetostrictive structure makes its dynamic behavior more complex [18, 19]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002575_125906-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002575_125906-Figure8-1.png", + "caption": "Figure 8. Force transferring relation.", + "texts": [ + " When the pre-tightening force Fp in the ball screw changes, the axial elastic deformation \u03b4 between the ball and the roller path will change. In this section, the experiments about measurement and adjustment of ball screw pre-tightening force are conducted, as well as the experiment of ball screw axial contact rigidity using the GMM structures. The solid cylindrical GMM structure is as shown in figure\u00a03(a), its parameters are as shown in table\u00a02. According to figure\u00a03(b), the structure\u2019s output force F1 is transferred to ball screw by hinge-levers, as shown in figure\u00a08. According to the principle of moment balance, the relation between acting force F1 of GMM structure and counter-acting force F2 of screw nuts is F2 = (l1/l2)F1,where l1 = 204.6 mm, l2 = 33.95 mm. The GMM structure\u2019s output force F1 is magnified 6 times by hinge-levers, that is, the pre-tightening force F2 will be magnified by the hinge-levers. The multi-field coupling relation in a giant magnetostrictive structure makes its dynamic behavior more complex [18, 19]. In order to verify its performance, we directly conduct electromechanical coupling measurement experiments" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002638_idt.2014.7038611-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002638_idt.2014.7038611-Figure8-1.png", + "caption": "Fig. 8. UWB Filter designed using Short-circuited Stub topology", + "texts": [], + "surrounding_texts": [ + "Figures 8 and 9, and Table V present a short-circuited stub topology and its performance as UWB filter. Observing its response we can conclude that this filter exhibits a wide bandwidth, a good flatness and matching as well as a relative ease in manufacturing. 12.2 mm 5. 4 m m via hole" + ] + }, + { + "image_filename": "designv6_24_0002184_embc.2015.7319462-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002184_embc.2015.7319462-Figure1-1.png", + "caption": "Fig. 1. The prototype of the robotic transtibial prosthesis.", + "texts": [ + " The prosthesis is designed to be light weighted and enable amputees to have larger movement range and better damping characteristics at the prosthetic ankle to adapt to terrain changes, instead of providing large power assistance for walking. A 50-W brushless DC motor (Maxon, EC 45-50W) equipped with a 5.8 : 1 reduction gearbox is used. The total weight of the prosthesis is 1.30 kg (without battery), and movement range is from 25-degree plantarflexion to 25-degree dorsiflexion. A uniaxial load cell (Interface, LBS), an angle sensor (AngtronRE-25) and two inertial measurement units (IMUs) are integrated with the prosthesis (see Fig. 1). B. Intrinsic Controller The intrinsic controller is based on damping control [9]. When the stator windings of the brushless motor are shorted, a braking torque is produced to prevent the motor from rotating. A pulse width modulation (PWM) signal is used to switch on/off of the motor-winding-short, and damping control is realized by controlling the duty cycle of the PWM signal. The damping output is determined by the ankle joint position and the measured ground reaction force (GRF) of the prosthetic side (see [7] for details)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002466_tcsi.2010.2046200-Figure9-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002466_tcsi.2010.2046200-Figure9-1.png", + "caption": "Fig. 9. Noise caused by 16-level (4-bit) quantization.", + "texts": [ + " 7 shows the frequency spectrum that is obtained if the input signal is sampled at the PWM carrier frequency . The spectrum now fills up with harmonics of the modulating signal. Apparently, sampling the input signal causes distortion. In a digital PWM modulator, the edges are synchronized to a high-frequency bit clock, e.g., 256 . Consequently, the pulsewidths are quantized to a limited number of discrete values. This can be modeled by inserting a sample-and-hold in series with the reference triangle operating at a multiple of the PWM frequency, as shown in Fig. 8. Fig. 9 shows the frequency spectrum that is obtained when 16 quantization levels are used (4-bit). As can be seen, the quantization in time results in noise. The SNR is improved by 6 dB for every additional quantization bit. However, in order to reach 100-dB SNR, at least 14-bit quantization is needed, which would require an unfeasibly high bit-clock frequency of 5.7 GHz kHz . Sampling and quantization effects can be separately or integrally handled. In a separated approach, as shown in Fig. 10(a), first, the distortion caused by sampling is corrected by either approximating NPWM using linear or higher order interpolation or applying precorrection based on a digital PWM distortion model" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003045_1.281-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003045_1.281-Figure4-1.png", + "caption": "FIGURE 4. A schematic depicting the flow patterns during valve closure forward motion, as seen in the central plane \u201ez\u00c40\u2026, the vertical plane \u201ey\u00c4\u00bf6 mm\u2026 on the minor orifice side and the vertical plane \u201ey\u00c4\u00c06 mm\u2026 on the major orifice side. Note that the actual flow map begins \u00c83 mm away from the actual valve seat.", + "texts": [ + " The valve then rebounds again, albeit with reduced amplitude, and closes again ;10 ms later. A third rebound was barely visible using the pressure and accelerometer curves. A schematic representation of the regurgitant flow patterns typically associated with valve closure forward motion and rebound in tilting-disk valves is provided as an aid to visualization in Figs. 4 and 5, respectively. In the central horizontal plane (z50 mm) there are two distinct flow regions: a minor orifice and a major orifice region. During the occluder\u2019s forward closure motion, as depicted in Fig. 4, the flow on the atrial side shows a striking minor orifice jet directed at an angle ~marked A!. Here the fluid is driven from the ventricle to the atrium by the global pressure gradient, against the local direction of motion of the occluder. On the major orifice side, the fluid ~associated with the closure volume! is traveling in the local direction of motion of the occluder and indicates a smoother parabola-like profile ~marked B!. In the transition between the minor and major orifice zones, a vortex ~;8 mm" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001572_tmag.1981.1061290-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001572_tmag.1981.1061290-Figure2-1.png", + "caption": "Fig. 2 Bow Coil Arrangement", + "texts": [ + " Their respective values however a r e funct ions of t h e c o i l a s p e c t r a t i o A equal to (R1 + R 2 ) / ( R 2 - R1) and the s t ruc tura l Suppor t arrangement. Since the cross-sectional area vailable t o support tension is l e s s a t R1 t h a n a t R2, methods of The author is w j t h Plasma Physics Laboratory, Princeton University, Princeton, NJ 08544, USA reducing f, E P,/P a re sought . For c i rcu lar co i l s having usefu l aspec t ra t ios th i s f rac t ion has a range 0 .6 ? f 7 0 .7 . For cons tan t t ens ion co i l s fp = 0.5. The s t ructural arrangement shown in Fig. 2 can P further reduce f P' A t ens ion l ink is a t t ached to each c o i l a t g with inc l ina t ion angle 0 and tangent to the co i l . The l i n k s a re a t tached to a cent ra l cy l inder which t ransmits a compressive force C directed towards the midplane. The compression C reduces the co i l t ens ion a t R, and, through t e l i n k s , increases the t ens ion a t R2. I\u20ac viewed i n a l i t e r a l ly g loba l s ense , t he equa to r i a l p ressure in the to ro id i s used t o produce compression from pole to pole ", + " 6 (next page) for toroidicities 0.5 < < 0.8. and radial halfthicknesses ch = 0.1, 0.2 and 0 . 3 . For the bow coil examples R is always 4 5 O . Note that the \u201cD\u201c coil and constant stress bow coil are identical for = 0.5. Representative values are also listed in Table I. STRUCTURAL ANALYSIS Two practical problems are now addressed ( 1 ) The effect of deformations (in both the coil and structural supports) on the internal force distribution, and ( 2 ) A realizable structure quivalent o the schematic arrangement shown in Fig. 2 The cylinder and pinned link configuration is difficult to construct due to the congestion where the links converge. The inverted arrangement shown in Fig. 7 is simpler. The coil structure is extended axially in the shape of a triangular blade ( B ) and a massive ring (A) restrains the radial motion. The blade then pushes towards the midplane at R1. A coil shape was geperated for = 0.7 ch = 0.2 and f = - .2. A structural model is shown in Fig. 8 . &e analysis, using $ structural code,6 indicates that distortions reduce the comprewive force at R1 to = - 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000559_amm.371.617-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000559_amm.371.617-Figure5-1.png", + "caption": "Fig. 5. Multi-tool processing system", + "texts": [ + " In the end the optimal solution is the one in which the objective function has the lowest value for the combination of vectors which led to choosing it. The results obtained point out that the most efficient constructive variants are the modular structures which certainly provide processing accuracy, high level of productivity and flexibility of the system. Integration of Multiaxis Heads into Processing Systems The most important and original contributions of the authors, presented in this paper and Fig. 5, refer to making more \u2265+ \u2264+ \u2208== \u2265++ \u2264++ \u2265++ \u2265++ \u2264++ ++++= \u2211 = ... ... ... (1,2,3)j 0;sau 1 ;1 ...min min662626161 max662626161 3 1 min1131312121111 max1131312121111 min1131312121111 min1131312121111 max1131312121111 6262131312121111 \u03b5 \u03b5\u03b5\u03b5 \u03b5\u03b5\u03b5\u03b5 xexe xx xx exexexe txtxtxt fxfxfxf pxpxpxp xxx xcxcxcxcC j ijij flexible the union flange 6 the multi-spindle end with the main spindle of the machine-tool, modularizing the supports 8 (19) and the frame-bars 17, 18 joining the gear box 1 and 2 with the supports-tool holder spindles 12 and 15, separating the gear box (1 and 2) as independent mode by multiplying the exit spindles to engage the cardan 9 and flexible 10 shafts, attaching some small special multi-spindle ends 14 on the supports-tool holder spindles 12, the additional stiffening elements 11 and 13, the modular structures and the flexible supports 16 joining to the guiding columns 20" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000888_ppc.2013.6627666-Figure12-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000888_ppc.2013.6627666-Figure12-1.png", + "caption": "Fig. 12. Comparison between 2D a) and 3D b) FEA induction plots for short circuit operation. For 3D results, coils on the right occulted to show the field magnitude on vertical and horizontal planes.", + "texts": [ + " The error on the rise time is low despite the fact that the most important errors related to the 2D assumption occur on the capacitance values. The rise time depends on both capacitance and inductance values and in the the absence of overshoot imposed by the specifications constraints, the errors on the inductance and capacitance values partially compensate each other in the computation of the rise time. Both 3D and 2D models are then in accordance when no pulse overshoot is allowed. The induction spatial distribution plot for short circuit operation (Fig. 12) and the electric field spatial distribution plot for the test V1 = 0 and V2 = 150kV (Fig. 13) with 2D FEA and 3D FEA can be compared. On the plane where 2D FEA is performed, one can notice that the results obtained by 2D and 3D FEA are in accordance. The differences between 2D and 3D FEA mainly occur in the windings corners where one can see that the magnetic field is more homogeneous than the electrostatic field. Consequently the error on the capacitances is bigger than the error on the leakage inductance" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001387_acdt.2018.8592940-Figure9-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001387_acdt.2018.8592940-Figure9-1.png", + "caption": "Fig. 9. Axial velocity of the flow at 0.3R after propeller disk of the actual model.", + "texts": [], + "surrounding_texts": [ + "The author would like to thank the aeronautical engineering research team at the Defence Technology Institute for their helpful discussions around the topic of the paper." + ] + }, + { + "image_filename": "designv6_24_0000782_ssp.198.301-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000782_ssp.198.301-Figure1-1.png", + "caption": "Fig. 1. Draft of the impact cylinder", + "texts": [ + " Because of the speed of the piston exceeding 10 [m/s] and the limited possibility of casing the measurement systems, contact (tactile) method cannot be applied. Previous research in the area of determining the kinetic energy of impact cylinders were conducted with the use of computer simulations [6] and non-contact experimental tests. Experimental tests were carried out with the use of the device developed for the determination and optimization of work parameters of pneumatic impact actuators. The method of laser triangulation was applied for the measurement of the piston shift [7]. The draft of the impact cylinder is presented in Fig. 1. The movement of the impact piston takes places when the product of P1 and S1 is higher than P2 and S2. After a slight shift of the piston, the area affected by pressure P1 significantly grows. This effect leads to an abrupt acceleration of the movement of the piston caused by the activity of compressed air accumulated in chamber 1. In order for the air flow obstruction to be avoided, the opening in chamber 2 needs to be of a big enough size, which in practice is realised with the use of a quick release valve" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001472_9781118354179.auto089-Figure10-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001472_9781118354179.auto089-Figure10-1.png", + "caption": "Figure 10. Twin clutch for a \u201ctwin clutch gearbox.\u201d (Reproduced with permission from LuK GmBH & Co. KG.)", + "texts": [ + "auto089 Also published in the Encyclopedia of Automotive Engineering (print edition) ISBN: 978-0-470-97402-5 a multi-plate diaphragm spring clutch design (using two diaphragm springs) with four friction discs. In addition to controlling the torque transfer for propelling the vehicle, some multi-plate clutch designs (with two or more friction discs) have two concentric output shafts, with one or more friction discs splined on each shaft. In the case of \u201ctwin clutch gearboxes,\u201d both shafts are entering the gearbox, as shown in Figure 10 (LuK). The two clutches are controlled independently and during normal driving one clutch (can be either of the two clutches) is fully engaged and the other completely disengaged. The disengaged clutch is interrupting power flow, enabling preselection of another gear ratio (pair), most likely to be used in the next \u201cgear change.\u201d The \u201cgear change\u201d is performed by controlled, simultaneous Encyclopedia of Automotive Engineering, Online \u00a9 2014 John Wiley & Sons, Ltd. This article is \u00a9 2014 John Wiley & Sons, Ltd", + " This power transfer without interruption has advantages in terms of vehicle acceleration times and driving comfort. However, slippage of both clutches uses energy and generates heat. Depending on the driving maneuvers and frequency of gear change, considerable heat can be generated, which must be dissipated in order to maintain required friction torque characteristics (coefficient of friction) and prevent excessive wear. As a result, dry twin clutch designs for twin clutch gearboxes (as shown in Figure 10) are only suitable for smaller vehicles and engines, typically under 350 Nm maximum engine torque. Clutches for engines with higher torque require wet clutches to satisfy increased heat dissipation requirements and ensure acceptable life of friction pair components. In agricultural and \u201cimplement carrier\u201d vehicles clutches have an additional function in separating the power flow from the engine flywheel toward two outputs shafts. Indicated by arrows in Figure 11 (Janic\u0301ijevic\u0301, Jankovic\u0301, Encyclopedia of Automotive Engineering, Online \u00a9 2014 John Wiley & Sons, Ltd" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003669_2014-01-2400-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003669_2014-01-2400-Figure1-1.png", + "caption": "Figure 1. Drive train layout of the \u201cGreen Wheel Loader\u201d", + "texts": [ + " Previous research activities focused on the substitution of individual subsystems of the machines. A consequent implementation of new drive technology has not been done so far. In the \u201cGreen Wheel Loader\u201d, the aforementioned technologies will be combined in one drive train for the first time. Besides efficiency advantages compared to conventional solutions, the subsystems' regenerative capabilities and the possibility to improve the machine system's adjustment are expected to result in fuel savings. The \u201cGreen Wheel Loader's\u201d drive train layout is shown in figure 1. It is based on the chassis of a 24-ton Liebherr L576 wheel loader [14]. A diesel engine optimised for low-speed operation powers the system. The engine based on a Deutz prototype machine TCD 7.8 L6 reaches its nominal output power of 200 kW already at 1200 rpm and is limited to a maximum rotational speed of 1600 rpm. The engine in combination with its exhaust after-treatment system (DPF and SCR) fulfils the exhaust emission limits of EU Stage 4 and US TIER 4 final. The hydrostatic-mechanical power-split transmission HVT used for the travel drive is supplied by Bosch Rexroth [15]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003210_icectech.2011.5941668-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003210_icectech.2011.5941668-Figure8-1.png", + "caption": "Figure 8. membership functions (a) inputs (b) outputs", + "texts": [ + " Reference current can be obtained by solving non-linear equation of flux and torque which are independent of Rs [11]: 2*2** 2*2** **** )()( )()( ])([ 4 3 qd qqfdd qddqqf p iiI iLiL iiLLi n T += ++= \u2212\u2212= \u03bb\u03bb \u03bb (14) The proposed Fuzzy Logic Identifier (FLI) has two inputs that are stator flux and current errors. The output of the Fuzzy estimator is the required change of resistance, R. The value of R is continuously added to the previous estimated stator resistance Rs[k-1]. The final estimated Rs can be used directly in the controller. The membership functions of the normalized input/output variables are given in Fig. 8. With two inputs and three membership values, the estimator contains nine rules listed in Table I. V. SIMULATION AND RESULTS The proposed control scheme was simulated with a MATLAB/SIMULINK model based on Fig. 5. Fig. 9 shows the simulated results for the PMSM torque without SRE. Motor operates at a low speed. The stator resistance changed to 1.3 times its actual value as a step function in 0.5 second. It can be seen that Rs deviations lead to torque error in the steady state. This error is due to the error in estimated stator flux" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003315_detc2005-84562-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003315_detc2005-84562-Figure5-1.png", + "caption": "Figure 5. Definition of Gear Tooth Crowning in Term of Cr", + "texts": [ + " Equation 19 entails to result in a smooth transition from unmodified involute to modified involute. The second method is tooth crowning. This type of tooth 5 Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 02/01/2016 Te modification is similar to tooth shaving but the amount of tip relief and length of modification at any cross section of gear tooth are varied along the tooth face direction. Therefore, if viewing the tooth face from the side, one can see a curve on the tooth face. If this curve appears to be parabolic, we call it as a parabolic crowning. Cr in figure 5 is a parameter affecting the gear tooth crowning and the equation of projected profile along the tooth face is defined by the following expression. y(z) = 4(Cr \u2212H)z2 L2 +(Radd \u2212Cr) (20) Where Cr is crowning parameter, H is length of profile modification, Radd is addendum radius, and L is gear face width, as shown in figure 5. Gear tooth profile optimization To our knowledge, the method of selecting the proper amount of tooth crowning and tooth shaving for any gear pair is still unavailable. Thus, we propose the systematic approach based on three dimensional rigid body dynamics, the optimization technique and the PISE method to predict these values. After reviewing all of optimization techniques, we choose the Box\u2019s rms of Use: http://www.asme.org/about-asme/terms-of-use Downl complex method [6] which is efficient and does not require an explicit objective function" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003929_rast.2013.6581342-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003929_rast.2013.6581342-Figure6-1.png", + "caption": "Fig 6. Subsystem Integration for Satellite Quetzal V.I", + "texts": [ + " The satellite bus consists of the following subsystems, which will be developed under requirements and constraints of the selected payload: navigation, attitude determination and control subsystem (NADCS), power supply subsystem (PSS), structural subsystem (SS), thermal control subsystems (TCS), telecommunications subsystem (TS), data handling and processing subsystem (DHPS), telemetry subsystem (TMS) and propulsion subsystem (PS). For the correct integration stage, it has been considered the interface subsystem. Also will be performed the reliability and electromagnetic compatibility analysis, and as part of our recent research field, the sustainability analysis. Preliminary work on the Quetzal satellite is shown in the following figures (Fig. 5, Fig. 6 and Fig. 7). The first proposals are currently used for mechanical, thermal and vibration analysis, which will be verified later on mock ups, while other systems are developed in a engineering model stage with COTS and low cost elements by the student groups at UNAM. Once the electronic design and software has the first round of test, we will evaluate if groups from other institutions will be invited for the integration of the full engineering model, and definition would be made for the flight model based on reliability, cost and execution" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001146_rfit.2007.4443911-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001146_rfit.2007.4443911-Figure4-1.png", + "caption": "Fig. 4. HFSS model of 2.4 GHz heatsink antenna", + "texts": [], + "surrounding_texts": [ + "To demonstrate the concept, heatsink antennas were designed using microstrip patch antennas on FR4 PCB. A square patch antenna was used as the mounting surface. The heatsinks were sized to fit within the planar area of the patch and were attached using silver, conductive epoxy. The planar layout of the patch and matching network are shown in Fig. 2. The patch antennas do not use section 3. This stub is for matching 1-4244-1307-9/07/$25.00 \u00a92007 IEEE with the heatsink added only. Section 1 is a quarterwave transformer and section 2 is a 50-ohm transmission line forming an edge-fed design. An inset feed cannot be used here since the heatsink base would effectively remove the inset. A coaxial feed, however, from the ground layer beneath can be used for potential 3-D integration. Each of the antennas was fabricated on a lossy, lowcost FR4 substrate with a thickness of 1.6 mm and a dielectric constant of 4.2. Ansoft HFSS was used for EM simulation since the heatsink antennas are 3-D structures. III. 2.4 GHZ PIN-FIN HEATSINK ANTENNA Based on the design and layout of patch antenna in Fig. 2, a 2.4 GHz pin-fin heatsink antenna was fabricated. The heatsink has 30 irregularly shaped vertical pin fins each having a height of 21.5 mm. The base of the heatsink was cut to match the planar dimensions of the square patch (29.1 mm x 29.1 mm). The thickness of the base is 3.5 mm. The length of the tuning stub (section 3 in Fig. 2) is 14.0 mm with a separation of 6.6 mm from the patch. Figs. 3 and 4 show the fabricated 2.4 GHz antenna and the HFSS model, respectively. Using FR4 PCB is a low-cost option for the fabrication of patch antennas. However, it has a relatively high loss tangent (tan( ) = 0.02) which reduces the radiation efficiency of patch antennas printed on this material. The simulated radiation efficiency for a 2.4 GHz patch antenna using an edgefeed with a quarter-wave transformer and a fifty-ohm trace also of a quarter-wavelength is 33%. Simulations show that the heatsink significantly improves the efficiency (increasing it to 79%). Table I summarizes other antenna parameters comparing the patch and heatsink antennas. The bandwidth more than doubles from 2.6% to 6.0% due to the heatsink. The peak directivity of the heatsink antenna decreases due to added end-fire radiation in the E-plane. The peak gain, however, increases due to the effect of the heatsink on the radiation efficiency [1]-[2]. IV. 5.8 GHZ EXTRUDED-FIN HEATSINK ANTENNAS In addition to the 2.4 GHz pin-fin heatsink antenna, extruded-fin heatsink antennas operating at 5.8 GHz were designed and fabricated. Unlike a symmetrical pinfin heatsink, the extruded-fin heatsink can have two orthogonal orientations: 1) fins parallel to the nonradiating edges (fins PNRE) of the patch and 2) fins parallel to the radiating edges (fins PRE). This orientation must be considered. Two heatsink antennas (one of each orientation) were fabricated and simulated. These antennas use the same layout in Fig. 2. The length and width of the square patch are both ~11.8 mm. The heatsinks have 3 fins each having a height of 12.5 mm (not including the base thickness which is ~2.4 mm). The fabricated heatsink antenna with fins parallel to the non-radiating edges is shown Fig. 5 while the HFSS model of the opposite orientation is shown in Fig. 6. The measured return loss for the patch antenna and the two orthogonally-oriented heatsink antennas is shown in Fig. 7. The heatsink altered the input impedance of the basic patch, so a tuning stub was used to match the heatsink antennas. The figure shows considerable bandwidth improvement over that of the patch antenna. Table II compares the bandwidth and other parameters of the three antennas. As shown in Table II the peak directivity for the two different heatsink orientations is quite different. With fins parallel to the non-radiating edges, the directivity is increases compared with the directivity of the patch. Measurements of the radiation pattern are similar to the patch antennas pattern with a focused broadside lobe. For the heatsink with fins parallel to the radiating edges, however, the directivity decreases compared with the directivity of the patch. In this case, end-fire radiation occurs in the E-plane, and as a result, the directivity is reduced. In this manner, the orientation of the heatsink can have a significant impact on the antenna performance. Also shown in Table II is the bandwidth improvement that comes with the heatsink antenna. The 10-dB bandwidth is increased nearly six-fold in the case with fins parallel to the non-radiating edges. For the case with fins parallel to the radiating edges, the bandwidth is nearly quadruple the bandwidth of the patch. V. INTEGRATED PA AND HEATSINK ANTENNA MODULE To demonstrate the concept of the heatsink antenna, a commercial power amplifier from RFMD was integrated with a heatsink antenna. The heatsink acts as the antenna while simultaneously being used for heat transfer. The IC is an RFMD 2126 high-power linear amplifier with a maximum output power of 1.3 Watts and a gain of 12 dB. The PA circuit was setup to operate at 2.45 GHz. The heatsink has a base dimension of 19.8 mm on each side with a base thickness of 2mm. The heatsink has 11 fins each with a height of 8 mm not including the base. The package has a slug on the backside which is used as a ground and a thermal sink. This slug was connected to the heatsink on the opposite side of the FR4 board with an electrically-insulating thermal via as shown in Fig. 8. A separate electrical via was used to feed the antenna with the RF signal. The placement of the heatsink with respect to the fixed feed determined the impedance matching which is similar to determining the feed point for matching a patch antenna. The return loss of the antenna is shown in Fig. 9 and the antenna displays multi-band behavior. Not only is the antenna multi-band, but the bandwidth at each resonant frequency is increased compared with the bandwidth of a basic patch antenna. The peak gain was measured using the absolute method [3]. A gain of 12.6 dBi was achieved with the heatsink antenna not including the gain of the power amplifier. Temperature measurements showed that the temperature of heatsink increased over ambient and confirmed that it functions properly as a heatsink. Heat transfer improvements can be made by using a material with a higher thermal conductivity for the thermal via." + ] + }, + { + "image_filename": "designv6_24_0000255_0094-5765(80)90086-7-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000255_0094-5765(80)90086-7-Figure4-1.png", + "caption": "Fig. 4. ERDSAT in orbit configuration (Version II).", + "texts": [], + "surrounding_texts": [ + "Another payload element is a Data Collection System (DCS), (CNES, 1977), equivalent to the French ARGOS-system. This system applies the random access technique and receives the in situ measurement data from ground based Data Collection Platforms (DCPs) in the UHF-band. Certain reference data can be used for MOMS and SAR data processing. The onboard equipment of the DCS includes the UHF antenna and receiver as well as control and processing units.\nSatellite concept The major requirements determining the ERDSAT concept are: - -ARIANE launcher compatibility (e.g. 3 m fairing) --Mission requirements as given before --Three years lifetime (preoperational) spacecraft platform", + "---Growth potential to a multimission ARIANE platform - - \" L o w cost\" approach.\nThe selected configuration (Figs. 3 and 4) features: --modularity by separate modules for payload and bus compartments --thermostable structure for the payload platform --ease of integration ---extensive use of space proven hardware elements ---design experience from German ZKS programme.\nTwo versions were analysed with differences in SAR antenna mounting and, consequently, arrangement of optical sensors as well as solar generator geometry. Both allow slewing of the SAR antenna minimum depression angle form 45 \u00b0 to 71 \u00b0. Figure 5 shows Version II.\nThe bus allocated mass is 880 kg including payload telemetry, attitude and orbit control system and power conditioning, whereas the payload mass budgeted is 450kg. Therefore a margin of more than 1200kg remains, with respect to the ARIANE payload capability of 2600 kg (see Fig. 6).\nThe maximum power (see Fig. 7) needed is 1750 W (peak), which will be provided by a 1330W (BOL) solar generator and batteries. Half of the power required is necessary for battery loading as a 6 rain, 1000 W SAR operation was assumed even during eclipse.", + "i\nz ~ i l l i i w z Z < w o w w\n~ z ~\nI\ni\n1 !\n- - , , - i \" - ~ , s . J j\ni r , / ~ i\nEw\nw ~\ni\nI\noto0o \u00ae \\ i ii 9 +w olo\u00ae . . . . ~ i t~ ' t :!ii\n' , ' ' 4 \\ / i ' / I i I ' L - - ~ = / /' _ _ - . . . . . ~\n0\n0\n<" + ] + }, + { + "image_filename": "designv6_24_0003937_978-1-4471-5104-3_5-Figure5.11-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003937_978-1-4471-5104-3_5-Figure5.11-1.png", + "caption": "Fig. 5.11 Equivalent circuits of the d-q axis: a direct axis; b quadrature axis", + "texts": [ + " The abc to dq transformation matrix for the balanced system is: Tdq0 \u00bc 2 3 cos he\u00f0 \u00de cos he 2 3 cos he \u00fe 2 3 sin he\u00f0 \u00de sin he 2 3 sin he \u00fe 2 3 1 2 1 2 1 2 2 4 3 5 \u00f05:7\u00de The voltage equation for the synchronous coordinates is: vd vq \u00bc Rs 0 0 Rs id iq \u00fe xe 0 1 1 0 wd wq \u00fe d dt wd wq \u00f05:8\u00de where wd wq \" # \u00bc Ld 0 0 Lq id iq \u00fe wpm 0 Ld \u00bc 3 2 L0 Lm\u00f0 \u00de Lq \u00bc 3 2 L0 \u00fe Lm\u00f0 \u00de The state equation is given by: d dt id iq \u00bc Rs Ld xeLq Ld xeLd Lq Rs Lq \" # id iq \u00fe 1 Ld 0 0 1 Lq \" # vd vq \u00fe xe 0 wpm Lq \u00f05:9\u00de From Eqs. 5.8 and 5.9, it can be built the equivalent circuit in synchronous coordinates like the circuits shown in Fig. 5.11. Digital technology, multi-phase converters, and position sensors are well suited for synchronizing the switching current with the rotor position. These sensors are often optical or Hall Effect devices using the rotor magnetism to sense its position. It is expected that the torque reduces as the speed increases since the rotor generates proportionally a counter e.m.f. (electromotive force) across the coil that is coming closer to the pole. There will be a current reduction and so, a reduction of the magneto-motive force (m" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003048_iceaa.2011.6046498-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003048_iceaa.2011.6046498-Figure1-1.png", + "caption": "Figure 1: Geometrical sketch of a coaxial cylindrical cavity resonator which has been filled by an ordinary DPS layer (region 1) and a metamaterial layer (region 2).", + "texts": [ + " In the last section, through the use of this novel resonator a novel miniaturized TM010 mode cylindrical cavity filter with the diameter and total length of just 0.232\u03bb and 0.123\u03bb respectively, is designed and simulated. The analysis begins by solving for the electromagnetic fields inside a coaxial circular cavity loaded by a concentric DPS layer with properties \u03b51, \u03bc1 (region 1) and an annular anisotropic MNG layer (region 2) with the permittivity equal to \u03bc2 and the permeability tensor equal to 2 0 0 2 \u02c6 \u02c6 \u02c6 \u02c6 0 \u02c6\u02c6zz , as seen in Figure1. The height and the radius of this structure are H and R respectively and the thickness of region 1, region 2, and the radius of inner coaxial part are R1, R2, and R3 respectively. It can be shown that TM mode without axial and angular variation of the fields (TM010) is the lower order mode in this structure provided that H \u2264 2R. 978-1-61284-978-2/11/$26.00 \u00a92011 IEEE Assuming H \u2264 2R we will continue the analysis considering only TM modes. Extracting the general form of Maxwell equations in a cylindrical anisotropic medium, combining these equations to formulate the scalar wave equation in such a medium and finally deriving the solution of such equations and matching the boundary conditions for the structure shown in Figure 1 provides the dispersion relation for different modes. These relations for the azimuthally symmetric TM mode can be expressed using equation (1): 2 20 2 0 2 0 22 1 1 2 0 2 0 1 1 0 1 0 1 1 0 1 ( ) ( ) . ( ) ( ) ( ) ( ) ( ) ( ) I K I K J Y J Y k k d k d k k d k d k d k d k d k d (1) where J0(.), Y0(.), I0(.), and K0(.) are the Bessel functions and the modified Bessel functions of the first and second kinds [6], and 2 2 2 2 1 1 1 2 2 2 2 3 0 1 3 1 0 1 3 0 2 2 0 2 , , J Y I K k k c c d R R k R k R k R k R (2) Since we are interested here in squeezing the resonant dimensions of the cavity we assume that \u03c9, R1, R2 and R3 are chosen such that small argument approximation can be used for the Bessel functions, and modified Bessel functions and their derivatives [6]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001573_j.apm.2014.11.058-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001573_j.apm.2014.11.058-Figure7-1.png", + "caption": "Fig. 7. Cross-section and geometric parameters of a coaxial magnetic gear mechanism.", + "texts": [ + " Therefore, the magnetic flux densities within the inner air gap and outer air gap can be calculated by finding the magnetic fluxes of all nodes at layer 5 and layer 3 divided by the areas of inner and outer air gaps, respectively. For numerical calculations, geometric parameters and the magnet\u2019s material properties of a coaxial magnetic gear mechanism, given in Table 1, are taken as an example to calculate the air gap flux densities. NdFeB magnets are used to increase the magnetic flux densities within the air gaps and the transmitted torque of the mechanism. The cross-section and geometric parameters of this coaxial magnetic gear mechanism are illustrated in Fig. 7. Fig. 8(a) and (b), respectively, present the cite this article in press as: Y.-C. Wu, B.-S. Jian, Magnetic field analysis of a coaxial magnetic gear mechanism by two-dimensional lent magnetic circuit network method and finite-element method, Appl. Math. Modell. (2014), http://dx.doi.org/10.1016/ 2014.11.058 waveforms for radial components of magnetic flux densities within the inner air gap and the outer air gap of the coaxial magnetic gear mechanism, when N is set as 3000, i.e. the coaxial magnetic gear mechanism is divided into 3000 parts in the circumferential direction" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001153_amr.189-193.1882-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001153_amr.189-193.1882-Figure3-1.png", + "caption": "Figure 3", + "texts": [ + "72, Rutgers University Libraries, New Brunswick, USA-03/06/15,16:14:29) Similar to the Step 1 in the establishment of the wheel position,click the icon to draw the wheel profile.Next click the icon to union the front 2 wheels as one part.Rename is as \u201cQianchelun\u201c.Similarly build the rear wheels.In order to see the wheels are rolling on the ground during simulation, some holes should be createed on wheels by using boolean subtract operator.Click the icon to build axle and connecting part.Rename them as \u201cQianzhoulian \u201c and \u201cHouzhoulian \u201c.The createed wheel model is shown in figure 3-(a). Click the icon in XOY view to build the long pole.Set \u201cradius=40\u201d, \u201cstart point\u201d as (850, 100, 0) and \u201cend point\u201d as (2800,600,0).Rename it as \u201cChanggan\u201d.Set point (850,100,0)as center to build slider1 and rename it as \u201cHuakuai1\u201d. Click the icon to build the short pole.The same set as the long pole, the start point is (1150,600,0) .The end point is (3500,100,0).Rename is as \u201cDuangan\u201d.Create the slider2 similarly and rename is as \u201cHukuai2\u201d.The model is shown in figure 3-(b). Click the icon and set \u201cLength=135 Radius=60\u201d.Draw a cylinder along the vertical direction and revolve it as 72 degree around Z axis.Rename is as \u201cJushengyougang\u201d.Similarly create the oil pole.Set \u201cRadius=40\u201d.Rename is as \u201cJushengyougan\u201d. \u201cQingxieyougang\u201d and \u201cQingxieyougan\u201d are set up by the same way.The model is shown in figure 4. Click the icon\u201c box\u201dto create the ground and cargo heap.Rename them as \u201cDimian\u201d, \u201cHuodui\u201d. Now the createed model of high-dump truck is shown in figure 4. Click the icon to Fix the Dimian and the ground,Huodui and the ground" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002247_j.jmatprotec.2004.07.117-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002247_j.jmatprotec.2004.07.117-Figure3-1.png", + "caption": "Fig. 3. Cross-section of the impeller wheel welded by classical method: (1) weld; (2) blade and hub disc; (3) cover disc.", + "texts": [], + "surrounding_texts": [ + "There are a few types of constructions of the shrouded impellers. The right selection of it benefits with the better efficiency of the particular stage and whole compressor and also operational properties. In practice there are following constructions: (a) strictly radial, with two-dimensional blades \u2013 the most principal construction, (b) two types of impellers with complex construction of blades: \u2022 with diagonal inlet and higher efficiency than two- dimensional, \u2022 with axial\u2013radial three-dimensional impellers (the highest efficiency). The field of application is big enough to use both types of impellers, depending on various criteria. If the shrouded impeller is made from forging, the cover is usually manu- \u2217 Tel.: +48 91 449 47 51; fax: +48 91 449 43 56. E-mail address: jnowacki@ps.pl. factured as a separate part (except cases, when channels are being manufactured using electro-erosion process or 5-axis milling). 924-0136/$ \u2013 see front matter \u00a9 2004 Elsevier B.V. All rights reserved. oi:10.1016/j.jmatprotec.2004.07.117 The blades can be also manufactured separately, but more often they are milled from the hub disc. Progress of centrifugal compressors, and especially of the application of a shrouded impellers with narrow flow passages, revealed the necessity of developing new methods of joining the impeller parts: the hub disk, blades and matching cover. These methods can be classified into three main groups: welding, brazing, and riveting. Application of the joining method is strictly dependent on blade geometry (two-dimensional or threedimensional), impeller dimensions (outer diameter and channel width) and operating conditions (mechanical stresses, kind of the flowing medium and temperature). At present the most often used method of joining wheel elements is welding [1]. In dependences to geometry of the wheel, and mechanical stresses various methods of welding can be used (Figs. 2\u20135). Area of welding application covers all types of impellers, which tangential velocities not exceed 340 m/s \u2013 practical limit of the closed wheel application. Welding does not find the application in construction of impeller with narrow flow passages, as to technological restrictions. In those cases for joining of wheel elements vacuum brazing is used [2\u20133]. At present the most often used method of joining wheel elements is welding. rocess for centrifugal compressor impeller. Welding is the most often technique applied for joining the impeller parts. Depending on the geometry of the flow channel and mechanical stresses there are following types of that technology (Figs. 3\u20135): \u2022 welding in the channel, \u2022 slot welding, \u2022 electron-beam welding, \u2022 laser welding. The range of application of welding method includes all the impeller types and the entire flowing medium. There is also no limitation from the point of view of circumferential speed which can reach 320\u2013340 m/s (such speed is limited rather by material mechanical properties). Welding is not used for production of shrouded impellers with narrow flow passages [4\u20135]." + ] + }, + { + "image_filename": "designv6_24_0000561_cp.2010.0023-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000561_cp.2010.0023-Figure3-1.png", + "caption": "Figure 3: Benchmark industrial motor FE model.", + "texts": [], + "surrounding_texts": [ + "This paper explores the use of a hybrid 'canned' induction rotor design in which the conventional cage of an industrial motor is replaced by a simple cage with a low radial depth, along with a series of air-gap cans to provide environmental shielding, and a conducting rotor can to simplify the rotor design. Figures 3 and 4 show a cross-sectional view of the 2D finite element (FE) models of a conventional (benchmark) industrial motor and the proposed simplified rim-driven canned motor, respectively. In a conventional industrial motor, the rotor bar is designed to use the 'deep-bar' effect to increase starting torque. The simplified bar design of the rim-driven motor therefore, would result in a significant loss of starting torque. This is partially circumvented, however, by including a simple conducting can directly underneath a containment (environmental) can on the rotor outer surface, as illustrated in Figure 5 . Canned rotor construction has been used successfully in the past in other applications [5]. The conducting rotor can (and also containment can) here acts as a secondary high-resistance cage whose main function is to boost the starting torque of the motor. The unusual combination of a simple rotor bar design and the conducting rotor cans allows both the radial depth and weight of the rotor to be reduced without significant loss of performance. The objective was to design a rim-driven motor with the following specification requirements: a rotational speed of approximately 575 rpm, 440 V 3-phase supply, 60 Hz and a torque of 3,000 Nm. The motor would be located on the outer rim of a nickel-aluminium-bronze (NAB) tube around the propeller blade tips. The outer diameter of the propeller was 0.8 m including a 20 mm radial thick NAB tube/ring for structural support. The inner diameter therefore, of the rim driven rotor was 0.84 m. A conventional industrial induction machine design was modelled using the Flux2D FE analysis software package. This was used as a benchmark to compare the performance between an industrial motor and the subsequent rim-driven motor design. The industrial motor specification was chosen to have approximately the same torque rating at full-load as the rim-driven motor. This was based on the reasonable assumption that the torque largely dictates the size and weight of a motor. One of the key differences between the benchmark industrial motor and the rim-driven design is the air-gap. The philosophy in a conventional induction motor design is to minimise the air-gap, subject to mechanical and thermal constraints. The benchmark industrial design had an air-gap of 1.5 mm. The rim-driven design on the other hand has to accommodate several non-magnetic cans within the air-gap and due to the larger diameter, mechanical requirements dictates a minimum air-gap of 3.5 mm. The rim-driven design therefore, has a relatively large air-gap. Firstly, there is a can on the inner stator surface; this is made of stainless-steel and is utilised for environmental protection. Secondly, there are two cans on the rotor outer surface: one conducting, which is copper in this instance, and one environmental, again stainless-steel. The stainless-steel environmental cans are 0.5 mm thick and the copper conducting can is 1 mm thick, producing a total air-gap length of 5 .5 mm. This has a significant impact on the design of the rim-driven motor and results in a large magnetising current. The environmental cans were chosen to be as thin as possible to essentially minimise the air-gap length and therefore keep the magnetising current down whilst still maintaining their structural integrity. The radial thickness of the copper can was chosen based upon its effects on the motor performance. The large diameter of the rim-driven design means that the axial length can be reduced significantly in comparison to a conventional machine. It is important, therefore, to minimise the overhang length of the end-winding basket to prevent the end-winding impedance having a strong detrimental effect on the performance of the motor. This would also help to keep the weight and volume to a minimum. A conventional 12-pole winding is used to meet the low-speed requirement of the rim driven design. The high pole number has the advantage of reducing the overhang length of the end-winding basket. A double-layer winding is utilised to further reduce the end winding basket. This would also increase the efficiency of the motor by reducing some of the low order mmf winding harmonics, as illustrated in Figure 6. This is additionally beneficial in reducing the eddy-current losses in the stator containment can; these can be significant because the can is stationary with respect to the rotating air-gap fields. Efficiency, however, was not the main driver in the design of this motor due to its low duty-cycle operation. In conventional induction motor design, the air-gap emf induced per phase in the stator winding is assumed to be approximately equal to the terminal phase voltage, which is then used to set the stator winding turns for a desired air-gap flux density. The large magnetising current of the rim-driven design, however, resulted in a lower specific magnetic loading using this approach. This was compensated for by increasing the required number of turns to bring the specific magnetic loading back up to the desired level of approximately 0.5 T, which produced a peak air-gap flux density of approximately 0.8 T, as shown in Figure 7. Furthermore, the electric loading of the machine was increased because of the availability of water cooling. This motor is to be operated in a sea-water environment. Consequently, the motor will be submersed during operation and, as a result, sea-water will be present in the air-gap in direct contact with the environmental cans on the stator and rotor. Furthermore, the increased surface areas of the rotor and stator cores means that substantially improved cooling of the machine is available when compared to a conventional air cooled machine. The stator conductors therefore, were designed to operate at up to 8 Almm2\u2022 This was almost three times more than the full-load current density of 2.89 Almm2 in the benchmark design. The rotor bar load current is related directly to the stator estimated load current. The rotor bars, however, are normally worked at a higher current density than the stator conductors. The rotor bar material was also changed from aluminium, used in the benchmark motor, to copper to further reduced the required bar area. The rotor bars therefore, were designed to operate at approximately 16 Almm2\u2022 Machines are usually designed to be worked just below saturation level to make effective use of the lamination steel in the core-backs. The stator and rotor core-back of the rim driven motor were designed therefore, to operate at a magnetic flux density of approximately 1.45 T. However, the approximation is less reasonable in the case of the conducting cans because in reality the currents are not constrained to flow axially in the solid material. As such, it cannot be identified as a circuit and hence, it is not possible to add an effective external end-ring impedance. As shown in Figure 8, the air-gap field produces circular current paths in the conducting can, which are not restricted and can flow in any direction. Currents in a caged rotor on the other hand are constrained to flow axially down the rotor bars and around the end rings. .... ..0:- \ufffd+-+-\ufffd\ufffd+- \ufffd ..0:- \ufffd - - ...... -4 \ufffd----=..\ufffd\ufffd----\"'----=..\ufffd _ _ _ I ,, <\"'\"\ufffd\ufffd......-+-\ufffd+-\ufffd \"' ..... \" .---\ufffd----=..-+\ufffd\ufffd\ufffd ..... \\ I / ,r<,\ufffd---\"--\" , \\ , I / /' /\ufffd\"\"'-'-410-.\",..-..,.\"\"\">.\" , \\ I I ///_ (g _L1)\\ 2 -(g +L1) 2 ii =0 (6) Here it is expected that the converter considering (2) controls the current of one pole somehow that torque is set at desired level and by controlling the other pole current such that the currents ratio satisfies (6), UMF become minimized", + " The main advantage of this control algorithm is that the role of S2 and S3 are interchangeable, in other words S2 can be in charge of controlling UMF and S3 minimizes torque ripple. Consequently it is illustrated that there is no need to identify fault direction and always currents are controlled in a way that yield the expected tasks. V. SIMULATIONS A workbench of the test for a 6/4 SRM has been established utilizing FEM. The machine dimension is as Table I. In this study, the stator and rotor cores are made up of non-oriented silicon steel laminations and, each phase winding consists of 120 turns with a nominal current of 2. 5 A. As shown in Fig. 2 in the considered faulty condition the rotor is displaced from its centric position along phase A in a way that air gap for three coils, namely CAl, CSI and Ce2 are increased and the other coils, namely CAb CB2 and CCi are decreased. Two eccentricity degrees, namely 30% and 70% are considered here. Table I: Motor Characteristics Parameter Value Stator core outer diameter 72mm Rotor core outer diameter 40.5 mm Length of air gap 0.25 mm Stack length 35mm Shaft diameter 10mm Rotor pole arc 32\u00b0 Stator pole arc 28\u00b0 Number of turns 120 A" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002010_jae-2008-999-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002010_jae-2008-999-Figure8-1.png", + "caption": "Fig. 8. Calculated result of a 6th-mode at the resonance frequency 246,200[Hz] with 14.14[Vrms] input voltage.", + "texts": [ + " The simulation result was obtained by 3-D FEM and Eq. (79), and experimental data by the impedance analyzer. The number of the element for the FEM analysis was 29,568. The tetrahedral mesh with 4-node is used for an element. As shown in Fig. 7, the data from the simulation was almost in agreement with the experimental data. This agreement proved that 3-D FEM routine used in this research was correct. The calculated resonance frequency value, which generated the desired mode properly, was 246,200[Hz] and the experimental result was 248,125[Hz]. Figure 8 shows the calculated result of the 6 th-mode wave at the calculated resonance frequency 246,200[Hz] with 14.14[Vrms] input voltage. Briefly, the vibrator of an 8.5[mm] outer diameter RUSM was analyzed and designed considering the resonance frequency, the mode, the size, and etc., using 3D-FEM at this design stage. In this research, four kinds of teeth were made for the vibrator of the USM, varying the number of teeth such as 12, 16, 18, and 22. This was because to find out the relation between the number of teeth and the speed of the motor and also to verify the Cutting Method (CM)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003121_12.826223-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003121_12.826223-Figure1-1.png", + "caption": "Figure 1: (a) NIRCam Strut Assembly (b) Strut Assembly Interfacing to NIRCam and to ISIM", + "texts": [ + " The structure must be able to survive launch loads of an Ariane V rocket and then accommodate a differential thermal contraction of the optical instrument as the temperature drops from ambient at 293K to operation at 35K. Rigid body motion of the instrument can lead to severe pointing errors and thermal distortion can create detrimental wavefront errors (WFE) that spoil the system\u2019s optical performance. To achieve optimum performance, one must exert care in the design of the space structure. The NIRCam instrument is mounted to the NASA ISIM structure (see figure 1b) via a mechanical support structure, also known as a strut assembly (see figure 1a). The strut assembly is designed to support and position the instrument Optical Assembly (OA) in all expected environments, including test, transport, launch, and on orbit operations. The strut assembly attaches to the OA at the OA-Strut Brackets. The strut pads that interface to the OA-Strut Brackets are the Strut-OA Pads. The struts pads that attach to the ISIM at the SI Interface Plates (SIIP) are the Strut-SIIP Pads. *paul.v.mammini@lmco.com; phone 1 650 424 2881 Astronomical and Space Optical Systems, edited by Penny G" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000197_elk-1306-206-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000197_elk-1306-206-Figure3-1.png", + "caption": "Figure 3. a) Solid model of the flywheel, b) prototype.", + "texts": [ + " The driving method of these machines is simpler and the high torque per ampere feature makes them the best candidate for this application. In the application a BLDC motor of Maxon (EC25) has been chosen. Its technical specifications are given in the Appendix. This is a 36 V machine with 6100 rpm maximum speed. A Maxon driver (DEC 70/10) was also used to drive the motor. The current reference was derived from the desired speed profile and transferred to this driver through a computer. A prototype of the machine was built and tested. The solid model of the flywheel and the prototype are shown in Figure 3. The wheel is placed on ball bearings on both sides. A high-speed flexible coupling was used for the motor\u2013flywheel connection. A functional diagram of the test bed is seen in Figure 4. The switch in the figure removes the supply connections and connects the terminals of the machine to a three-phase load. The energy measurements are carried out directly at the generator outputs. The motor/generator and flywheel set was vacuumed down to 5 mbar for the experiments. The motor was driven for approximately 10 min at constant current to reach 5000 rpm" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002905_mop.4650030208-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002905_mop.4650030208-Figure1-1.png", + "caption": "Figure 1 Coupler configurations", + "texts": [ + " INTRODUCTION Microstrip antennas in \u201csuspended-substrate\u201d or in \u201cinverted\u201d configurations are often used where a large bandwidth, low loss, and light weight are required. No simple analytical methods, such as the transmission line or cavity models, exist which allow the determination of their resonance frequencies, unless the effective permittivity is very close to that of the air. One must resort to a more rigorous integral equation method involving a multilayer Green\u2019s function [l, 21. Even then the Figure 1 shows the setup used to perform the resonance frequency measurements. It consists of the following: 1. 2. 3. 4. 5. a ground plane made of a brass plate and fitted with SMA receptacles; a foam spacer of varying thickness, which sets the substrate height; a sample substrate with the patch-radiator to be measured; a Styrofoam block of height at least 2.5 - 3 wavelengths; an absorber layer sufficiently large to avoid reflections from the walls of the room. The foam materials used must be of hard (rigid) type, and be cut away around the radiator so as to minimize the field \u2019 distorsions" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003581_012042-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003581_012042-Figure5-1.png", + "caption": "Figure 5: Assembly schematic of proposed antenna [18]", + "texts": [ + " Furthermore, the changing structure of the antenna will affect the operating frequency and antenna performances. Another method to configure frequency is by using metasurface instead of the embedded switching circuit. Zhu et al. [18] invented frequency-reconfigurable antenna using metasurface. The proposed antenna consist of a simple circular patch antenna, design on Rogers substrates RO4350B with full ground plane. The metasurface of the proposed antenna is designed on a single-side substrate and composing of a number of rectangular-loop unit cells as shown in Figure 5. The metasurface of the proposed antenna constructed by placing rotating the metasurface around the center and relative to the patch antenna. As the rotation angle of metasurface increase from 10\u00b0, 25\u00b0, 35\u00b0, 55\u00b0 and 80\u00b0, the resonant frequency shifted to 4.77 GHz, 4.9 GHz, 5.07 GHz, 5.31 GHz and 5.5 GHz, respectively. This kind of method categories is mechanically reconfigurable antennas, where the parts of the antenna structures (metasurface) consist of movable parts for turning the frequency. The drawback of such designs was the difficulties of the fabricated process to ensure the moveable part can be still attached to the antenna and the same time can be flexible tuneable" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000223_1.4030653-Figure15-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000223_1.4030653-Figure15-1.png", + "caption": "Fig. 15 Free body diagram of human lower limb", + "texts": [ + " The spatial motion of saw-bones knee-model is measured by motion capture camera (OptiTrack V100:R2 [36]) which is tracking reflective markers attached on anatomical landmarks of femur and tibia bone in real time (100 Hz). The force and moment during flexion/extension motion are measured by 6DOF force transducer (ATI Delta [37]) which is rigidly connected to tibia bone of knee model. The midpoint between the head of the fibula and the medial epicondyle of tibia is defined as knee joint center [38]. Figure 15 shows the free body diagram of knee joint and load at knee joint are developed by a Newton\u2013Euler formulation as P Fx;Knee \u00bc Rx1 \u00fe Rx2 \u00bc mtaxtP Fy \u00bc Ry1 \u00fe Ry2 mtg \u00bc mtaytP Fz \u00bc Rz1 \u00fe Rz2 \u00bc mtaztP M \u00bc M1 \u00feM2 \u00fe L1t F1 \u00fe Lt2 F2 \u00bc Itat (11) In our testing, tibia is grounded, and various knee motions are performed relatively slowly and thus the above equations are reduced to the quasi-static equivalent. First, we measured the load at knee joint without PCCP/ PEB system and repeated same process (flexion and extension motion) with knee exoskeleton as shown in Fig", + " For example, this could help address limbbrace hyperstaticity which has traditionally been alleviated by an accommodating cuff\u2014poor cuff-designs have in the past led to undesirable chafing effects between soft tissue and knee brace. The formal examination and verification of the improved ergonomics and wearability benefits remains part of our ongoing work. This work was supported by the National Science Foundation (NSF) Award No. CNS-1314484. ax\u00f0y;z\u00det \u00bc acceleration of tibia in x (y, z) direction (Fig. 15 and Eq. (11)) b \u00bc longitudinal displacement of elastic band (Fig. 10 and Eq. (6)) dll \u00bc elongation of nonlinear spring only (Fig. 13 and Eq. (10)) dlnl \u00bc elongation of preload adjustor (prismatic actuator, Fig. 13 and Eq. (10)) dlnlp \u00bc overall elongation of nonlinear spring with preload adjustor (Fig. 13 and Eq. (10)) Fn \u00bc resultant force of PEB spring (Fig. 10 and Eq. (6)) Fp \u00bc force of an elastic band (Fig. 10 and Eq. (6)) F1\u00f02\u00de \u00bc force at ankle (F1) and knee joint (F2) (Fig. 15 and Eq. (11)) g \u00bc normal distance between inner and outer plate of PCCP mechanism (Fig. 5 and Eqs. (1) and (2)) It \u00bc inertial of tibia (Fig. 15 and Eq. (11)) Knl \u00bc spring constant of nonlinear spring (Figs. 7 and 11 and Eqs. (3)\u2013(5)) Journal of Mechanisms and Robotics NOVEMBER 2015, Vol. 7 / 041024-11 Downloaded From: http://mechanismsrobotics.asmedigitalcollection.asme.org/ on 01/29/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use Knlp \u00bc spring constant of nonlinear spring of semi-active model (Figs. 7 and 11 and Eqs. (3)\u2013(5)) lunit \u00bc length of fixed-guided segment (Eq. (5)) Lt1 \u00bc proximal length of tibia (Fig. 15 and Eq. (11)) Lt2 \u00bc distal length of tibia (Fig. 15 and Eq. (11)) l0 \u00bc initial length of nonlinear spring with preload adjustor (Fig. 13 and Eq. (10)) l0 \u00bc initial length of spring (Figs. 7 and 11 and Eqs. (3)\u2013(5)) L0 \u00bc length of plate of PCCP mechanism/Link length of pseudo-rigid-body model (Figs. 5, 7, and 9 and Eqs. (3)\u2013(5)) m \u00bc number of fixed-guide segment (Eq. (5)) mt \u00bc mass of tibia (Fig. 15 and Eq. (11)) M1\u00f02\u00de \u00bc moment at ankle (M1) and knee joint (M2) n \u00bc number of elastic bands (Fig. 10 and Eq. (6)) p \u00bc vertical distance of center of PEB (Fig. 10 and Eq. (6)) Rx\u00f0y;z\u00de;1 \u00bc reaction force at tibia end (ankle) in x (y,z) direction (Fig. 15 and Eq. (11)) Rx\u00f0y;z\u00de;2 \u00bc reaction force at knee joint (Fig. 15 and Eq. (11)) R1 \u00bc radius of inner plate of PCCP mechanism (Fig. 5 and Eqs. (1) and (2)) R2 \u00bc inner radius of outer plate of PCCP mechanism (Fig. 5 and Eqs. (1) and (2)) a1 \u00bc arc angle of inner plate of PCCP mechanism (Fig. 5 and Eqs. (1) and (2)) a2 \u00bc arc angle of outer plate of PCCP mechanism (Fig. 5 and Eqs. (1) and (2)) b \u00bc inclined angle of elastic band (Fig. 10 and Eq. (6)) c \u00bc characteristic radius factor (Figs. 7 and 11 and Eqs. (3)\u2013(5)) H \u00bc joint angle of pseudo-rigid-body model (Figs. 7 and 11 and Eqs" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001128_compel.2008.4634668-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001128_compel.2008.4634668-Figure2-1.png", + "caption": "Figure 2. 2kW class experimental system.", + "texts": [ + " ( ) Lqddq re TII L PM dt d P J \u2212\u03a6\u2212\u03a6=\u239f \u23a0 \u239e \u239c \u239d \u239b 2121 2 \u03c9 (5) rerestMss \u03c9\u03c9\u03c9\u03c9 \u2212+= * (6) Where, \u03c9re: rotor angular frequency, \u03a62d: rotor d-axis flux, \u03a62q: rotor q-axis flux, J: value of inertia, TL: load torque, \u03c9s: slip angular frequency. The inadequate slip frequency disturbs the rotor flux and sometimes reduces and increases the torque, which may result in much more errors of the estimated rotor frequency. So, we carry out experimental tests by a 2kW class experimental system. Experimental tests are carried out to clarify the influence of the mechanical parameter error by using a 2 kW class inverter and an induction motor with an inertial load. Fig. 2 shows their configuration and Table I shows their specifications. This experimental set has little enough rotational mechanical resistance to be regarded that TLset is zero. Fig. 3 is the test result in the case of Js=J, in which Js is the nominal value J(=10[kgm2]). Where, Te: torque, Te *: torque reference (calculated torque), fre: rotor frequency, frest: estimated rotor frequency. In this paper, the d-axis current and the q-axis current satnd for the stator d-axis current and the stator q-axis current respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000116_itoec.2018.8740499-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000116_itoec.2018.8740499-Figure6-1.png", + "caption": "Fig. 6. Imaging diagram of the binocular camera shooting back image", + "texts": [ + " We can use the binocular camera combined with the binocular head to directly pass back the depth of field map, because the indoor reconstruction does not need to use the higher cost map, so the Raspberry Pie can be used to connect the Wifi channel to return the depth of field image. The splicing hardware flow chart is shown in Figure 5. In terms of image stitching, we pick out the same part of the image of successive frames by image registration and overlapping feature matching, and obtain the spatial transformation matrix of image stitching and then image stitching. The imaging principle diagram is shown in Figure 6. As shown in Fig. 6, the shaded part in the figure is the overlapping part of the two frames of images. We determine the position of the shadow part by the image stitching algorithm, and perform image registration and stitching [15] to remove the overlapping areas of the two images. The image stitching algorithm can be used to obtain a longer image after multi-frame stitching, and then the depth map backed by the binocular camera is rendered using Open CV, and the threedimensional topographic map represented by the final depth map can be obtained" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000555_amm.486.239-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000555_amm.486.239-Figure3-1.png", + "caption": "Fig. 3 Model system in the ADAMS", + "texts": [ + " For systematic analysis of different cases and the possibility of comparing the results of each solution method, a simple (but sufficiently general) 3D model was primarily defined. This model met the required prerequisites. The model was designed to depict the laboratory model system as much as possible. The numerical model for investigating vertical vibrations was created in accordance with the model for experimental investigation. The base consists of a rigid plate flexibly stored on four springs, which was supplemented by two identical weights. Designed model for solving of vertical vibrations of the mechanical system in the ADAMS program can be seen in Fig. 3. The following figures show selected results of experimental and numerical solutions. In order to compare numerical and experimental solutions the results of both methods for each case are inserted into one figure. Each graph corresponds to the same asymmetry and the same kinematic excitation. All 3 position sensors are shown. The above results show good compliance between numerical results and experimental solution. Slight differences are caused by different material damping and spring stiffness" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002862_detc2007-34856-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002862_detc2007-34856-Figure6-1.png", + "caption": "Figure 6: Pivot-Arm CVT [8]", + "texts": [ + "org/ on 02/04/2016 T (diametral pitch = 16) Standard bicycles overcome the non-integer tooth problem by having a finite number of sprockets, both driving and driven, which all have the same diametral pitch. Each of these sprockets has an integer number of teeth, which allows them to mesh properly throughout their complete rotation. The pivot-arm CVT, originally developed by Mortensen [7], and later analyzed and modified by Christensen [8] at Brigham Young University, is an embodiment that employs compliant members to provide a mechanism that will change its active diameter to create a continuous range of mechanical advantage. Fig. 6 shows a specific design of the pivot-arm embodiment, meant for application in a bicycle drive train. The design consists of seven arms, called pivot arms, which are allowed to rotate about their connection to a common carrier about which they are attached. The pivot arms are connected to each other through compliant members which resist the rotation of the arms, and which cause all of the arms to rotate the same amount. As the arms rotate, causing the compliant members to deflect, the effective diameter of the CVT (front sprocket) is changed" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003460_esars.2015.7101415-Figure12-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003460_esars.2015.7101415-Figure12-1.png", + "caption": "Fig. 12. 3D thermal Analysis of the machine for 10000 rpm and full load operation (classic)", + "texts": [], + "surrounding_texts": [ + "Iron loss calculation of a Halbach array permanent magnet motor is presented. To reduce the iron losses calculation time six critical points are considered and iron losses are calculated in these points instead of all machine elements as it is done in pure finite element methods. This method significantly reduces the iron loss computation time in Halbach array permanent magnet motors without a considerable reduction of accuracy. The method is then used to calculate the iron losses in a high speed machine with different speeds and loads conditions. The results show that the proposed method predicts iron losses with less than 2% error with respect to FEM. Quantity Value (oC) Without Frame With Frame and Water Cooling Rotor 172.5 102.5 Stator 204.3 82 PM 173.1 102.6 Conductor 205.7 83.6 A 3-D FEM is then employed for thermal analysis of the motor. The analyses show that the temperature rise in permanent magnet section is more than permissible limit. Therefore a water cooling system is designed to reduce stator temperature. The FE analysis shows the cooling system reduces the permanent magnets temperature by more than 70oC degree and prevents thermal demagnetization." + ] + }, + { + "image_filename": "designv6_24_0003058_1.5082626-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003058_1.5082626-Figure1-1.png", + "caption": "FIG. 1. (a) Schematic view of the DSRR array. (b) Displacement of one DSRR owing to Ampe\u0300re\u2019s force and liquid resistance under an incident electromagnetic wave, where c is the wire radius of the enameled wire, d0 is the distance between two SRRs without an incident wave, d1 is the reduced distance with an incident wave, and r0 is the radius of the DSRR.", + "texts": [ + " Different from the previously reported cases,13,21 the force that remained within the metaatom immerged in a liquid-phase environment is liquid resistance. So, the force balance properties between the electromagnetic induction force and the liquid resistance force are completely new. Moreover, due to the fluid feature of the liquid host medium and the changeable distance of the metaatom, the proposed DSRR has more flexible structure controlling ability. It can be widely used to design the intelligent functional materials and devices in the future, and has the application of viscosity sensing for viscous media. Figure 1(a) shows the schematic DSRR array configuration we studied in this paper. Note that those parallel DSRRs are not printed on the conventional substrate but immerged in low-loss liquid-phase media. We fabricated the simple metaatom with an enameled wire and then hung it with an acrylic rod. Assuming there is an incident electromagnetic wave acting on the metaatom, the DSRR will generate an Ampe\u0300re\u2019s force due to electromagnetic induction.21,22 Because the splits of the DSRR are in-phase, the induced currents generated in the two SRRs have equal magnitude and phase, resulting in an Appl. Phys. Lett. 114, 144101 (2019); doi: 10.1063/1.5082626 114, 144101-1 Published under license by AIP Publishing inter-attraction Ampe\u0300re\u2019s force between the two rings. When the intensity of the incident electromagnetic wave is large enough, the produced Ampe\u0300re\u2019s force will cause the DSRR\u2019s mechanical displacement, and then the structure will be changed, as shown in Fig. 1(b). However, all these changes will be very tiny, so in the actual measurement, it will be shown and characterized by electromagnetic resonance frequency offset.23 For both the gas-phase environment and the liquid-phase environment, the impetus acting on the DSRR is Ampe\u0300re\u2019s force, just with different magnitudes because of the difference in the dielectric parameters of the host media. However, it is noteworthy that the resistance of the DSRR in both environments can be very different.24 The detailed theoretical derivation for the Ampe\u0300re\u2019s force calculation has been extensively studied in many works" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000535_0954406214544726-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000535_0954406214544726-Figure8-1.png", + "caption": "Figure 8. Segments of Welding used in sequential welding.", + "texts": [ + " The run-outs have been measured for all the nine numbers of such bowls, against three different clamping conditions, four different speeds and three different sequences for a single pass. In order to compare the effect ofmulti-pass welding against single pass, the same has been also checked for one case as well. In total, there are three circumferential welds required to complete each bowl, and it is necessary to provide the weld joint totally from outside. It is proposed to divide in to three segments spacing 120 each as shown in Figure 8. A quantitative measurement of experimental transient temperatures histories at three thermocouple locations starting at 0 and ending 120 is presented in Figure 9(a). Similarly, during welding of the second segment in addition to those three thermocouples, previous thermocouple data (which is at 90 from weld starting point) are also plotted as shown in Figure 9(b). A typical result of thermal simulation is presented in Figure 10, in which the 3D temperature profile on the outer surface at 45 s for the simulation condition of point\u2013point clamping condition with 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000546_amc.2006.1631751-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000546_amc.2006.1631751-Figure1-1.png", + "caption": "Fig. 1. Rigid motion of an object on a conveyor belt under a scaled orthographic camera", + "texts": [ + " f \u2192 \u221e, the perspective projection is reduced to an orthographic projection given by x = X, y = Y (2) When the depth of objects, \u03b4z, is much smaller than their average distance, Z\u0304, from the camera along the optical axis, i.e. |\u03b4z| Z\u0304, the full perspective camera model can be replaced by the weak-perspective or so called scaled orthographic camera model. An example for this is the camera located at a sufficiently large distance, d, above a plane of a conveyor belt, with moving piece-parts over it as in Figure 1. In this case Z\u0304 = d. The perspective projection given in (6) takes the following form: x = f X Z\u0304 = f X d , y = f Y Z\u0304 = f Y d (3) If a rigid body is moving with instantaneous translational velocity, T , and rotational velocity, \u03a9, then the 3D instantaneous velocity of points on the surface is given by\u239b \u239dX\u0307 Y\u0307 Z\u0307 \u239e \u23a0 = \u03a9 \u00d7 \u239b \u239dX Y Z \u239e \u23a0 + T (4) \u239b \u239dX\u0307 Y\u0307 Z\u0307 \u239e \u23a0 = \u239b \u239d 0 \u2212\u03c93 \u03c92 \u03c93 0 \u2212\u03c91 \u2212\u03c92 \u03c91 0 \u239e \u23a0 \ufe38 \ufe37\ufe37 \ufe38 [\u03a9]\u00d7 \u239b \u239dX Y Z \u239e \u23a0 + \u239b \u239db1 b2 b3 \u239e \u23a0 \ufe38 \ufe37\ufe37 \ufe38 T (5) where \u03a9 = (\u03c91, \u03c92, \u03c93)T and T = (b1, b2, b3)T " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002388_2015-01-0090-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002388_2015-01-0090-Figure3-1.png", + "caption": "Figure 3. FE model of the truck frame with meshes of each member separately assigned.", + "texts": [ + " For this reason, a shell element type was selected to generate Finite Element (FE) model of the truck frame. In most cases, appropriate shape of elements of two-dimensional configuration should have an Aspect Ratio below five [5]. For the simulations, the frame CAD model was manually converted into a Finite Element (FE) model using FE pre-processor HYPERMESH. To achieve appropriate mesh quality, Aspect Ratio and Skewness of each element were limited to 5 and 0.45 respectively. Details of the complete FE model of the truck frame (Figure 3) are displayed in Table 1. The FE model was later imported into ANSYS for simulations. Multi-Point Constraint (MPC) contact formulation was selected to create bonded contacts between adjacent parts. MPC can be employed for bonded contacts between parts meshed with shell elements [6] and it is effective for bonded contact regions [7]. In this case, local mesh adjustments could be achieved without affecting the global mesh density. This illustrated the advantage on local mesh adjustments. Yield Strength and Tensile Strength of a material were 458 MPa and 594 MPa respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001922_aps.2016.7696205-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001922_aps.2016.7696205-Figure1-1.png", + "caption": "Fig. 1.Geometry of the (a) multilayer antenna design and (b) top patch. The optimized dimensions are Ls= 52.85 mm, rp1= 6.94 mm, hs1= 2.662 mm, hs2= 2.921 mm, hs3= 2.745 mm, lc= 47.82 mm, wc= 35.23 mm, ld= 3.01 mm, wd= 1.74 mm. The substrates permittivities are \u03b5r1=10.2, \u03b5r2=4.5, and \u03b5r3=10.2.", + "texts": [ + " Several wearable antenna designs have been developed over the years for onbody communications; for example in [2-4] microstrip patch antennas designed to excite a higher order modewere studied. Also, a button-like antenna was designed in [5]. All of these designs have nearly omnidirectional radiation patterns parallel to the body. In this paper, we propose a compact, low profile multilayer stacked-patch antenna for on-body communication without the need to excite higher order modes. II. ANTENNA DESIGN AND PERFORMANCE The antenna structure considered throughout this paper is shown in Fig. 1.The proposed stacked-patch antenna employs a coupled feed at its center comprised by a circular patch with a radius rp1 in the lower layer of the substrate. A cross-cut parasitic patch is located in the middle layer, and a substrate is added on top of this layer (top layer) to improve the impedance matching. The geometrical parameters of the antenna were optimized using BORG [6], a multi-objective algorithm that was employed to provide a trade-off solution between the maximum realized gain in XZ- and YZ-planes, a null at boresight, an omnidirectional pattern in the XY-plane, and the substrate thickness (Ls)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003997_asms-spsc.2010.5586910-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003997_asms-spsc.2010.5586910-Figure2-1.png", + "caption": "Figure 2. Explanatory diagram for the intersection points.", + "texts": [ + " Since the plane of the satellite orbit is fixed in the fixed equatorial coordinate system due to the assumption, that there are no perturbations, the SSP moving in the plane of the satellite orbit creates a circular path in this plane on the sphere of Earth. In order to transform the three dimensional problem into a two dimensional one, the right ascension of the intersection points of cgs and cEOC have to be found as a function of the elongation of the satellite in the fixed equatorial coordinate system. Fig. 2 shows an example, in which the two small circles intersect each other. The variables \u03b11 and \u03b12 denote the right ascension of the above mentioned intersection points, whereas \u03c9 \u2208 [0, 2\u03c0] denotes the elongation of the satellite. Due to the fact, that the right ascensions of the intersection points are not related to the rotation of the Earth, the rotation of the Earth is not relevant at this point. The rotation of the Earth will be taken into account in Subsection II-D. The right ascensions of the intersection points can be plotted in a two dimensional diagram as a function of the elongation of the satellite, and since there can be zero, one or two intersection points at certain satellite position (at a certain elongation of the satellite), this function will be a multiplevalued function" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003391_ijvnv.2012.046175-Figure11-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003391_ijvnv.2012.046175-Figure11-1.png", + "caption": "Figure 11 Rolling motion with regards to time, inspired by passive walking", + "texts": [ + " By passing v through a neuron with a small time delay \u03b5v as in Zaier and Nagashima (2004), so that v + \u03b5vdv/dt = \u2013\u03b7(t, x)sgn(w) with wT = [\u2202V/\u2202x]G and \u03b7(t, x) \u2265 \u03c1(t, x)/(1 \u2013 \u03ba0, the perturbed system of equation (35) becomes stable and the chattering problem due to term sgn(w) will be solved. With regards to the motion along the pitching direction, besides the gyro feedback control we consider the robustness against rough ground (ground with some surface irregularity) by controlling the time-varying parameters of the virtual spring damper system bs(t) and ks(t) in equation (37) during the gait as described in Figure 11. That is, to minimise the force of collision at the landing time t1, the damping coefficient bs(t1) is set at its maximum value while ks(t1) is at its minimum value. At the lifting time t2, we inject spring energy into the system by setting bs(t2) at its minimum while ks(t2) is at its maximum. We consider linear time-parameter changes between lower/upper limits. Hence, a smooth commutation between equilibrium points of equation (8) is achieved and we obtain a passive like walking behaviour (Zaier and Kanda, 2007)", + " Second, we calculated the rolling motion parameters by investigating the dynamics of the robot during stepping motion, using the phase portrait of the ZMP position in the lateral plane. The phase portrait of the ZMP position during stepping motion is shown in Figure 13, which is similar to the simulation result in Figure 9. Note that the more we increase the derivative gain the better convergence we will have, however, we are limited by the high frequencies resonance that can be triggered at high feedback gain as shown in Figure 12. To demonstrate the effect of the virtual damper-spring system of equation (37), which is implemented to work as illustrated in Figure 11. Hence, Figure 14 demonstrates how the robot uses gravity for landing and the spring energy for lifting. Figure 14(a) shows that the lifting phase starts when the angular position of the rolling joints is almost zero. In other words, the actuators of the hip and ankle rolling joints do not contribute in moving the ZMP to the supporting leg. These joints, therefore, can be considered as locked joints. Figure 15 shows the rolling of the hip and the pitching motions of the knee, hip and ankle. It demonstrates also how smooth the approximate solution is using the proposed pattern generator with sensory feedback" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002571_isem.2018.8442835-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002571_isem.2018.8442835-Figure3-1.png", + "caption": "Fig. 3. Magnetic flux density for I=35A; a) for zero position of the rotor in relation to the stator =0\u00b0; b) for the stator angle in relation to the stator of =14\u00b0", + "texts": [ + " With the aim of reducing the numerical costs required for analysis of the 3D model, the symmetry in the XY plane of the analyzed machine was used and the constructed model was limited to half of the machine volume. The conducted preliminary calculations of electromagnetic field show that the armature plays the role of the magnetic pole without winding. The lack of armature breaks the magnetic circuit resulting in the drop of the generated torque. A fundamental aspect of the stator design is the length of the armature, which should not exceed the stator diameter. Fig. 3 presents the distribution of the magnetic flux density for two positions of the rotor in relation to the stator. 978-1-5386-5210-7/18/$31.00 \u00a92018 European Union Fig. 4 presents the relation of the electromagnetic torque for phase C with power supply in the range 0 \u2013 35A (DC) gained on the basis of the 3D model. The differences in the results of calculated electromagnetic torque gained by application of 2D and 3D models do not exceed 3.3%. The maximum torques determined on the basis of the examined models amount to Te_max2D=9" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003809_0921-5093(94)09732-1-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003809_0921-5093(94)09732-1-Figure4-1.png", + "caption": "Fig. 4. Schematic diagram of the specimens tested in shear with the gripping arrangement keeping them aligned. All dimensions in mm.", + "texts": [ + " Although it does not appear explicity, the small angle, O, which can be satisfactorily approximated by tan 0, depends upon yield strength as the ratio of yield stress divided by Young's modulus. 3. E x p e r i m e n t a l p r o c e d u r e s The material under investigation is commercial aluminium alloy, similar to the heat-treatable IADS 2618. The chemical composition is given in Table 1. Cylinder samples of thickness L = 6 mm were cut off from a round bar of radius 40 mm, which was initially rectified to a surface finish of 0.3 mm. The samples were machined and drilled parallel to the cylinder axis very carefully, as indicated in Fig. 4. The load was applied in a vertical direction and the specimen was aligned horizontally and remained so while it was being stressed. This specimen design is a modification of the original development of Schroth et al. [20], and was chosen to ensure that the specimen remained aligned as it was being stressed in pure mode III. Two series of specimens were examined. All of them were drilled on a straight line at equal distances with three 5 mm holes as shown in Fig. 4. The first series of specimens were pre-cracked at length ~ = 2 mm, thickness 0.35 mm and radius 0.15 mm, whilst the second had a crack of the same length with a crack thickness 0.16 mm and a flat end. The cutt ing was done using a low-speed d iamond wheel and, finally, the specimens were metal lographically prepared on both top and bo t tom surfaces down to a d i amond polish o f 1/4 gm. For a shear testing the specimens were moun ted into the testing machine with a gripping ar rangement schematically shown in Fig. 4, the outer pair o f holes being at tached to the upper grip, the central hole to the lower. In order to study the extent o f the plastically deformed region at the end of the crack, the polished surface was examined before loading and at various stages o f loading. Metal lographic examinat ion was carried out both on the as-polished surface and on an etched surface using appropr ia te etchant [21] so as to reveal the microstructural features. Both top and bottom surfaces o f each sample were observed" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000120_cicc.2007.4405676-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000120_cicc.2007.4405676-Figure1-1.png", + "caption": "Fig. 1. Composition of the Flex-PG SRAM cell. Pass gates, MN3 and MN4, consist of 4T-FinFETs and the flip-flop consists of usual \u201c3T\u201d-FinFETs.", + "texts": [ + " In this article, a FinFET-based 6-transistor SRAM, which enhances both the read and the write SNMs without increase in the cell size, is proposed, and its performance in enhancing SNMs is demonstrated by using a TCAD simulation [1]. II. FLEX-PG SRAM CONCEPT The pass gates in our proposing SRAM cell consist of double-\u201cindependent\u201d-gate FinFETs, which have four terminals of first gate, second gate, source and drain as four-terminal- (4T-) FinFETs. The 4T-FinFET has a variable threshold voltage controlled by the second gate voltage, VG2. The topology of the SRAM circuitry is shown in Fig. 1. Since two 4T-FinFETs of pass gates have the tunable threshold voltage, a ratio of drivability of the driver to the pass gate (\u03b2-ratio) is flexibly changed. In the read operation, the \u03b2-ratio is increased by raising Vth of the pass gates. On the other hand, in the write operation, a write SNM is enhanced by lowering Vth of the pass gates. Consequently, the Vth flexibility in the pass gate makes an SRAM cell free from a trade-off relation between the read and the write SNMs. We named it \u201cFlex-pass-gate (PG) SRAM\u201d" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000829_ted.2010.2066279-Figure13-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000829_ted.2010.2066279-Figure13-1.png", + "caption": "Fig. 13. Experimental microphotographs (a) for buried n-islands and (b) for the HVIC with a BNIL LDMOS [different scales in (a) and (b)]. (c) Tilted section of the fabricated BNIL LDMOS, with a tilted angle of 2\u25e652\u2032 to the wafer surface.", + "texts": [ + " The key processes of a BNIL SOI wafer include the following: 1) patterned implantation of arsenic to form nislands, followed by thermal oxidation on the device wafer; 2) oxidation on the handle wafer; 3) bonding two oxide surfaces; and 4) thinning the device wafer. The impurities in the n-islands diffuse during the thermal oxidation and bonding high-temperature processes. In order to avoid the dualsided alignment and to improve the alignment tolerance, the n-islands are patterned as close-packed hexagon arrays, owing to the best symmetry in a random direction compared with any other geometric arrays. The top view of the layout with close-packed hexagon arrays is illustrated in Fig. 13(a). The microphotograph for the HVIC is given in Fig. 13(b), in which the circular BNIL LDMOS is designed. The different alignments in the experiment, which are denoted by different values of LBN, La, and Lb, are investigated in 2-D simulation. Fortunately, the alignment tolerance is large, as demonstrated in Fig. 10. The dual-sided alignment process can hence be removed in this experiment. In comparison with the conventional silicon directly bonding technology for SOI wafers, only the patterned arsenic implantation followed by a short-time annealing process is added to the BNIL SOI wafer fabrication. Fig. 13(c) gives the tilted section (not a cross section and a top view) of a fabricated BNIL SOI LDMOS, with a tilted angle of 2\u25e652\u2032\u2032 to the surface. Fig. 13(c) partially shows the structure in the z-direction, in addition to those in the x- and y-directions, displaying the vertical structure of the n-drift/p-SOI/n-islands/BOX/Si substrate, as well as the closepacked hexagonal n-islands in the x- and z-directions. The measured OFF- and ON-state I\u2013V curves for the BNIL LDMOS is shown in Fig. 14. A 660-V LDMOS is obtained. Fig. 14 has verified the feasibility and validity of the new concept. A 660-V BNIL LDMOS in an SOI HVIC has been proposed by introducing additional positive charges at the interface of the SOI/BOX to enhance the field strength at the interface" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000849_biorob.2008.4762805-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000849_biorob.2008.4762805-Figure2-1.png", + "caption": "Fig. 2. Tensegrity model of the cell.", + "texts": [ + " Tensegrity model The cytoskeleton model is composed of an interconnected structure of various cross-linked and interlinked filamentous biopolymers that extends from the center to the cell surface (bio-membrane). Three major filamentous components composes the cytoskeleton, i.e. the actin microfilaments, microtubules and intermediate filaments, and are physically interlinked. In our simulation, we choose a simplified tensegrity structure with six compressing struts (two in each orthogonal direction), these struts aggregate behavior of the microtubules, as viewed as beams as shown in Fig.2. These struts are attached to 36 cable segments: 12 cables representing the intermediate filaments which are connected to the parallel struts and 24 cables representing the actin microfilaments connected to the end points of each strut. Real time deformation are usually derived using either mass-spring approach or Finite Element Method (FEM). Mass-spring models are less \u201dphysically based\u201d. They comprise a set of nodes connected by springs, with point masses attached at each node. Real time performance can be achieved with a limited number of nodes but the behavior is often unrealistic and can be unstable", + "m j I ) (4) The equation (3) defines the force f i for one tetrahedron T K , but for global mesh, we must add the contribution by all adjacent tetrahedra of T K , the resulting force F i can be expressed as follows : F i mesh = IKii .ui + \u2211 j\u2208O(Pi) IKi j .u j (5) IKii and IKi j design respectively the sum of tensor Kii associated with the tetrahedra adjacent to node i and with the tetrahedra adjacent to edge (i, j), O(Pi) is the neighborhood of vertex Pi. These tensors, depending only on the rest geometry and Lames coefficients, are constant and can be pre-computed. B. Cable behavior for the tensegrity structure In this section, we present the description of the cable behavior about our tensegrity model (Fig.2). These cables are assumed to be have as viscoelastic mass-spring-damper. Each cable is modeled with two masses interconnected via spring and damper in parallel (Voigt-Kelvin model, Fig.4). In the local frame, the relation between the stress and the strain can be written as follows: \u03c3 = E \u03b5 +\u03b7 \u03b5\u0307 (6) with \u03c3 = F S and \u03b5 = l\u2212 l0 l0 , where F : is the applied load in the extremity of the cable, with l and l0 and S are respectively the resting length of the cable, the initial length and the section of the cable" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003699_edssc.2012.6482775-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003699_edssc.2012.6482775-Figure4-1.png", + "caption": "Figure 4. (a) Small-signal equivalent model of single TIA stage. (b) Equivalent RLC network.", + "texts": [ + " It is noted that this enhancement is at the expense of signal delay which should be acceptable in a typical optical receiver front-end. The effective low-frequency gain of the \u03c0peaking TIA when the outputs are combined is similar to its conventional shunt-feedback counterpart, which is 1 1 1 1 1 d m F D m D v g R R i g R \u2212 = \u2212 + (1) A more complete bandwidth enhancement operation of the proposed TIA can be examined by first considering the RLC equivalent network model at the input terminal of the TIA stage. Consider the small-signal equivalent circuit for one of the single feedback TIA stage as shown in Fig. 4(a) where it is assumed that the gate-drain capacitance Cgd of the MOS amplifier transistor can be omitted. In the model, Cgs is the gate-source capacitance and gm is the transconductance of M1. It can be shown that the frequency characteristic of the input impedance Zi, can be modeled by the RLC network of Fig. 4(b). The capacitance C1 in the RLC network is equal to the gate-source capacitance Cgs of the amplifier transistor, i.e., 1 gsC C= (2) Also, the following parameter relations are derived, 1 1 (1 ) 1 F F D m F R R R R g R \u2212+ = \u2212 (3a) 2 FR R= (3b) 2 1 1 F L m F R C L g R = \u2212 (3c) In summary, the input impedance Zi, of the single TIA stage can be modeled by the RLC network using the parameter relations in (2), and (3a) to (3c). With the use of the RLC network in Fig. 4(b), the equivalent network model for the proposed TIA of Fig. 3 can be depicted in Fig. 5. Upon inspecting the figure within the dash box, it is recognized as the \u03c0-type inductive peaking network in [6], and hence the proposed circuit is called \u201c\u03c0\u2013 peaking TIA\u201d. In operation, the input current signal from the photo diode successively flows into the divided TIA stages due to the effect of the series inductance L , yielding a bandwidth extension as previously explained. In addition, within each of the stages, there exists an effective inductor L1 in series with the resistance R1", + " Since the capacitance C1 is determined by the gate capacitance of the MOS amplifier transistors as given in (2), it follows that the bandwidth of the \u03c0\u2013peaking TIA can be optimized by proper choices of the inductors L and L1. Fig. 6 shows a practical schematic of the \u03c0\u2013peaking TIA with the cascaded amplifier stages (M2) and the signal combining network (Lo1, Lo2 and Co) at the output. It is noticed that a series inductor LL is employed between the TIA and the second stage. This enables us to adjust the effective load capacitance CL of the TIA [cf. Fig. 4], and hence the value of the effective inductance L1 as indicated by (3c). In this way, the bandwidth of this practical \u03c0\u2013peaking TIA can be optimized by proper selection of the inductors L and LL. The inductors Lg are also included in the practical circuit to compensate for the second-order effect, such as the effect of Cgd which is omitted in the description of the circuit operation. For the output network, it makes use of the non-uniform constant-k LC network developed in [8] to efficiently combine two time-shifted signals from the cascaded TIA amplifier stages" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003634_powereng.2015.7266300-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003634_powereng.2015.7266300-Figure3-1.png", + "caption": "Fig. 3. Variation of magnetic field density at 3 ms while the motor is rotated at approximately 1057 rpm.", + "texts": [ + " The tasks accomplished can be summarized as follows; \u2022 In order to determine distribution of the magnetic flux in stator, stator windings rotor and magnet, a 2D simulation model has been derived. \u2022 Also, the electromechanical parameters like electromotive force, torque and stator resistance and inductance are obtained from the model. Electric machine modelling and numerical methods to solve them have been realized by using ANSYS Ansoft EM 2014 package. Some quantities such as magnetic flux density has been calculated and visualized as seen in Fig.3. \u2022 Furthermore, a Simplorer/Maxwell Co-simulation model (Fig.4) for the analysis of motor performance in dynamic conditions is developed and results are given. IV. DYNAMIC STUDY OF THE BLDC MOTOR A final design consideration of the motor is the analysis of the dynamic analysis of the motor. For this a physical electromechanical model of the motor was implemented in MatlabSimulink Environment. The screenshot of the model is seen in Fig. 7 below. In the simulation, the motor model is implemented with the parameters obtained as the result of the design process described earlier" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003536_978-94-009-5063-4_1-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003536_978-94-009-5063-4_1-Figure8-1.png", + "caption": "Figure 8. Different types of load-displacement relations. A is the load; wb is the buckling modal displacement.", + "texts": [ + " Wbo would characterize the sensitivity of the maximum load AS to initial geometrical imperfections. According to the jargon that has become accepted over the years, the structure to which the curves in Fig. 7(b) correspond is called \"imperfection sensitive\" because imperfections reduce its maximum load carrying capability. (Of course, it is not the structure that is sensitive to imperfections, but the maximum load it can safely support!) Neither all structures nor mathematical models of them behave as shown in Fig. 7(b). Figure 8 shows various types of post-buckling behavior. A linearized model of elastic stability, that is an eigenvalue formulation of the buckling problem, implies load-deflection behavior shown in Fig. 8(a): The amplitude of the eigenvector, the bifurcation buckling mode, is indeterminate, which implies that the load A remains constant A = AC with increasing deflections Wb' The equilibrium path for the slightly imperfect structure follows the rectangular hyperbolic path, (1) shown dotted in Fig. 8(a). If nonlinear post-buckling effects are accounted for, equlibrium paths for most structures have the forms shown in Fig. 8(b, c, d). The asymmetric nature of the curves in Fig. 8(b) indicates that the structure continues to carry loads above the bifurcation load AC if it is forced to buckle one way, but collapses if allowed to buckle the other. An example of this type of behavior is a structure with parts that move relative to each other as buckling proceeds in such a way that these parts come in contact and support each other for positive deflections but move away from each other forming gaps for similar negative deflections. Specifically, a built-up panel consisting of a flat sheet riveted to a corrugated sheet is such a structure. Roorda [5] has demonstrated this asymmetric post-buckling behavior for perfect and imperfect frames with eccentric loads. His results are presented in [4]. The symmetric stable post-buckling behavior displayed in Fig. 8(c) is typical of axially compressed isotropic flat plates. The perfect plate loaded precisely in its neutral surface buckles either way with equal ease and the post-buckled equilibrium state is stable. The symmetric unstable post-buckling behavior shown in Fig. 8(d) is typical of the early post-bifurcation regimes of axially compressed thin cylindrical shells and externally pressurized thin spherical shells. Capsule of recent progress in buckling analysis Recent progress in our capability to predict buckling failure can be categorized into three main areas: (1) development of asymptotic post-buckling theories and applications of these theories to specific classes of structures, such as simple plates, shells, and panels [6-8] ; (2) development of general-purpose computer programs for calculation of static and dynamic behavior of structures including large deflections, large strains, and non linear material effects [8-10] ; (3) development of special-purpose computer programs for limit point axisym metrical buckling and nonaxisymmetrical bifurcation buckling of axisymmetric structures [11-14, 171]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003634_powereng.2015.7266300-Figure12-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003634_powereng.2015.7266300-Figure12-1.png", + "caption": "Fig. 12. A commercial in-wheel electric motor by Protean Company.", + "texts": [ + " Future studies going to focus on eddy current loss minimization in permanent magnets as well as copper loss optimization in stator windings as studied in [14-17]. Also load tests will be carried out by coupling the motor in the enclosure between the wheel rim and the drum brake housing, where it will be tested for performance to assess whether the design is suitable as part of an electric car conversion kit. The new design has been compared with a commercial in-wheel electric motor designed and manufactured by Protean Company. The differences of both designs can be evaluated in Fig.11 and Fig. 12. It is clear that new in-wheel electric motor is specifically designed for the use electric vehicle conversions of existing cars. Since the design is suitable for placing in the enclosure between the wheel rim and drum brake housing, it differs from existing technologies. This work is jointly funded by Small and Medium Enterprises Development Organization of Republic of Turkey (KOSGEB) under the project number 2014-692-7/02 and Turkish Scientific Research Council (TUB TAK) under grant number 113M070" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002665_icscs.2008.4746910-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002665_icscs.2008.4746910-Figure1-1.png", + "caption": "Fig. 1. Whirling pendulum.", + "texts": [ + " \u2016x(t)\u2016 < \u03ba\u2016x(t0)\u2016e\u2212\u03b3(t\u2212t0), \u2200t0 \u2264 t \u2264 t1 \u2016x(t)\u2016 \u2264 b , \u2200t > t1 For t \u2265 t1 , \u03ba = \u221a c2 c1 , \u03b3 = c3 2c2 , b = c2 c3 \u221a c2 c1 \u03c3 (38) Suppose now the perturbation term satisfies \u2016d(t, x)\u2016 \u2264 r b \u03c3 (39) Applying (32), (33), (34) and (35) to (31) the expression of the perturbation can be expressed by : d(t, S) = \u2212\u03b7sgn(S) (40) with S\u0307 = \u2212kS + d(t, S) So the nominal system S\u0307 = \u2212kS has an exponentially stable equilibrium point at S = 0, since k > 0. The norm of perturbation (38) is bounded and can be expressed \u2016d(t, S)\u2016 \u2264 \u03b7 (41) And when drive the lyapunov function we get V\u0307 = S(\u2212kS \u2212 \u03b7sgn(S)) V\u0307 \u2264 \u2212k\u2016S\u20162 \u2212 \u03b7\u2016S\u2016 and when takes c1 = c2 = 1c3 = k, c4 = 1 and b = \u03c3 k (42) from (37), (39) and (40) d(t, S) \u2264 kr S \u2264 \u03c3 k To demonstrate the effectiveness of the propose approach, we consider the 2-DOF whirling pendulum (Figure 1). The whirling pendulum is a planar pendulum whose suspension point is attached to another masse M by means of a vertical shaft .We will ignore frictional effects acknowledgement. the parameter of system presented as follows: l : pendulum length,I : Bob inertia around its center of gravity, m : pendulum bob mass,M : whirling mass, g : gravitational acceleration, R : radius of arm, \u03c4 : shaft torque. \u03b8 : angle of pendulum from the upward vertical \u03c6 : angle of mass from -4- a fixed vertical plane. The differential equations describing the system are then [MR2 + m(R 2 + l2 sin2(\u03b8)]\u03c6\u0308 + mRlcos\u03b8\u03b8\u0308 + mRlcos\u03b8\u03b8\u0308 +ml2 sin(2\u03b8)\u03c6\u0307\u03b8\u0307 \u2212 mlR sin \u03b8\u03b8\u03072 = \u03c41 mRlcos\u03b8\u03c6\u0308 + (I + ml2)\u03b8\u0308 \u2212 ml2 sin \u03b8cos\u03b8\u03c6\u03072 \u2212 mglsin\u03b8 = 0 The dynamic can be written as (7) where" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002441_embc.2013.6609739-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002441_embc.2013.6609739-Figure3-1.png", + "caption": "Fig. 3 (a), (b) and (c) shows the longitudinal shear stress distribution and (d), (e), (f) shows the circumferential shear stress distribution at the limb\u2013socket interface when loadings simulating the three walking phases were applied with pre-stress considered. The maximum value is 25.65 kPa and 103.6 kPa over the posterior of the socket brim region for the longitudinal shear stress and circumferential shear stress, respectively.", + "texts": [ + " The stress distributions can be shown in socket and stump, but focus on the surface of the stump with the socket\u2013stump interaction will be made. Fig. 2(a) shows the normal stress distributions predicted from the first step to simulate the pre-stress, which were more evenly distributed. Fig. 2, (b), (c) and (d) displays the normal stress distribution obtained from the second step analysis in the bottom surface when loadings simulating the three walking phases were applied with pre-stress considered. The normal stresses up to a Figure 3. The longitudinal shear stress for loading conditions at foot flat (a), mid-stance (b) and heel off (c) and the circumferential shear stress for loading conditions at foot flat (a), mid-stance (b) and heel off. (c) stress maximum of 80.57, 52.41 and 73.37 kPa, respectively. In all this three cases, the highest normal stress was produced at the ischial bearing areas up to 119.3, 89.98 and 104.1 kPa, respectively. It was assumed in the FE model that the femur position did not change within the soft tissue at different loading cases" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002453_978-3-030-20131-9_174-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002453_978-3-030-20131-9_174-Figure2-1.png", + "caption": "Fig. 2. Large-deflection beam model (left) and pseudo-rigid body model (right). The large displacement of the beam\u2019s extremity, in the beam support frame, is a circle centered at K.", + "texts": [ + " A precise model of the system is highly coupled and does not lend itself to a closedform solution; the orientation of the force on the idler depends on the idler position and the resulting belt angles, but the orientation of the force influences the beam deflection and therefore the idler position itself. Therefore, a model is pursued which approximates beam deflection behavior based on a widely accepted model for large displacements of compliant beams, and assuming that stiffness of the belt is much greater than that of the compliant beam (i.e., the belt acts as a kinematic constraint (inextensible), with no pretension at the neutral position of the idler). As we can assume equal angles of the input and output wheels (neutral, or no relative displacement), the path of the idler (point P in Figure 2) allowed by the belt is an arc of an ellipse. The considered ellipse is described by its major and minor axis lengths, respectively ae and be, as shown in Figure 3. . With the focal distance c prescribed (the half-spacing of the input and output pulleys), and the minor axis set equal to the y-component of the undeflected beam (i.e., no preload in the beam as the idler travels across the top-most point of the ellipse), the ellipse axes are given by be = Lbsin( ) (1) ae = (c2 + be 2)1/2 (2) According to the widely adopted pseudo-rigid body model of a compliant fixedpinned beam of length Lb under tip loading with large deflections (nonlinear beam) [17], it has been shown that the beam tip traces an approximately circular deflection path, with the center of that path located a distance Lb(1 \u2013 ) from the base of the beam, and an equivalent torsion spring K located at the circle\u2019s center (see Figures 2 and 4)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003952_978-3-642-33457-3_7-Figure7.1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003952_978-3-642-33457-3_7-Figure7.1-1.png", + "caption": "Fig. 7.1 Section view of the AMB spindle, taken from [1]", + "texts": [ + " The Chapter will start with presenting development of new spindle concepts in the Microfactory project group\u2014Sect. 7.2. Thereafter, geometric and electromagnetic design of the spindle motor will be explained in Sects. 7.3 and 7.4, then evaluated using FEM in Sect. 7.5 and, lastly, optimization of the rotor retaining sleeve is presented in Sect. 7.6. In the Microfactory project a small spindle in active magnetic bearings (AMB) was realized by Kimman [1, 2]. A relatively slender rotor is supported by two radial bearings and an axial bearing which exerts force over a small rotor disc (Fig. 7.1). A. Borisavljevic\u0301, Limits, Modeling and Design of High-Speed Permanent Magnet Machines, 129 Springer Theses, DOI: 10.1007/978-3-642-33457-3_7, \u00a9 Springer-Verlag Berlin Heidelberg 2013 Rotation of the spindle is controlled by a commercially available PM motor. The maximum attained speed of the spindle is 150.000 rpm. It was shown that miniaturization of a spindle has positive effects on actuation and cutting force monitoring [1, 3]. The first flexural critical speed (approximately 180000 rpm) of the spindle has represented an obstacle of utilizing the motor up to its maximum speed constrained by the structural limit of the PM rotor (250000 rpm)", + " However, that was hardly achievable with the same motor-bearings configuration. Hence, Kimman et al. [4] proposed a whole new approach for high-speed spindles: to use a short (disc-shaped) rotor suspended in AMB. The inspiration was found in an idea of 3DOF combined axial and radial magnetic bearings envisaged in [5]. Kimman et al. [4] proposed using such bearings for supporting 5DOF of a disc thus benefiting from reducing rotor tilting and higher resonance frequencies (Fig. 7.2). In essence, it would mean that all the bearings from the original setup\u2014Fig. 7.1\u2014would be grouped around the axial-bearing disc and that would, in turn, drastically reduce the spindle volume. Advantages of using short rotor follow also from analyses of Chap. 5. A rigid short rotor has one critical speed less than its long/slender counterpart and it may also benefit from increase of critical speeds as a result of gyroscopic stiffening (see Fig. 5.12 in Sect. 5.5). Still, the greatest advantage of such a rotor clearly comes from the increase of flexural resonance frequencies as a result of the great reduction of the rotor slenderness" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001835_j.mechmachtheory.2012.11.004-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001835_j.mechmachtheory.2012.11.004-Figure6-1.png", + "caption": "Fig. 6. Constant-breadth cam mechanism with parallel flat-faced double oscillating follower.", + "texts": [ + " 1 corresponds to the function s(\u03b8) of this example. In Cardona and Clos [13] the parametric expression for obtaining the profile of a cam that drives a flat-faced oscillating follower is shown. This expression is: OP \u03b8\u00f0 \u00def g1;2 \u00bc l1 \u00fe l2up sin s \u03b8\u00f0 \u00de\u00f0 \u00de\u2212 l1 cos 2 s \u03b8\u00f0 \u00de\u00f0 \u00de 1\u00fe s\u03b8 \u03b8\u00f0 \u00de\u00f0 \u00de l2up cos s \u03b8\u00f0 \u00de\u00f0 \u00de \u00fe l1 cos s \u03b8\u00f0 \u00de\u00f0 \u00de sin s \u03b8\u00f0 \u00de\u00f0 \u00de 1\u00fe s\u03b8 \u03b8\u00f0 \u00de\u00f0 \u00de 8>< >: 9>= >; 1;2 OP \u03b8\u00f0 \u00def gx;y \u00bc cos \u03b8\u00f0 \u00de sin \u03b8\u00f0 \u00de \u2212 sin \u03b8\u00f0 \u00de cos \u03b8\u00f0 \u00de OP \u03b8\u00f0 \u00def g1;2: \u00f010\u00de In constant-breadth cam mechanisms that drive a parallel flat-faced double oscillating follower (Fig. 6), the displacement function of the follower s(\u03b8), Fig. 7, is made up of one segment which is designed and another that is calculated from it, as shown in [12]. This function must comply with the following expressions: sin s \u03b8\u00f0 \u00de\u00f0 \u00de \u00fe sin s \u03b8\u00fe \u03b4\u00f0 \u00de\u00f0 \u00de \u00bc l2 low\u2212l2 up l1 \u00bc c \u03b4 \u03b8\u00f0 \u00de \u00bc s \u03b8\u00f0 \u00de\u2212s \u03b8\u00fe \u03b4\u00f0 \u00de \u00fe \u03c0: \u00f011\u00de Therefore: \u03b4 \u03b8\u00f0 \u00de \u00bc s \u03b8\u00f0 \u00de\u2212 arcsin c\u2212 sin s \u03b8\u00f0 \u00de\u00f0 \u00de\u00f0 \u00de \u00fe \u03c0 s \u03b8\u00fe \u03b4\u00f0 \u00de \u00bc arcsin c\u2212 sin s \u03b8\u00f0 \u00de\u00f0 \u00de\u00f0 \u00de: \u00f012\u00de The authors, in reference [12], show the flowchart that explains the design process of the global displacement function for these followers", + " The designed segment of the displacement function is defined by the explicit function s(\u03b8) and the calculated segment is defined parametrically by Eq. (12), where (\u03b8+\u03b4) is the rotation angle of the cam and s(\u03b8+\u03b4) is the rotation angle of the follower. In the case of parallel flat-faced double oscillating followers, the displacement function can be freely designed for an interval \u03b8\u2208 [0, \u03b40] in which \u03b40 is given by the initial inclination s(0) of the follower and also by its geometry \u2013 the distance between the center of rotation l1 and the length of the upper l2 up and lower l2 low arms, Fig. 6. \u03b40 is less than \u03c0 rad when s(0)=0 \u2013 the common situation \u2013 as will now be shown. The value of the abscissa \u03b40 of the first point of the calculated segment, and therefore that of the last point of the designed segment, is obtained by making \u03b8=0 in the Eq. (12) [12]. \u03b40 \u00bc s 0\u00f0 \u00de\u2212 arcsin c\u2212 sin s 0\u00f0 \u00de\u00f0 \u00de\u00f0 \u00de \u00fe \u03c0: \u00f013\u00de If s (0)=0, then \u03b40\u2264\u03c0. A special situation arises when c=0 (double symmetric follower with l2 up= l2 low), in which case \u03b40= \u03c0 and, therefore, the interval of \u03b8 where the function is freely designed is [0, \u03c0]", + " (8) and (27): s\u03b8\u03b8 \u03b40\u00f0 \u00de \u00bc sn\u22122sn\u22121 \u00fe sn\u22122\u00f0 \u00den n\u22121\u00f0 \u00de \u03b420 \u2192 sn\u22122 \u00bc \u2212sn \u00fe 2sn\u22121 \u00fe s\u03b8\u03b8 \u03b40\u00f0 \u00de \u03b420 n n\u22121\u00f0 \u00de \u00bc \u2212sn \u00fe 2sn\u22121 \u00fe s \u03b8\u00fe\u03b4\u00f0 \u00de \u03b8\u00fe\u03b4\u00f0 \u00de \u03b40\u00fe \u03b420 n n\u22121\u00f0 \u00de s \u03b8\u00fe\u03b4\u00f0 \u00de \u03b8\u00fe\u03b4\u00f0 \u00de \u03b40\u00fe \u00bc s\u03b8\u03b8 \u03b40\u00fe 1\u00fe s\u03b8 0\u00f0 \u00de\u00f0 \u00de\u2212s\u03b8 \u03b40\u00fe s\u03b8\u03b8 0\u00f0 \u00de 1\u00fe s\u03b8 0\u00f0 \u00de\u2212s\u03b8 \u03b40\u00fe 3 s\u03b8\u03b8 \u03b40\u00fe \u00bc cos2 s0\u00f0 \u00de c\u2212 sin s0\u00f0 \u00de\u00f0 \u00des2\u03b8 0\u00f0 \u00de 1\u2212 c\u2212 sin s0\u00f0 \u00de\u00f0 \u00de2 3=2 \u00fe sin s0\u00f0 \u00des2\u03b8 0\u00f0 \u00de\u2212 cos s0\u00f0 \u00des\u03b8\u03b8 0\u00f0 \u00de 1\u2212 c\u2212 sin s0\u00f0 \u00de\u00f0 \u00de2 1=2 s\u03b8\u03b8 0\u00f0 \u00de \u00bc s2\u22122s1 \u00fe s0\u00f0 \u00den n\u22121\u00f0 \u00de=\u03b420: \u00f034\u00de Fig. 8 shows a numerical example of the design of a displacement function s(\u03b8), and its derivatives to the second order, for a parallel flat-faced double oscillating follower driven by a constant-breadth cam equal to dc=60 mm. The values of the remaining parameters of the mechanism, Fig. 6, are: distance between the centers of rotation l1=80 mm, upper arm l2up=20 mm and lower arm l2low=40 mm. The designed segment of the displacement function is defined by a B\u00e9zier curve with 8 control points, the set of free ordinates being {0 0 0 0,27 0,27} rad. Eq. (13) determines that \u03b40=0, 9196\u03c0 rad and therefore the interval of design is [0 0,9196\u03c0] rad. The ordinates of the last three control points are calculated in the way shown in the previous section, with which the set of B\u00e9zier ordinates is: {si}={0 0 0 0,27 0,27 0,2527 0,2527 0,2527} rad. The first graph of Fig. 8 shows the control points of the B\u00e9zier curve and the displacement function obtained. The second and third graphs show, respectively, the first and second derivatives of the displacement function with regard to the rotation angle \u03b8 of the cam; it can be seen in both that the continuity C2 is fulfilled. Fig. 6 shows the corresponding cam profile for this example. This work exposes the process of direct synthesis of the constant-breadth cam mechanism with a parallel flat-faced double follower, both translating and oscillating, where firstly the displacement function is designed and then the cam profile is obtained from this and the parameters of the mechanism. In this work the viability of using non-parametric B\u00e9zier curves for defining the displacement functions in the said mechanisms has been demonstrated" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001083_978-81-322-2638-3_72-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001083_978-81-322-2638-3_72-Figure1-1.png", + "caption": "Fig. 1 a Proposed antenna top view. b Proposed antenna back view", + "texts": [ + " The complete antenna size is 30 mm3 \u00d7 30 mm3 \u00d7 1.6 mm3. This paper also shows that approach of EBG structure to create tunable notch is better over CSRR. Here authors substantiate their own work presented in [27] with measured radiation patterns and group delay. Following Table 1 provides some useful information about the EBG structures. All the simulation and optimization of the proposed antenna has been done with the Ansoft HFSS 13. Configuration and geometry of suggested antenna is presented in Fig. 1. Fabricated prototype antenna is presented in Fig. 2. This antenna is designed on the FR-4 dielectric material, which has thickness of 1.6 mm, dielectric constant \u03b5r = 4.4 and loss tangent of 0.02. (a) CSRR Antenna Design (WiMAX Notched Band) A CSRR provides filtering characteristic so, we have used a CSRR slot on radiating patch to create notch in WiMAX band. Figure 1a shows the antenna with CSRR slot, and its length is approximately \u03bbg/2. Proposed length of circular split ring resonator can be intended from the Eqs. (1) and (2). Leq \u00bc 2p R1 g \u00f01\u00de fc \u00bc C 2 Leq ffiffiffiffiffiffiffiffiffi er \u00fe 1 2 q \u00f02\u00de where g is 5.5 mm, R1 is 5 mm and C is speed of light. The value of equivalent length Leq varied according to variation in gap \u201cg\u201d which is 5.5 mm. It is optimized to create notch at WiMAX band. All optimized dimensions of the proposed antenna are listed in Table 2" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000978_2004-01-1151-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000978_2004-01-1151-Figure1-1.png", + "caption": "Figure 1. Cross Section of Drive Clutch", + "texts": [ + " Finally, a case study shows this optimal design approach improved clutch design. Starter drive clutch is a one way roller clutch and a key component in a starter motor that is used to crank internal combustion engines. The starter drive clutch transmits torque from an electrical motor to a ring gear mounted on a cranking shaft in an engine thus cranks the engine. The clutch also prevents the whole starter from damage caused by extremely high load and/or extremely high speed applied starter pinion from the engine. Figure 1 shows the cross section of the drive clutch. The clutch consists of one barrel, five springs, five rollers, and one race. The barrel has an internal spline to connect to external spline of an output shaft of a starter motor, and race is part of the starter pinion. Thus, power is transmitted from motor to pinion through the clutch during cranking engine. Sketch (a) In Figure 1 shows the rollers set at their initial position, and right one shows the rollers at their working position. When the starter is not energized, rollers set at their initial position and slightly contact the race and barrel due to the spring force. When the starter motor is energized, an electrical motor drives the barrel, and the race is driven by the barrel though the friction between the barrel/rollers/race. When the pinion engages to the ring gear, RPM of the barrel is greater than RPM of the race thus the rollers are forced to move to their working positions shown in sketch (b) in Figure 1 where the space between the barrel and race is smaller. The high contact stress between race surface and rollers is established due to reduced space. The high contact stress is required so that the clutch can get high torque capacity and transmit enough torque. High torque capacity, however, increases hoop stress in the barrel. Drive slippage and barrel crack are two major failure modes for starter drive failure[1]. Insufficient torque capacity results in the drive slippage while excessive high hoop stress on the clutch barrel ring causes barrel crack" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001120_6.1997-1198-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001120_6.1997-1198-Figure1-1.png", + "caption": "Figure 1. Conventional Wing Model", + "texts": [], + "surrounding_texts": [ + "The concept of an adaptive aircraft wing, i.e., whose shape parameters such as camber, span-wise twist, and thickness can be varied to optimize the wing shape for various flight conditions, has been extensively studied by numerous researchers [1-8]. While the aerodynamic benefits (in terms of increased lift/drag ratios, improved maneuverability, and delayed flow separation) have been analytically and experimentally established, the complexity and weight penalty of the designs and actuation mechanisms have limited their practical implementation. Recent developments in sensors and actuators using smart materials could potentially alleviate the shortcomings of prior designs, leading the way to a more practical \"smart\" adaptive wing which responds to changes in flight and environmental conditions by optimally modifying its shape.\nThis paper presents the results of recent work conducted under a Defense Advanced Research Projects Agency (DARPA) contract entitled \"Smart Structures and Materials Development - Smart Wing\". In particular, development and testing of the smart wing wind tunnel model are presented. Limitations and potential benefits of adaptive wing designs are also discussed, along with recommendations for future work required to develop an operational smart adaptive wing.\n1.0 INTRODUCTION AND BACKGROUND\nSince the dawn of manned flight, aircraft engineers have dreamed of adaptive wings to provide optimal flight performance over a wide range of flight conditions. The terms active and adaptive are used to broadly convey a family of concepts wherein the structure senses the environment and responds actively to optimize performance. For aircraft, concepts include: (1) active feedback control systems for flutter suppression, load alleviation, and improvements in\nride quality; and (2) changing the shape of the wing (to vary camber, span-wise twist or airfoil cross-section) for optimal performance at different flight conditions (takeoff, landing, maneuver, and multiple cruise conditions).\nWhereas active load alleviation systems are quite well developed and installed on several commercial and military aircraft, active flutter suppression systems have yet to be incorporated in operational aircraft. Current experimental efforts are based on actively deploying conventional control surfaces.\nThe theoretical benefits of active control of wing shape are well known and have also been experimentally validated. Two extensive studies in this area are the mission adaptive wing (MAW) and the active flexible wing (AFW) programs [1-5]. The MAW design used a mechanical actuation system to smoothly deploy leading and trailing edge control surfaces which were fully enclosed by flexible wing skins to provide increased efficiency by elimination of discontinuities in the airfoil cross-section. Performance benefits over a conventional fixed camber wing in the subsonic regime were demonstrated in flight tests on a modified F-lll. However, the complexities of the mechanical actuation system and increase in overall weight rendered the design impractical for fleet operations.\nThe AFW concept on the other hand involves reducing the wing flexibility (and hence weight). To improve maneuver performance, the wing was twisted using aerodynamic torque provided by control surface deflections. Aeroelastic performance degradation was offset using active controls. While the anticipated aerodynamic performance benefits were somewhat compromised by the increased drag due to the use of control surfaces to both twist the wing and for normal flight control operations, the concept has sufficient benefits and a detailed flight test is currently being planned [5].", + "The smart wing concept is based on both the AFW and MAW designs and potentially improves the benefits by making judicious use of smart materials and structures technologies. Under an DARPA/WL contract to Northrop Grumman, the smart wing concept is being investigated incorporating new ideas in integrated sensing and actuation systems. Details of the program are discussed below.\n2.0 SMART WING REQUIREMENTS, DESIGN, AND TESTING Under the smart wing program, three key features are being studied: 1) hingeless, smoothly contoured trailing edge (TE) control surfaces, 2) variable wing twist, and 3) realtime pressure distribution data for feedback control. To evaluate the concepts and quantify performance improvements, two 16% scaled models (of a present generation fighter aircraft), one conventional and the other incorporating the above features (Figures 1 and 2), have been fabricated and tested in a wind tunnel to quantify performance benefits of the smart wing concept. Prior to undertaking the design, actuation requirements for the smart wing were established.\n2.1 Requirements Details of the requirements analysis performed in the program are provided in References 8 and 9. Figure 3 shows the actuation rates needed for various flight operations. Figure 4 shows calculated (nominal) values of torque at wing mid-span and tip to achieve 2 and 5 degrees of twist for a full-scale aircraft and scaled models. (The torque requirements increase essentially as the fourth power of the geometric scaling factor - the values shown in the figure are slightly different because of differences in the materials used.) While it is feasible to achieve the torque requirements for the models, it is obvious that meeting scaling requirements will be a significant challenge to transition this technology to a full-scale aircraft. This is discussed further in Section 3.\n2.2 Design Wing Twist: Several design concepts (Figure 5) were considered for twisting the wing for the wind tunnel models. Initial trade studies indicated that the integrated torque box concept was structurally most efficient. However, on further examination, the design presented severe manufacturing difficulties and appears to be somewhat impractical. Hence the shape memory alloy (SMA) torque tube actuation was chosen and a design with two concentric tubes as shown in Figure 5A was implemented. This technique functioned well in the tunnel, but because the final wind tunnel model was significantly stiffer than the scaled model (primarily due to escalation of wing skin and spar web thickness from the original scaled values to prevent local panel buckling), maximum wing tip twist of only about 1.25 degrees was realized. If the stiffness were scaled exactly, 3 to 5 degrees of twist at the wing tip could easily have been achieved. Further details of the torque tube design are presented in Reference 10 and 11.", + "Adaptive Control Surfaces: The aerodynamic benefits of contoured hingeless surfaces are well known [1-3]. To implement these types of control surfaces, SMA based actuation systems are ideal because of their high force and high strain capabilities [12, 13]. Figure 6 shows a schematic of the adaptive control surfaces with embedded SMA wires in top and bottom face sheets which provide two-way \"antagonistic\" actuation. Figure 7 shows the final system used for the wind tunnel model. Approximately forty 20 mil diameter wires were used to obtain the equivalent of ten degrees of rotation. Because of the complex thermo-mechanical behavior of the SMA wires, i t was essential to incorporate sensors to determine the true position of the control surfaces. The most suitable sensors were fiber-optic sensors, and a suite of extrinsic FabryPerot interferometric (EFPI) strain sensors were embedded in the control surfaces and calibrated to provide an accurate measure of control surface actuation. This information was used for feedback to command, achieve and maintain a desired deflection.\nA modified version of the EFPI strain sensor was also developed for pressure sensing; the design, testing and performance of this sensor is discussed in detail in Reference 14.\n2.3 Fabrication and Assembly\nThe basic planform of the wind tunnel models consists of four spars and three ribs and is shown in Figure 8.\nDetails of the SMA torque tube fabrication, training and integration into the structure are discussed in Reference 11; a few pertinent issues are summarized here.\nThe SMA torque tubes were manufactured from a hot rolled Unimet metal rod obtained from Special Metals company. Due to the inherent difficulty in machining TiNi, the rods were cut into 4 inch lengths using EIcetrospark Discharge Machining (EDM), followed by gun drilling to produce a rough tube, and then further EDM processed to obtain the final tube.\nBecause of a lack of accurate knowledge of the complex thermo-mechanical behavior of SMA torque tubes, their training was a trial and error process, requiring many iterations.\nConnection of the torque tubes to the wing structure to ensure maximum torque transfer as well as provide easy assembly required an innovative design. Figure 9 shows a" + ] + }, + { + "image_filename": "designv6_24_0002571_isem.2018.8442835-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002571_isem.2018.8442835-Figure2-1.png", + "caption": "Fig. 2. Numerical model in FLUX 3D", + "texts": [], + "surrounding_texts": [ + "978-1-5386-5210-7/18/$31.00 \u00a92018 European Union\nKeywords\u2014 switched reluctance motor; electric motor; electric vehicle; finite element method\nI. INTRODUCTION Over the past years an increase in the demands regarding environmental protection has been notified. One of the most important tasks imposed on new combustion engine vehicles is reduction of harmful exhaust gases [1]. Electric motors offer an alternative to internal combustion engines. Electrical vehicles can be driven by energy from renewable sources. The use of an electric motor eliminates exhaust emissions and reduces noise [2,3,4]. Despite a number of advantages, electric vehicles offer a limited range on a single battery charge [5]. The solution to this problem could be inter alia the state support in creating modern infrastructure of stations designed to charge electric vehicles. According to the data from the Alternative Fuels Observatory in Poland, there are currently around 150 charging stations for electric vehicles [6]. Depending on the size of the electric vehicle, it is necessary to adapt the powertrain to particular design requirements. In electric or hybrid vehicles, there is a tendency to use propulsion of the highest torque density [7]. At present, the motors with permanent magnets are the most popular featuring\ngood exploitation properties. However, they require considerable investment due to high prices of rare earth elements [4,7,8]. Other machines used for vehicle propulsion are induction motors, synchronous motors and switched reluctance motors (SRM).\nSRM motors are characterized by high torque density, a wide range of speed regulation, operation characteristics similar to the serial motor and have a simple design [9, 10]. These types of motors can be used in vehicles where assembling space is not a critical condition on account of their volume and mass. Definitely the SRM motor in electric vehicles is most often mounted directly on the drive wheels [1,8]. The paper presents the SRM motor with geometry of the stator modified by reducing the number of power poles and diminishing the stator radius in the modification. The design changes result in reduction of the machine volume and mass. Thanks to the applied characteristic geometry of the stator, it is possible to install the motor in the bicycle's support. The design assumptions do not exclude the application of a gear between the motor shaft and the shaft applied in the drive of the bicycle chain. In addition, the presented design of the machine should meet the requirements specified for drives applied in electric bikes. Assuming that the propulsion of the electric vehicle should not propel the bike more than 25 kms per hour while using 28-inch wheel rims, the speed of the electric motor will not exceed 200 rpm. Moreover, considering that the total weight of the bicycle with the cyclist is 100kg, the torque delivered to the wheel should be at least 3 N\u00b7m to ensure adequate ride comfort [11, 12].", + "978-1-5386-5210-7/18/$31.00 \u00a92018 European Union\nII. MODIFIED DESIGN OF SRM MOTOR The proposed design of the SRM motor is presented in Figure 1. Preservation of the original assembling space can be combined with an increase of the rotor radius resulting in an increase of the torque value.\nTable 1 contains a summary of selected parameters of the analyzed motor.\nThe presented modification of SRM motor enables to install it in vehicles and devices with limited assembling space, including mounting it in the lowest construction point\nof light vehicles, bicycles in particular. The use of the armature made of ferromagnetic material forming a closed magnetic circuit has a beneficial effect on the operation of the motor. In order to avoid losses and too much saturation in the magnetic circuit of the stator, the thickness of the armature is similar to the width of the stator yoke.\nIII. ANALYSIS OF THE MAGNETIC FIELD For calculating electromagnetic parameters of the tested motor two and three-dimensional models were developed which include the nonlinear magnetization characteristics and a constant current density across the entire coil section. Basing on the models, a number of magnetic field calculations were made and the characteristics of the electromagnetic torque (Te) were determined depending on the rotor angle position ( ) for the applied current values in the phase given. The electromagnetic torque for the 2D models was determined using the Maxwell Stress Tensor and Virtual Work method for the 3D models. In addition, the calculations provided the torque density (1) as well as ripple (2) [13, 14, 15]\nV T avT d = (1)\nWhere Td \u2013 torque density, Tav\u2013 mean value of electromagnetic torque, V \u2013 motor volume.\n%100 2\nminmax \u22c5 \u2212\n= avT\nTT \u03b5 (2)\nwhere Tmax\u2013 max torque value, Tmin\u2013 min torque value.\nWith the aim of reducing the numerical costs required for analysis of the 3D model, the symmetry in the XY plane of the analyzed machine was used and the constructed model was limited to half of the machine volume.\nThe conducted preliminary calculations of electromagnetic field show that the armature plays the role of the magnetic pole without winding. The lack of armature breaks the magnetic circuit resulting in the drop of the generated torque. A fundamental aspect of the stator design is the length of the armature, which should not exceed the stator diameter.\nFig. 3 presents the distribution of the magnetic flux density for two positions of the rotor in relation to the stator.", + "978-1-5386-5210-7/18/$31.00 \u00a92018 European Union\nFig. 4 presents the relation of the electromagnetic torque for phase C with power supply in the range 0 \u2013 35A (DC) gained on the basis of the 3D model.\nThe differences in the results of calculated electromagnetic torque gained by application of 2D and 3D models do not exceed 3.3%. The maximum torques determined on the basis of the examined models amount to Te_max2D=9.52N\u00b7m, Te_max3D=9.21N\u00b7m. The differences in the values of torque for the 2D and 3D models result from the machine construction, namely the proportion of the motor diameter to its active length, which in this case accounts for 3/1.\nIncluding the simplest control algorithm of the motor, which involves the initiation of the phases in the sequence A C B A (Control 1) Fig. 1 shows the determined cyclicality of the electromagnetic torque for the two of analyzed models.\nRaising the turn-on angle of the particular phases by 5.6\u00b0, and thereby modifying the control algorithm to the following sequence (Fig.1): A AC C CB B BA (Control 2), an increase of mean torque value is achieved with simultaneous reduction of ripples by 10%.The cyclicality of the torque is presented in Fig. 6.\nThe summary of the results can be found in Table II.\nTABLE II. COMPARISON OF CONTROL ALGORITHMS\nTav [N\u00b7m] Td [kN\u00b7m/m3] [%]\n2D 3D 2D 3D 2D 3D\nControl 1 8.13 7.72 10.6 10.1 32 33\nControl 2 9.07 8.64 11.8 11.3 22 23 In the next stage the authors determined dependence of the incremental self-inductance vs. the rotor angle position. The mathematical relation defining this problem is as follows:" + ] + }, + { + "image_filename": "designv6_24_0003647_j.msea.2014.02.078-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003647_j.msea.2014.02.078-Figure8-1.png", + "caption": "Fig. 8. Surface plot showing the length fraction of \u03a33 boundaries following a one hour anneal at 1025 1C of the IN600 samples deformed to various strains at different temperatures.", + "texts": [ + " Deformation of the samples to 11% strain followed by annealing at 1025 1C did not result in any recrystallization phenomena and strain annealing was responsible for a significant increase in both the \u03a33 length fraction and grain size, Figs. 6 and 7. Although increasing deformation temperature did result in an increase in the average grain size following annealing, Fig. 4a, d, g and j, the extent of grain growth in the deformed and annealed samples was significantly smaller when compared to an as-received, undeformed sample of IN600 that was subjected to a nominally identical annealing treatment, Fig. 8. As the deformation temperatures increased, increasing levels of dislocation recovery occurred and led to smaller degrees of intragranular misorientation occurring within the microstructures of the samples. Both grain growth and the degree of dislocation recovery were found to effect the formation of S3 boundaries in this set of samples. Compared to the average total length of \u03a33 boundaries measured in the as-received or undeformed samples, a modest increase in the overall length of \u03a33 boundaries was measured following annealing of the samples deformed to 11% strain at 25 1C, 538 1C and 760 1C", + " Unlike the highly strained samples deformed at 25 1C and 538 1C where static recrystallization results in a strain or dislocation free microstructure where the reduction in surface or interfacial energies drive grain growth, dynamic recrystallization produced a refined grain structure containing a sufficiently high strain level or dislocation density that was able to enhance the formation of \u03a33 boundaries during annealing. Although the overall length of the \u03a33 boundaries measured in the sample deformed to 80% at 982 1C was nominally identical to that measured in the same sample following annealing, the average grain size increased from 14.5 \u03bcm to 24.8 \u03bcm as a result of the annealing, Fig. 7. The larger grain size also corresponds to a significant reduction in the overall length of random grain boundaries and results in a net overall increase in the length ratio of \u03a33 boundaries, Fig. 8. Hence, strain-annealing and dislocation grain boundary interactions that occur within fully, dynamically recrystallized microstructures do appear to increase the overall length ratio of \u03a33 boundaries at large magnitudes of strain. A deformation temperature of 982 1C corresponds to a homologous temperature, Tm, of 0.77 for IN600. Considering that the microstructure of IN600 consists of a solid solution strengthened, single phase austenitic structure, dynamic recovery mechanisms can dominate and accommodate high levels of strain during high temperature deformation [36,37]", + " The subsequent annealing of the specimens deformed at 982 1C was found to induce some additional grain growth and result in a corresponding reduction of the random grain boundary length, but did not serve to generate additional \u03a33 boundary length. Based on these observations, it appears that under deformation conditions where dynamic recovery dominates, \u03a33 boundaries are neither created nor destroyed. This permits samples deformed at 982 1C to 11% strain followed by an annealing cycle to exhibit a \u03a33 length fraction of 81% at a grain size of 52.2 \u03bcm. When compared to the grain boundary character of undeformed samples of IN600 that were annealed using the same thermal cycle, Fig. 8 (79% \u03a33 length fraction with a grain size of 69.9 \u03bcm), mechanisms associated with dynamic recovery at 982 1C were responsible for increasing the \u03a33 length fractions while maintaining a modestly reduced grain size. The formation of denser subgrain networks and following deformation at strains between 25% and 80% further restricts the grain boundary mobilities during annealing as the average grain sizes in these samples are on the order of 22.2\u201325.8 \u03bcm. Despite the smaller average grain sizes, \u03a33 length fractions between 59% and 63% were obtained in these samples" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002550_tmag.2012.2226468-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002550_tmag.2012.2226468-Figure3-1.png", + "caption": "Fig. 3. Magnetic flux distribution of IM: (a) healthy and (b) two broken bars IM [26].", + "texts": [ + " By increasing the NBB, the nonlinearity of the motor behavior rises and the rate of variations of the amplitude of both right and left sides will be slower. Magnetic field in the rotor broken bar region leads to the magnetic saturation. Since this saturation does not cover the whole machine and occurs in particular regions of rotor and stator cores and the air gap, it is called local saturation. Local saturation has several effects and the most important one is the impact on the location of the hot spot within the core. In addition, saturation increases the core losses. As shown in Fig. 3, increase of the magnetic flux density happens in the rotor core, around the broken bar, stator core and the air gap [26]. The air gap flux is proportional with the current and therefore, induced EMF is proportional with the . Therefore, in order to include the local saturation, attention must be paid to the rotor bars current. By comparing the amplitude of the bar current with the highest current in faulty case with the amplitude of the same bar in the healthy case, coefficients are obtained to include local saturation in the following model: (5) (6) where is the faulty rotor bar current, is the current of the same bar in the healthy case, is the local saturation factor and is the induced EMF", + " Considering (17), (16) is expressed as follows: (18) Rewriting (8) as a discrete equation, EMF is denoted by the following: (19) By substituting (19) in (18), the frequency of the equal sinusoidal flux density is derived as follows: (20) where is the sampling frequency computed by: (21) Replacing (20) in (7), the specific energy loss due to bars breakage is derived as: (22) If the remagnetization is repeated with the period , the power losses due to rotor broken bars are estimated as follows: (23) Fig. 9(a) depicts core losses of the IM versus NBB.According to Fig. 9(a), even one broken bar increases core losses considerably. Because, broken bar distorts the flux density distribution in IMs (see Fig. 3) and distortion of flux density magnifies harmonic components in the flux density profile. Moreover, it is seen that the fault extension raise core losses. However, incremental rate of the loss increase in two, three and four broken bars is less than that of the faulty case (one broken bar). Fur- thermore, it is seen that the incremental rate of the core losses in the IM with one broken bar in the DTC mode is less than that of the line-start and CV/F modes. It is due to competent control of flux density in the DTC mode in comparison with other modes" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002082_mop.10661-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002082_mop.10661-Figure2-1.png", + "caption": "Figure 2 (a) Schematic drawing showing the fabrication of the antenna; (b) side view and top view of the basic cone; (c) schematic model of the unwrapped log-periodic antenna ( o 30\u00b0); (d) modulation functions for various tooth shapes", + "texts": [ + "igure 2 shows the variation of the bandwidth, BW, as a function of r3. The results are obtained using the following structure parameters: A 25 , R 20 , t1 5 x, t2 3 x, r1 2.53, and r2 1.1 (Foam). The case of r3 1.0 represents a truncated dielectric layers; they do not cover the entire ground plane. The value of R 20 is used after studying the effect of changing R on the center frequency 4.5 GHz. We found that increasing R more than R 20 has a very little effect on the center frequency of the structure. However, using smaller values of R has the effect of shifting the center frequency upward or downward, depending on the value of r1 and r2. The bandwidth (2:1 VSWR or 10 dB) is maximum at r3 7.0 with a value of 31.5%. The case of truncated dielectric layers ( r3 1.0) gives a BW 16.5%, and the conventional structure gives a BW 21.8%. The center resonance frequency is 4.5 GHz. Also, Figure 2 shows that the BW increases for r3 15, but for high values of r3 the antenna efficiency decreases, due to the high coupling of energy to substrate and surface waves. For r3 from 1.0 to 7.0, the BW increases rapidly with r3. The value of r3 7.0 is chosen as the optimum value that gives the largest bandwidth for this structure. Also shown in the same figure are the curve-fit variations of the data using a 4th degree polynomial. The calculated S11 as a function of frequency is shown in Figure 3. These curves are obtained using: R 20 , t1 3 x, t2 4 x", + " Like all frequencyindependent antennas, the geometry of this antenna is mainly described by angles, and lengths are introduced to specify the smallest and largest dimensions of the (finite) antenna. In section 2, the antenna geometry is presented, including various antenna arm designs, and the modeling of the antenna using the finite-difference time-domain (FDTD) method is briefly described. Results for the impedance, gain, and pattern are then presented in section 3. The two-arm, conical log-periodic antenna, shown in Figure 1, consists of two metallic strips or arms placed on the surface of a cone. A practical approach for constructing this antenna is shown schematically in Figure 2(a). The metallic arms are formed on a MICROWAVE AND OPTICAL TECHNOLOGY LETTERS / Vol. 36, No. 1, January 5 2003 29 planar, flexible circuit board through a wet chemical etch. The circuit board is then wrapped around a conical mandrel, and soldered or taped along its seam. The basic cone, shown in Figure 2(b), is characterized by the half angle of the cone o and the diameters d and D that limit the extent of the antenna at the small and large ends. In Figure 2(c), the antenna arm geometry is shown on the planar sector obtained by unwrapping the conical antenna. This drawing is for the angle o 30\u00b0. Simple geometry shows that any angular width on the planar sheet translates into an angular width on the conical structure using the relation sin o. (1) Hence, as shown in Figure 2(c), the total angular width of the unwrapped cone is 180\u00b0 when o 30\u00b0. For the design shown in Figure 2(c) the notches and teeth of constant angular width 2 max are inscribed in the circular sector of angular width . Based on the principles of log-periodic antennas, the dimensions of these notches/teeth vary periodically, in particular, the radii measured from the apex to two adjacent edges of an arm are related to each other by ri 1 ri , (2) where is the geometric ratio for the log-periodic antenna. On the cone, the second arm is symmetrically located to the first (diametrically opposite), and at any height z the angular width of an antenna arm is 2 max. All of the designs presented here are for the case 2 max 90\u00b0, which is shown in Figure 2(c) for the unwrapped case* and in the bottom of Figure 2(b) for a planar cut through the conical antenna. For this choice of angles, the metallic arms are identical in size and shape to the open regions. The main objective of this work is to study the effect of different tooth shapes on antenna performance. To describe the modulation of the teeth, the angle (r) is introduced (see Fig. 2(c)). Since we assume that the minimum and maximum values of the modulation angle (r) are independent of the radius r, this angle is measured from the center of the sector (with angular width 2 max) that makes up the tooth geometry, so that (ri) max. The modulation angle (r) for the rectangular tooth geometry is rect r max i 1 N r ri ri 1 ri 1 i 1, (3) for the triangular tooth geometry tri r max i 1 N r ri ri 1 ri 1 i 1, (4) and for the sinusoidal tooth geometry sin r maxcos log r/r1 /log . (5) In Eqs" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001770_tim.2014.2364699-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001770_tim.2014.2364699-Figure3-1.png", + "caption": "Fig. 3. Three-phase inductance vectors and the rotating coordinate transformation.", + "texts": [ + " The three-phase current slope difference and the full-cycle inductance profiles of the tested 12/8 SRM in an electrical cycle are shown in Fig. 2. It needs to be noted that Fig. 2 is a diagrammatic sketch, the inductance L and phase current slope difference I are not exactly the sinusoid signals. The three-phase inductance of the tested 12/8 structure SRM can be represented by three positive vectors with fixed phase difference of 120\u00b0 electrical angle, and the norm of which is the magnitude of the phase inductance. The phase inductance magnitude of each phase will change simultaneously according to the rotor position. As shown in Fig. 3, the three-phase inductance vectors La , Lb , and Lc are relatively static in the a \u2212 b \u2212 c plane coordinate system, where La 0, Lb 2\u03c0/3, and Lc 4\u03c0/3 represent the relative positions of each phase inductance vectors respectively. Rotating the a \u2212 b \u2212 c coordinate system in counterclockwise of \u03b4, the new a\u03b4 \u2212 b\u03b4 \u2212 c\u03b4 coordinate system could be constructed. As shown in Fig. 3, the sum projection value of the three-phase inductance La , Lb, and Lc in the a\u03b4-axis of the new coordinate system is defined as La\u03b4 . Similarly, Lb\u03b4 and Lc\u03b4 can be defined, respectively. The relationship between the two coordinate systems can be represented by \u23a1 \u23a3 La\u03b4 Lb\u03b4 Lc\u03b4 \u23a4 \u23a6 = \u23a1 \u23a2\u23a2\u23a2\u23a2\u23a2\u23a3 cos \u03b4 cos (2\u03c0 3 \u2212 \u03b4 ) cos (2\u03c0 3 + \u03b4 ) cos (2\u03c0 3 + \u03b4 ) cos \u03b4 cos (2\u03c0 3 \u2212 \u03b4 ) cos (2\u03c0 3 \u2212 \u03b4 ) cos (2\u03c0 3 + \u03b4 ) cos \u03b4 \u23a4 \u23a5\u23a5\u23a5\u23a5\u23a5\u23a6 \u00d7 \u23a1 \u23a3 La Lb Lc \u23a4 \u23a6. (2) For cos \u03b4 + cos ( 2\u03c0 3 \u2212 \u03b4 ) + cos ( 2\u03c0 3 + \u03b4 ) = 0 (3) it can be obtained that La\u03b4 + Lb\u03b4 + Lc\u03b4 = 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003121_12.826223-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003121_12.826223-Figure8-1.png", + "caption": "Figure 8: (a) NIRCam FEM (b) Strut to Bench Bracket FEM Detail", + "texts": [], + "surrounding_texts": [ + "A necessary but exhaustive structural, dynamic, and thermal analysis was conducted to support the design. Much of the analysis was done concurrent with the design process which helped to mature the design quickly while maintaining its technical integrity. For example, there were a number of iterations for the pad support designs so that all degrees of freedom were properly constrained; yet maintaining the ability to assemble/disassemble the pads. In addition to the concurrent analysis and design, once the design had matured a more traditional approach to evaluate the structural and thermal integrity was taken." + ] + }, + { + "image_filename": "designv6_24_0000118_2011-01-0962-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000118_2011-01-0962-Figure1-1.png", + "caption": "Figure 1. Schematics of the two-axle rigid vehicle model", + "texts": [ + " The conclusion of the paper is the summary of the effect and the significance of each source of non-linearity on the overall vehicle response, for an advanced design of vehicle dynamic performance, handling and control. NON-LINEAR VEHICLE MODELS Rigid Vehicle Model Numerous models, from fairly simple to increasingly comprehensive, have been adopted for rigid vehicle dynamics analysis [10],[11],[12],[13],[14],[15],[16]. In this study, an eight Degree-Of-Freedom (DOF) non-linear vehicle model has been employed. This model is characterized by the maximum possible simplicity in relation to the non-linear effects it should evaluate. Figure 1 depicts the twodimensional schematics of the rigid vehicle model. In this formulation, vehicle sprung mass has four DOFs, i.e. two translational (longitudinal and lateral displacement) and two rotational motions (yaw and roll). The other four DOFs are wheel rotations. It has been assumed that vehicle characteristics are symmetrical along the center line of vehicle. (1) (2) (3) (4) where i = 1, 2, 3 \u2026 n. n is the total number of wheels of the vehicle The longitudinal and lateral tire forces, denoted by Fx and Fy respectively, are non-linear functions of tire slip angles, slip ratios, camber angles and vertical loads", + " speed characteristics (properly converted into ) have been tested in simulations for comparisons and analyses, including linear, non-linear progressive and non-linear regressive behaviors. The generic model described by the above equations considers all the axles to be steered, but for simplicity this analysis is assuming only the wheels of the front axle to be steered. The lateral (transversal) load transfer and its distribution between the axles of the vehicle is one of the crucial components for the computation of non-linear lateral tire force. From Figure 1 the difference in the load shift from one side of an axle to the other side may be represented by the following equation: (5) where nj is the total number of tires on the jth axle of the vehicle, yj is the track width of the jth axle. The longitudinal load transfer, due to vehicle forward acceleration and aerodynamic drag (here applied to the vehicle center of gravity) is modeled using the following equation in case of a 2-axle vehicle. (6) SAE Int. J. Passeng. Cars - Mech. Syst. | Volume 4 | Issue 1 723 where xji is the longitudinal distance of centre of the ith tire of the jth axle from CG, nj is the total number of tires of the jth axle and L is the wheelbase of the vehicle. For rigid vehicles characterized by more than 2 axles (or articulated vehicles), the equations of the longitudinal load transfer have to consider the suspension stiffness and damping distribution along the vehicle. From the vehicle geometry in Figure 1, a general non-linear tire slip angle equation is constructed as described in Equation (7), where the following contributions are incorporated: \u2022 static toe angle; \u2022 steering angle as imposed by steering system kinematics (different for each wheel); \u2022 toe angle variation as a function of vehicle roll motion; \u2022 toe angle variation induced by half-track variation as a function of vehicle roll motion; \u2022 toe angle variation due to the suspension compliance induced by tire lateral force. (7) where i = 1, 2, 3 \u2026 n" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002713_9780470611111.ch3-Figure3.1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002713_9780470611111.ch3-Figure3.1-1.png", + "caption": "Figure 3.1. Example of a spaceborne SAR: the RADARSAT satellite \u00a9 CSA", + "texts": [], + "surrounding_texts": [ + "The main characteristic of regular Earth coverage by imaging satellites depends on their particular type of orbit (orbital plane inclination, phasing, cycle, etc.). This section provides complementary information on some orbitography notions that can improve our understanding of several essential parameters of SAR systems and our approach to some satellite location issues. Several acquisition elements playing a Chapter written by Jean-Marie NICOLAS and Fr\u00e9d\u00e9ric ADRAGNA. Processing of Synthetic Aperture Radar Image2 Edited by Henri Maitre Copyright 0 2008, ISTE Ltd major role, for example, in the implementation of relief reconstruction techniques are also specified. 3.1.1. Remote sensing satellite orbits Among the numerous parameters characterizing orbits we will only discuss those that are relevant to Earth observation satellites. Such satellites must first of all have near-polar orbits in order to ensure their coverage is as fully global as can be. As for altitude, most of them operate at around 800 km. They are often sun-synchronous, although this requirement is mainly useful for optical remote sensing satellites. 3.1.1.1. Inclination Orbital inclination is the angle between the equatorial plane and the orbital plane of the satellite in ascending direction (i.e., when it is headed for the North Pole; see Figure 3.2). For sun-synchronous satellites, this angle has to be higher than 90 , as will be shown below. It is actually 98.5 in ERS and 98.6 in RADARSAT. 3.1.1.2. Period The velocity of a satellite in a circular orbit depends only on its altitude H: .sat E E g V R R H [3.1] Its orbital period, which is the time it takes to travel one full orbit, depends on its altitude H and the Earth\u2019s radius RE as follows: 2 3 2 E E sat gR HR T Regarding remote sensing satellites, the altitude frequently ranges from 500 to 800 km: they have orbital periods of 100 min, approximately 14 orbits per day. 3.1.1.3. Eccentricity The ellipticity of an orbit is characterized by its eccentricity: 2 22 a ba e , where a and b represent semi-major and semi-minor axes of the orbit. Remote sensing satellites often have quite circular orbits (i.e., their eccentricity is very low, e.g., for ERS, e = 1.165\u00b710-3). 3.1.1.4. Sun-synchronism A satellite is said to be sun-synchronous (or heliosynchronous) when it is synchronized with the Sun, i.e., when there is a constant angle between its orbital plane and the Earth-Sun axis. In these conditions, the satellite benefits from nearly identical illumination for every scene that it observes (except for latitude and seasonal variations). While being sun-synchronous has no relevance for radar satellites, they are often so. As a result, they can use the same platforms (solar panel orientation) as their optical counterparts and have a simpler operating pattern. ERS 1&2 satellites for example, are referred to as \u201c10:30\u201322:30\u201d like the SPOT satellites whose platform they share (see Figure 3.2). This means they cross the equator (descending node) at 10:30 in the morning (local sun time) and again (ascending node) at 22:30. They go down on the sunlit side of the Earth and go back up on the shadowed side (during the night). RADARSAT is a \u201c06:00\u201318:00\u201d satellite, meaning that it does not enjoy good ground illumination by the Sun, which it does not really need. Still, its solar panels are constantly illuminated, which allows it to perform acquisitions on both its ascending and descending orbits without using too much battery power. Sun-synchronism is made possible by an anomaly of the Earth potential. If the Earth were a perfectly spherical homogenous body, orbital planes would be timeinvariant (with respect to a reference point in the sky). The Earth\u2019s flattening at the poles is, among other things, responsible for a rotation of this plane, known as precession. By choosing orbital parameters carefully, we can make this plane spin at a rate of one turn per year to make up for the Earth\u2019s rotation around the Sun. For a circular orbit, this implies the following relation between altitude H and orbital inclination i (which has to be wider than 90\u00b0 for the cosine to be negative): 3609.97 cos 0.985 365.24 T T R i R H 3.1.1.5. Cycle For the sake of comparison, it is desirable that successive images of a same region are acquired at the same look angle. We therefore want the satellite to pass over the same point on the Earth\u2019s surface once it has completed an orbital cycle involving an equal integer number of days (same Earth\u2019s position) and equal integer number of orbits (same satellite position). Such orbits are described as phased or geo-synchronized. ERS and RADARSAT cycles are 35 days and 24 days long, respectively. Phased orbits often have a sub-cycle, which is the time it takes the satellite to pass again (though not exactly with the same look angle) near a given point. The RADARSAT platform, for example, travels (14 + 7/24) orbits per day. In seven days, it makes (7 14 + 7/24) revolutions, a nearly whole number, and is thus very close to the orbit it followed on the first day. In 24 days, it completes an integer number of orbits and begins another orbital cycle which is entirely identical to the previous one. 3.1.1.6. Phasing Sun-synchronous orbits are often phased (existence of a cycle), which leads to some confusion between these two attributes. Still, not all sun-synchronous orbits are phased, nor are all phased orbits sun-synchronous. From a mathematical perspective, phasing implies H, but not i. 3.1.1.7. Orbital drift and correction At an altitude of 800 km, the atmosphere is rarefied, yet friction continues to exist, albeit rather low. It leads to loss of altitude and, according to equation [3.1], acceleration. The result is an eastward drift: the Earth always spins at the same speed, so the satellite arrives in the East a little earlier than planned. The ERS, for example, loses around 1 m daily. In order to keep orbital properties unchanged, the satellite needs to be repositioned to its nominal altitude at regular intervals. Adjustment operations are ground-controlled and involve the use of thrusters mounted on the platform and onboard propellant." + ] + }, + { + "image_filename": "designv6_24_0001839_2015-01-1363-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001839_2015-01-1363-Figure1-1.png", + "caption": "Figure 1. A typical park lock system", + "texts": [ + " Among them, the park gear is positioned on the transmission output shaft, the pawl is installed on a pawl shaft that is connected to the transmission casing, and the detent lever mechanism connects the park pawl to the shift lever via a park rod. Whenever the driver shifts the lever to the \u201cPark\u201d position, the detent lever mechanism activates the pawl actuation motion. At higher speeds, the pawl ratchets or slips between gear teeth; while at lower speeds, it locks the park gear and thus immobilizes the vehicle. Depending upon the specific vehicle, different type of park lock systems are available. A typical park lock system is shown in Figure 1. In the actual model, there is an axial spring (known as actuator spring as shown in the left inset of Figure 1) between the park rod and the actuator and there is a torsional spring (known as Pawl Return Spring as shown in the right inset of Figure 1) between the pawl and the pawl shaft that goes through the pawl hole. The axial spring is fixed at park rod stopper at one end and the other end sits on top of the actuator. The torsional spring sits on the pawl shaft and controls the pawl actuation motion (up and down motion). These springs hold the pawl out of the park (locking) position when the shift lever is not in the \u201cPark\u201d position. The design of the park pawl system depends on the transition point of the drivetrain ratcheting to locking sequence", + " Additionally, the accuracy of such a manual iteration cannot be guaranteed. With the change of any system parameter, an analyst needs to reestablish the engagement speed. The current objective of this project is to automate the engagement speed calculation procedure. The software Isight enables automation of such a procedure, and hence is used here to accomplish this objective. For the current study, Abaqus Explicit is used as the solver. A typical park lock model is selected for engagement speed automation study as shown in Figure 1. In the current model, vehicle mass, driveline stiffness, driveline damping, linear and torsional spring details, and other information are embedded. The details of such data and their calculation methods are out of scope of the current study. However, all of the parameters affect the engagement speed calculation. Each of the components shown in Figure 1 is treated as a rigid body. In standard practice, engagement speed is calculated using multi body dynamics and then the model is fine-tuned using deformable analysis. Though the current study is based on the rigid body analysis only, the proposed automation procedure is also applicable to the deformable FE analysis. The engagement speed from rigid body analysis slightly differs from the deformable FE analysis. So the result from the rigid body analysis is used as the starting point for the deformable analysis with adjustment no more than 10%. Each of the parts shown in Figure 1 is meshed using rigid elements. System level information such as spring stiffness, driveline damping and stiffness, vehicle mass, and etc. are treated as 1-D features (elements and connectors). Penalty type contact is used for the current study. In addition, boundaries, detent rotation, and vehicle and gear initial rotational velocities are used as loads. Note that for all cases, vehicle and gear rotational velocities are considered as equal at the start of the analysis. To save time and space, the explicit dynamic analysis is run only for 0", + " ADAMS based model can also be implemented by modifying the Isight components. The Task component can be changed to other design exploration method, such as DOE, optimization and six sigma is used in other studies where the linear and torsional spring stiffness values are varied to study their effect on the park lock system. In addition, Isight has elegant data processing capabilities and some of them are used here to derive conclusions, as discussed in the subsequent section. Using the model shown in Figure 1, and incorporating it inside Isight, using T = 0.001, and \u03b5= 0.0099, analyses were run. The analysis required 2.3 hours to complete eleven successful iterations as shown in Table 1. Note that iteration 10 gives the last successful result, and this should be regarded as the final result. This value is exactly the same as the value obtained via manual iterations. Another way to verify the result is to look at the rotational speed plots at the gear center, similar to Figure 2. All the gear center rotational velocity data were stored automatically while performing Isight iterations" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001717_j.biosystemseng.2008.02.010-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001717_j.biosystemseng.2008.02.010-Figure5-1.png", + "caption": "Fig. 5 \u2013 Schematic diagram of the method used for measuring impingement force on the surface of halved fruit.", + "texts": [ + " After each extraction test was completed, the extracted arils were sorted into physically damaged and undamaged arils using visual inspection and the percentage of physically damaged arils was calculated. Arils with visible damage to their flesh were considered as damaged. An experimental and analytical investigation was conducted to determine the force generated by an air-jet impinging normally on the surface of fruit. The effect of nozzle diameter (at 2.5, 3, 3.5, and 4.5 mm) and supply air pressure (at 300, 500, 700, and 800 kPa) on air-jet impingement force was assessed. Fig. 5 shows a schematic diagram of the apparatus used to measure the forces generated by the air-jets on the surface of halved pomegranate fruits. The apparatus consisted of a flat smooth plate attached to a balance, a regulated air compressor, nozzles, a pressure regulator, and a pressure gauge. The momentum of each flow was determined by measuring the force normal to the fruit surface held near the exit of the nozzle. Chamber pressures were measured with a pressure gauge for all the tests. Following the methods used by Janos and Hoffman (1968) and Crafton et al" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001074_iccv.1995.466882-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001074_iccv.1995.466882-Figure5-1.png", + "caption": "Figure 5: (a) Occlusion graph for the mechanism in Fig. 4, and (b) the visibility order produced by topologically sorting the graph.", + "texts": [ + "2 Occlusion Graphs The occlusion relations for a multi-body system with no ambiguities can be represented by a directed occlusion graph. The graph is a pair ( V , E ) , where the vertex set V contains all of the bodies. To construct the edge set, E , consider all pairs t,y E V . Since there are no occlusion ambiguities, one of 2 y, x + y, or y b 3: must be true. In the first case no edge is added, while the other two cases add the directed edges (2, y) and (y, 2) respectively. Consider the collection of 2D rigid bodies viewed by a 1D camera which is illustrated in Fig. 4. Figure 5 (a) shows v (b) The first step in analyzing the existence of binary occlusion relations for an arbitrary pair of bodies is to model the bounded motion between them. We fix A and let M ( B ) denote the union of all possible positions of B. Its convex hull, C H [ M ( B ) ] , can be partitioned from A by a separating plane if the occlusion is unambaguous. This is illustrated in Fig. 3 (a) for two 2D bodies viewed by a 1D camera. The partition creates two half-spaces. If the image plane projections of A and CH[M(B)] don\u2019t overlap, A E B", + " For a specific object like the hand, techniques like velocity-based prediction can be used to handle ambiguous configurations. the occlusion graph for the system under bounded translations in the plane. When the object configuration admits a visibility ordering, it can be obtained by searching the occlusion graph. In general, the occlusion graph must be acyclic to induce a natural order on the set of objects. When the occlusion graph is acyclic, it can be topologically sorted by depth-first search to produce a visibility ordering. Figure 5 (b) shows the ordering produced by sorting the sample occlusion graph. The sorted graph has the property that all edges are directed left to right. Taking the vertices in that order guarantees that no object will be occluded by an object that follows it in the list. These results give sufficient conditions for the existence of a visibility ordering for an arbitrary object. Existence hinges primarily on the absence of occlusion ambiguities, which is determined by the relative motion and the temporal sampling rate" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001338_polyeng-2020-0326-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001338_polyeng-2020-0326-Figure4-1.png", + "caption": "Figure 4: Three different shunt threads and profiles.", + "texts": [ + " Ordinary thread has considerable conveying capacity, but it is not conducive tomixing and exhaust. The shallow groove shunt thread, which mixing capacity is between the ordinary thread and the pin shunt thread, usually appears in the exhaust zone of conical counter-rotating twin screw to improve the exhaust capacity. The mixing capacity of pin shunt thread is the strongest, but its conveying capacity is very low. Pin shunt thread usually appears in the homogenizing zone and occupies 1/3 of the length of the homogenizing zone [21]. Figure 4 shows three common threads and profile of each one viewed from the direction perpendicular to the conical helix of shunt thread. Gray area in Figure 4 represents the flank of the ordinary thread, and blank area enclosed by dotted line and solid line represents the profile of the shunt thread. The meaning of the parameters H1, R, W1 is shown in Figure 4. In the conical counter-rotating twin screw, the shunt thread is an ordinary thread adding a reverse swept cut thread with a certain profile. In order to facilitate the drawing of the shunt thread profile, a new sketch plane is created, which is perpendicular to the tangent line at the conical helix starting point and passes through the starting point of the shunt thread conical helix. Finally, the structure of the shunt thread is generated by \u201cswept cut\u201d command in SolidWorks. As a visual language, VB can be used to create a userfriendly interface" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003458_s12239-019-0096-6-Figure10-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003458_s12239-019-0096-6-Figure10-1.png", + "caption": "Figure 10. (a) Mesh of generator; (b) Flux density distribution of generation at 1000 r/min.", + "texts": [ + " The 2-D model of the generator and the coupled field circuit are shown in Figures 9 (a) and (b). The rotor of generator is NO 10 steel, the stator of generator is 50J250, and the permanent magnet is N38SH. The axial length of generator stator and rotor are 65 mm, the axial length of permanent magnet is 6mm, the resistance value of retarder coil is 3\u03a9. Other design parameters are shown in Table 1. Considering the influence of ambient air, the mesh model is set to 1.25 times of the stator. The rotor speed is set at between 250 and 2000 r/min in FEA. Figure 10 shows the magnetic density and mesh distribution of generator for the analysis model. Based on the above analysis for the SLB-EMR, we developed a prototype of the novel retarder as shown in Figure 11. To test the performance of the SLB-EMR, the bench test was carried out. The test bench was mainly composed of the high power driving motor, the N = bmN hmN\u2013 BrAm 10 4\u2013 bmN = n 1 f \u2013 n 1+ -------------------- hmN = n f 1+ n 1+ ----------------- f = fad n ----------- C = Sc Dil 2 lefnN --------------- A = mNIN Dil ------------ transmission, the torque sensor, the electric control cabinet, the data acquisition system, and the cooling system, as shown in Figure 13" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001120_6.1997-1198-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001120_6.1997-1198-Figure7-1.png", + "caption": "Figure 7. Details of Smart Trailing Edge Design", + "texts": [ + " Further details of the torque tube design are presented in Reference 10 and 11. Adaptive Control Surfaces: The aerodynamic benefits of contoured hingeless surfaces are well known [1-3]. To implement these types of control surfaces, SMA based actuation systems are ideal because of their high force and high strain capabilities [12, 13]. Figure 6 shows a schematic of the adaptive control surfaces with embedded SMA wires in top and bottom face sheets which provide two-way \"antagonistic\" actuation. Figure 7 shows the final system used for the wind tunnel model. Approximately forty 20 mil diameter wires were used to obtain the equivalent of ten degrees of rotation. Because of the complex thermo-mechanical behavior of the SMA wires, i t was essential to incorporate sensors to determine the true position of the control surfaces. The most suitable sensors were fiber-optic sensors, and a suite of extrinsic FabryPerot interferometric (EFPI) strain sensors were embedded in the control surfaces and calibrated to provide an accurate measure of control surface actuation" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000474_bf01171588-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000474_bf01171588-Figure1-1.png", + "caption": "Fig. 1. Improveddesignofrol ler bearing.", + "texts": [ + " If, for instance, the row of rollers 2 is overloaded in comparison with the second row, then the axial component of the radial load occurring in this case moves race 3 toward the second row until the load between them is balanced. If the second row is overloaded relative to the third, then the axial load acting on the \"floating\" shoulder 6 moves it toward the third row until the load is balanced, etc. As a result, the radial load is balanced completely with respect to the rows. At present two type 77752 bearings, whose design was changed to that shown in Fig. 1, are being tested in the back-feed rollers of the 140 and 250 automatic mills of the Azerbaidzhan Tube-Rolling Plant. The tests are showing that the new bearings perform capably and have a longer life than the existing ones. The introduction of the improved bearing will effect a considerable economy. Azerbaidzhan Tube-Rolling Plant. Khar'kov Polytechnic Institute. Translated from Metallurg, No. 10, pp. 38- 39, October, 1974. 9 1975 Plenum Publishing Corporation, 227 West 17th Street, New York, iV" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001576_383082.383106-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001576_383082.383106-Figure2-1.png", + "caption": "Figure 2. Configuration of the mixer (a) and its small signal model (b).", + "texts": [ + " Permission to make digital or hard copies of all or part of this work for personal or classroom use is granted without fee provided that copies are not made or distributed for profit or commercial advantage and that copies bear this notice and the full citation on the first page. To copy otherwise, or republish, to post on servers or to redistribute to lists, requires prior specific permission and/or a fee. ISLPED\u201901, August 6-7, 2001, Huntington Beach, California, USA. Copyright 2001 ACM 1-58113-371-5/01/0008\u2026$5.00. The topology of the mixer is shown in Figure. 2. A p-MOSFET with an inherent LPNP P1 is used as a four terminal device. The RF input signal VRF is applied to the gate of the MOSFET and LO signal VLO is fed to the base of the LPNP. The IF output signal Vout output is accessed from the drain. VG1 , VB1 and Vdd are the bias and supply voltages respectively. The n-MOSFET M2 acts as an active load and the conversion gain of the mixer can be adjusted by changing its gate voltages VG2 . LS is used to shift the ac voltage level at the source of M1 and creat coupling between the two devices. The small signal equivalent circuit of the mixer is shown in Figure. 2 (b). Vgs is the signal voltage across Zg which is the impedance associated with the gate of M1, gm is the transconductance of M1, ZS is the impedance of LS , \u03b2 is the current gain of P1 and Zb is the impedance associated with the base of the LPNP. To simply the circuit, the n-MOSFET is replaced by a resistor RL . VS is the voltage at the source of the M1. M1 should be biased in the saturation region where the gate voltage can control the source to drain current. P1 must be biased in the active region so that a small variation of the base voltage can result in a large variation in the collector current \u03b2ib and thus vary VS effectively by changing the current flowing through LS " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003355_j.jmatprotec.2014.02.005-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003355_j.jmatprotec.2014.02.005-Figure1-1.png", + "caption": "Fig. 1. Principle of double-sided tube hydroforming in", + "texts": [ + " In addition, numerical simulation was conducted using the Abaqus/Explicit software. The thickness distribution and effective strain distribution were validated detailedly with experimental, theoretical and simulation study. Finally, the deformation mechanism in the transition zone under different external pressure is clarified by the electron back scattering diffraction (EBSD), fracture morphology and micro hardness. 2. Principle of double-sided tube hydroforming in a square-section die The principle of double-sided tube hydroforming in a squaresection die is shown in Fig. 1. Compared with the conventional tube hydroforming of square section in which the tube is deformed toward to the die corner only by the function of internal pressure, external pressure is introduced to the tube outside simultaneously in the double-sided tube hydroforming of square section and the tube is deformed under the hybrid effect of internal pressure and external pressure. The external pressure, as a supporting role to increase the normal compressive stress, could change the stress state and increase the hydrostatic pressure for the tube, thus improve its deformation ability" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003634_powereng.2015.7266300-Figure11-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003634_powereng.2015.7266300-Figure11-1.png", + "caption": "Fig. 11. On-drum in wheel BLDC electric motor design to be prototyped.", + "texts": [ + " Future studies going to focus on eddy current loss minimization in permanent magnets as well as copper loss optimization in stator windings as studied in [14-17]. Also load tests will be carried out by coupling the motor in the enclosure between the wheel rim and the drum brake housing, where it will be tested for performance to assess whether the design is suitable as part of an electric car conversion kit. The new design has been compared with a commercial in-wheel electric motor designed and manufactured by Protean Company. The differences of both designs can be evaluated in Fig.11 and Fig. 12. It is clear that new in-wheel electric motor is specifically designed for the use electric vehicle conversions of existing cars. Since the design is suitable for placing in the enclosure between the wheel rim and drum brake housing, it differs from existing technologies. This work is jointly funded by Small and Medium Enterprises Development Organization of Republic of Turkey (KOSGEB) under the project number 2014-692-7/02 and Turkish Scientific Research Council (TUB TAK) under grant number 113M070" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001594_9783527646982.ch6-Figure6.5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001594_9783527646982.ch6-Figure6.5-1.png", + "caption": "Figure 6.5 (a) Geometrical parameters of a horseshoe pattern stretchable interconnect; (b) examples of the horseshoe interconnect with different turning degrees ( \u03b8 = 0 \u00b0 and 45 \u00b0 ).", + "texts": [ + " However, the stresses are still concentrated in a small region of the track, which depends on the radius of curvature of the rounded region. Furthermore, similar to the square design, the straight vertical lines are detrimental for deformations perpendicular to the axis of the meander. In the optimal shape (Figure 6.4 d), the stress is distributed in an extended part of the conductor. In this case a larger radius can be used, keeping the amplitude of the designed meander constant. 6.2.1.4 Optimization of the Horseshoe Shape of Conductor The meandering horseshoe pattern is created by joining a series of circular arcs, as shown in Figure 6.5 , where R is the inner radius of the circle, W is the width of the copper trace, and theta ( \u03b8 ) is the angle, measured clockwise, where the two arcs of circles intersect. When \u03b8 = 0 \u00b0 , we have a semicircle design, if \u03b8 > 0 \u00b0 , we obtain the horseshoe design [1 \u2013 3] . In general, the metal interconnections in a stretchable electronic circuit can fail due to different mechanical factors. First, if the stresses and the strains are high and exceed the ultimate strength of the metal, the conductor track will break causing a loss of continuity" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000839_978-981-10-6553-8_51-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000839_978-981-10-6553-8_51-Figure3-1.png", + "caption": "Fig. 3 Design model of actuator in Solidworks", + "texts": [ + " In order to avoid too much braking force resulting in stop-and-go trips, the STM32 can control the braking force to be precise. Besides, the actuator uses a fast response servomotor as the power source to shorten the response time. Finally, in order to reduce the modification of the original car, the actuator installation structure can be designed based on the original car. In order to verify the correctness of this design, a test prototype of the intelligent speed limit system has been manufactured, referred to Fig. 3. In the workflow setting, when the horn of the servomotor 1 is rotated, the arm 6 is rotated about the rotary shaft, thereby bringing the brake pads 4 and the friction cylinder 5 closer or away. When the frictional resistance caused by the pressing of the brake pads and the friction cylinder passes to the wheel, the car will slow down or stop, that is, to complete the speed limit task. By setting the rotation angle of the servomotor, the user can change the pressing force between the brake pad and the friction cylinder that affects the friction force as a result, the speed limit force can be adjusted as needed" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003052_1.3604662-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003052_1.3604662-Figure8-1.png", + "caption": "Fig. 8 S e g m e n t of p i e c e w i s e l inear a p p r o x i m a t i o n for r ise portion of m a g n e t i z a t i o n curve", + "texts": [ + " The actual solenoid and its driving circuit must be used in this measurement as will be subsequently apparent. The measured mmf which is designated (iV7)/ must be regarded as the mmf necessary to produce the total flux field in the solenoid at the armature closed position. A sketch of the \"magnetization curve\" is shown in Fig. 7. The rise portion is used when the calculated value of (d/dt) is positive and the fall portion when (d/dt) is negative. T o obtain a functional relation between (<\u00a3) and (,Ni)It the rise portion is approximated b y a piecewise linear function. With reference to Fig. 8, a general portion of the curve can be described by Wh = Z{4> - fr) + (Ni), (32) 444 / A U G U S T 1 9 6 8 Transact ions of the AS M E Downloaded From: http://manufacturingscience.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/journals/jmsefk/27526/ on 06/14/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use The fall portion is approximated by two linear functions and a polynomial. The function (Ni)j can now be solved explicitly in terms of the instantaneous value of () from the appropriate function" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001395_ijvp.2016.075351-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001395_ijvp.2016.075351-Figure6-1.png", + "caption": "Figure 6 Tracked vehicle contact models (see online version for colours)", + "texts": [ + " The tracked vehicle model shown in Figure 1 will be used in the following section to compare between post-processing decoupled rigid body/deformation stress analysis based on the simplified FFR equations and the ANCF fully coupled stress analysis that takes into account the effect of the deformation on the rigid body motion. The tracked vehicle modelled is an armoured personnel carrier consisting of a chassis, idler, sprocket, 5 road-wheels, and 64 track links on each track side (right and left). Figure 6 further shows the contact engaged between track links and other components such as the sprocket, road wheels, and ground. The vehicle model also contains a suspension system made up of road arms placed between the road wheels and the chassis. A shock absorber is connected to each road arm at an initial angle. The road arms and sprocket are connected to the chassis by revolute joints and the road arms are connected to the road wheels through revolute joints as well. Each track system is made up almost entirely of revolute joints with the exception of a single bearing element used to define a secondary joint in order to avoid the singularities associated with closed loop chains" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000263_eiconrus.2019.8657035-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000263_eiconrus.2019.8657035-Figure1-1.png", + "caption": "Fig. 1 A test structure to measure temperature rise on the metal sensors", + "texts": [ + " The material stack-up file is integrated into the PDK for each technology and is provided to the designer. The preciseness of the film dimensions and thermal conductivity values denotes the accuracy of the electro-thermal modelling. The dimension values are defined in the process specification provided by the foundry. However, the conductivity values are technology dependant and shall be accurately extracted. III. TEST STRUCTURES The test structures consist of two metal electrodes and silicon dioxide between them, as shown in Fig.1. The metal electrodes represent a metal resistor. There are five metal layers in the IHP 0.25 \u03bcm SiGe BiCMOS technology: metal1(M1), metal2(M2), metal3(M3), top metal1(TM1) and top metal2 (TM2). The test structures for all metal resistors combinations were designed and measured. The heat is generated by applying some power at the top resistor (heater) and it flows down to the bottom one (sensor). For the measurements the input power of 0.4 W and 1 W was used. The ratio between width (W) and length (L) of the heater and the sensor shall meet the criteria WHEATER >> WSENSOR and LHEATER >> LSENSOR to avoid temperature effects on the sides of the heater", + " Conductivity W/(m\u22c5K) 20 1 237 174 The materials mentioned in Table I have fixed thermal conductivity values. However, the \u0198SI is temperature and technology dependent. To extract real thermal conductivity values of silicon a method described in [2], [3] was performed. The extracted values are different from the values described in the literature [12-14] which can be correlated with the high doping of silicon in IHP 0.25 \u03bcm SiGe BiCMOS technology process. The schematic circuit used for the simulation is similar to the one shown on Fig.1. The sensor current (IDC) is set to 30mA. As there are no power source available for circuit simulator the voltage source was used. However, the input power of the heater for the circuit simulation must be matched with the input power used for the measurements by adjusting the input voltage VIN. Identically, to the measurements the input power of 0.4 W and 1 W was applied for the simulation. The electro-thermal simulator calculates the temperatures on both heater and sensors. The temperature rise on the sensor can be calculated using formula (3)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001942_aim.2011.6027029-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001942_aim.2011.6027029-Figure2-1.png", + "caption": "Fig. 2. Hopkinson method", + "texts": [ + " RthermalRotor/Stator = TRotor \u2212 TStator Pin \u2223\u2223\u2223\u2223 steady\u2212state (5) Rthermal Stator/Ambient = TStator \u2212 TAmbient Pin \u2223\u2223\u2223\u2223 steady\u2212state (6) Where: PR - total mechanical losses PJ - Joules losses \u03c4thermalRotor - thermal time constant of rotor \u03c4thermal Stator - thermal time constant of stotor mwinding - mass of winding mmotor - mass of motor The next section describes methods for determining the efficiency of an engine by separating global losses. This principle can be used to access Joules losses. Therefore, if Joules losses are known, electrical resistance variation can be determined, which is important for DC motor thermal modeling. There are different scenarios for determining thermal parameters. They can be divided into two main categories: with non blocked rotor and with blocked rotor. A. Identification methods with nonblocked rotor The Hopkinson method is the method where it is necessary to have three motors [12] (figure 2). PM + PG = \u03b7aux \u00b7 Uaux \u00b7 Iaux (7) There are two identical engines, one that should be known and must be more powerful compared with the first two. In order to find necessary data and making thermal model, equation 7 is used, where \u03b7aux is efficiency of auxiliary motor, PM and PG are electrical powers of motor and generator, respectively. This equation cannot be used, because of mechanical noise due to coupling and because auxiliary motor parameters should be well known. The Hutchinson method [12] uses two identical DC motors and two sources, where the first source is used to apply desired speed, and the second source is used to apply desired torque (figure 3)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001338_polyeng-2020-0326-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001338_polyeng-2020-0326-Figure3-1.png", + "caption": "Figure 3: Schematic of thread profile.", + "texts": [ + " From formulas (1) and (4)\u2013(7), the calculation formula of the thread lead T of the ith segment is Ti = 4iVi 2\u03c0a(D \u2212 a) \u2212 D 2 \u00b7 arccos(a/D) + a \u0305\u0305\u0305\u0305\u0305\u0305 D 2 \u2212 a2 \u221a , (8) a = as + (2L \u2212 xi1 \u2212 xi2)tan(12 \u03b6), (9) D = rs + hs + (L \u2212 1 2 xi1 \u2212 1 2 xi2)tan(12 \u03b4), (10) where xi1 is the abscissa of the initial position of the ith thread, xi1 = \u2211i\u22121 1 ( li + si), and xi2 is that of the end position of the thread of segment i, xi2 = \u2211i 1li +\u2211i\u22121 1 si, li is the length of the ith thread and si the nonthread length. In the process of conical screwmodeling, it is very important to obtain the coordinates of the four vertices of the screw profile. Figure 3 presents a half-view of any cross-section through the axis of the screw. Among them, the quadrilateral surrounded by points ABCD is the thread profile, and the quadrilateral surrounded by the point EFGH is half of the cross-section of the screw root. For each thread segment, scanning the thread structure along the conical helix requires calculating the coordinates of the four vertices of the thread profile at the starting position of each thread segment, and then connecting the four points to forma closed quadrilateral. Assuming that the thread starting angle \u03b1 is the clockwise rotation angle along the positive direction of x, the coordinates of points E, F, G, and H are (0, 0, 0), (L, 0, 0), (L, rscos\u03b1, rssin\u03b1), and (0, Rscos\u03b1, Rssin\u03b1), respectively. According to the geometric relationship in Figure 3, the 3D coordinates of the four points A, B, C, and D of the thread profile at the starting position of the ith thread can be calculated as follows: Ai{xi, [rs + (L \u2212 xi)tan(12 \u03b6)]cos\u03b1, [rs + (L \u2212 xi)tan(12 \u03b6)]sin\u03b1}, Bi{ \u2212m, (ya +m)cos\u03b1, (ya +m)sin\u03b1}, Ci{xb \u2212 wi \u22c5 cos(12 \u03b4), [yb + wi \u22c5 sin(12 \u03b4)]cos\u03b1, [yb + wi \u22c5 sin(12 \u03b4)]sin\u03b1}, Di{xc \u2212 n, (yc \u2212 n)cos\u03b1, (yc + n)sin\u03b1}, where xi =\u2211i\u22121 1 ( li+di), m=h1 \u22c5sin(\u03c8 \u2212 1 2\u03b6 )/cos\u03c8, n=h2 \u22c5sin(\u03c8+ 1 2\u03b6 )/cos\u03c8, and h1 =hs+(L \u2212xi)tan( 1 2\u03b4 \u2212 1 2\u03b6 ), h2 =h1 +wi \u22c5sin( 1 2\u03b4 \u2212 1 2\u03b6)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000072_j.ijmecsci.2004.06.005-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000072_j.ijmecsci.2004.06.005-Figure1-1.png", + "caption": "Fig. 1. Schematics of the beam model considered in the present study. (a) Beam subjected to a distributed axial load due to friction. (b) Free body diagram of a beam element. (c) G = 0. (d) G = 0.", + "texts": [ + " The physical model is discussed in Section 2 and the governing equation of motion of the beam is shown to be of a Leipholz column subjected to a parametric excitation with two fundamental frequencies. A solution methodology is outlined in Section 3 and the instability analysis is considered in Section 4. Numerical results and their implications to the brake instability problem are discussed in Section 5. Consider a beam placed on an elastic medium (lining) that is subjected to external excitation by means of a moving, distributed, oscillatory motion as depicted in Fig. 1(a). The moving oscillatory 1 A more accurate pcr was found to be 40:05EI=L3 by Sugiyama et al. [15]. motion may be considered as a result of vibration of a neighboring elastic body (e.g., a beam or plate) in contact with the lining; for instance, the interaction between the back-plate of the brake pad and a rotor spinning with zero nodal circle 2 through the friction material. The relative sliding between the lining and neighboring elastic body generates a distributed frictional traction on the surface of the lining", + " In this study, the following assumptions are employed to examine the conditions for the onset of instability of the transverse 2 In general, the disc modes measured in experiments are observed to have zero nodal circles at squeal [23,24]. motion of the beam: (1) the lining and moving elastic body are in a state of conformal contact such that there is no loss of contact in the interface; (2) the material properties of the lining and remain constant; (3) the coupling between the transverse and axial motions of the beam is neglected. The free-body diagram of the system is shown in Fig. 1(b). Note that the distributed axial load is a slope-dependent nonconservative follower-type force due to friction. Denoting W as the transverse displacement of the beam, X and T the spatial and temporal variables, respectively, the corresponding equation of motion governing W of the beam can be derived as A @2W @T 2 + EI @4W @X 4 + @ @X ( \u222b L X p( ; T ) d @W @X ) + p(X; T ) @W @X +Ks Ns\u2211 i=1 (X \u2212 Xi)W = p(X; T ); (1) reducing to A @2W @T 2 + EI @4W @X 4 + \u222b L X p( ; T ) d @2W @X 2 + Ks Ns\u2211 i=1 (X \u2212 Xi)W = p(X; T ); (2) where denotes the e8ective mass density, A the cross-section area, EI the =exural rigidity, L the span length of the beam, Ks the elastic sti8ness of the support against the beam, (\u2022) the Dirac\u2013 Delta function, Ns the number of intermediate supports, and Xi are the locations of the supports", + " Based on the aforementioned assumptions, the instantaneous distributed normal pressure p can be written as the sum of the static (ps) and dynamic (pd) normal pressures p(X; T ) = ps + pd(X; T ): (3) Denoting K and G as the transverse and shear moduli of the lining, respectively, p can be expressed in terms of the relative displacement of the beam, lining, and its elastic properties p(X; T ) = K( +W \u2212 We)\u2212 G @2(W \u2212 We) @X 2 ; (4) where denotes the static deformation of the lining due to ps and We the transverse displacement of the neighboring elastic body. It should be noted that if G = 0 in Eq. (4), the deformation of the lining depends only on the local pressure load as illustrated in Fig. 1(c). However, for most elastic materials, a deformation pro 2000 (rise and fall time increase) The simulation result for SPICE3 MESFET model did not support this behavior as shown in f i g u r e 9 where V(3) ,the output of the inverter of f i g u r e 4 is stuck at one when R= 10K (compared with fault free output in f i g u r e 7 ) . fault #4 ........... when O< R <50000 For SPICE3 MESFET model when R= 10k an increase in the rise and fall times were obtained as shown in f i g u r e 10 (compared with fault free output in F i g u r e 7) fault #5 ............ when O< R > 50k For SPICE3 MESFET model when R< 18k a proper operation is reported and the rise time was the same as shown in f i g u r e 11 (compared with f i g u r e 7) , fault #3 ........... when For SPICE3 MESFET model the DC response as shown in f i gu re 9 was drasticaly different to the point one can say, it is almost stuck at one ( for such slow fall in the output against the input compared with fault free of f i g u r e 8a,b,c ). In f i g u r e 8d the transient response is shown when Rd=lOK . Clearly the output is stuck at one as it is suggested for the D.c. response. Inserting different types of faults showed that there is a minimal deviation from the results in section 5. However one should not hastely drow a conclusion that the results are essentiallly the same. In the above simulation, we used the same DCFL technology but with different transistors. HEMT is very sensitive to delay faults since it is supposed to operate at higher frequencies and some variation in fault mapping is possible as shown under fault #10 and fault #5 . (rise time increase) (rise time increase) R> 1 OK (output stuck at one) VDD A*8 Figure 6 the DCFL inverter with faults numbers E Proceedings - 1989 Southeastcon 4b: m 6 > 9 > 3 !l! '1 , > . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . - - . . - - - + . . - - - - - + - - - - - - > 3 N 9 Input Voltage V(I) Figure 8a D.c. seep for the inverter of Fig. 6 from o to 2 Volts Vin is the input and V(3) is the output . . . . . . - - - - - . . + - - - - - - I - - - - - - > 3 0 ? Input Voltage V(1) Figure 8b D.C. sweep for the inverter output V(3) when Rd= 10K low output response due to invrease in Rd In order to conduct a careful analysis for the fault behavior, the following considerations must be observed and used as a cornparision standards: 1) The circuit technology ( BFL, SDFL, DCFL etc) 2) Transistor technology ( HEMT, MESFET, 3) The simulation model (SPICE2, 4) The frequency band of operation 5) The transistor parameter variations ( VBI, VTH, tD, B, Lg, W etc) 6) Using and following the same design rules for the logic gates. HBT etc) (switching speed) SPICE3, etc) CMOS Modeling Approach for GaAs There has been no major effort to investigate the appropriateness of using modeling techniques in CMOS698 for timing and stuck-open faults to model GaAs timing faults. Delay faults in CMOS logic circuits may occur whenever a transistor path of CMOS gate has a lower conductance than assumed. The propagation delay from a gate input to the gate output will exceed a specified maximum value. Some reasons for the reduced conductance are a poor contact, a narrow interconnection line or a threshold voltage shift. These types of physical faults do occur in GaAs and are considered more crucial for ultra-high speed applicatjons. Proceedings - 1989 Southeastcon To detect a timing fault in CMOS6i8, an ordered two pattern input sequence is required, these patterns will also detect a stuck-open faults in the same transistor path. Timing faults in CMOS are detected on the logic gate level using two pattern test (initialization + test) applied under worst case clocking condition. In the case of an open fault the gate output will be stuck at the previous value for one cycle. The sequential behavior is well known from stuck open faults. Hence, it follows, that a delay fault is detected by an input sequence which detects a stuck-open fault in the same transistor path, if and only if the fault-initialization pattern is followed directly by the fault-sensitization test pattern (forcing the fault's effect to an observable output of the circuit). As an example, a two-input CMOS NAND gate and its logic gate model, equivalent to the model proposed by Jain and Agrawa17 is shown in f i g u r e 1 2 . An extended logic gate model for a two-input CMOS NAND gate suitable for delay fault simulation is shown in f i g u r e 13 It includes the known logic gate model in f i g u r e 12. This model also covers stuck-open faults. The RS-flipflop, at the top, will retain its previous value in the current test cycle. In this case D becomes \"1\". Otherwise the fault's effect is not exposed with the current input pattern and either S or R is \"1\" resulting in D= \"0\". D is fed to the memory element. At the begining of the simulation the a FF(top) holds a \"0\". If the R S flipflop retains its previous value for more than one cycle both D and D' are \" I \" , and then M is \"0\". In this case the output multiplexer controlled by M passes the output REF of a faultless reference of the modeled CMOS gate to the output of the logic gate model OUT and prevents a fault detection. Otherwise the model behaves as shown in f i g u r e M=l, and hence the output of the RS-flipflop Q is passed to the model output OUT. i i i / Proceedings - 1989 Southeastcon A software implementation in a parallel fault simulation program results in an average decrease of simulation speed of about 15% compared to the known stuck-open model. However logic gate models will always require a massive overhead in both memory and time. GaAs Gate Modeling As a result of an extensive study of the fault mechanisms of GaAs transistor using switch-level simulation(SP1CE) we can develop better understanding about the behavior of GaAs transistors during physical failures. The next step will be investigate the usefulness of the proposed CMOS models in the literature and see their capability of maping accurately the physical faults in GaAs. The CMOS model ,discussed above, might prove useful ,in most cases, for high speed applications in GaAs. However, further research effort is needed to answer some of these concerns. An economical decsion is also needed to balance the choice of transistor model, gate-level model and the simulation algorithm. These choices will greatly influence the simulation cost (memory and computer time) also as well as the accuracy and the size of fault coverage , For instance, the transistor model should imply a reasonable trade off in parameter values. And the gate-level model complexity should not prove to be a formidable task to program the library routines. No single major effort has been devoted to the development of an effecient GaAs fault model at the gate or transistor-level . The need for such a model is crucial for maping the crucial timing faults that are inherent in this emerging technology. Conc lus ion This report serves the purpose of reviewing the technology of GaAs from a testing-engineer point of view. It overviewed the features and the fault behavior of GaAs. transistors. Fault characterization of physical failures was shown and timing faults were shown to be the most serious faults in GaAs. Timing algebra was presented to propagate and detect timing faults in the gate-level simulation. Simulation results for the same type of faults in the DCFL inverter was shown using the modified SPICE3 MESFET model and it was compared with the correspondent results for HEMT simulation model. It was shown that there exists some difference in fault behavior due to a difference in the simulation model and ,subsequently, some standards for fault behavior study were needed to assist in drawing a useful comparisons among different GaAs technologies. These standards were listed and test engineers should consider such variations during the study and comparison of fault behavior. Finally it was suggeted that it might be possible to make use of exsting CMOS fault models to detect both timing and traditional stuck at faults in GaAs. A timing fault model for CMOS gates was presented to detect both timing and stuck-open faults in CMOS. The need for developing a gate-level fault model based on good knowledge about the behavior of GaAs in the transistor-level is essential for an accurate physical-failure maping to this model. Designers should develop an evaluation measure for a suitable trade off bettween accuracy, complexity, and cost. M. J. Howes And D. Morgan GaAs Materials, Device And Circuits John Wiley & Sons Ltd. New York, 1985. Guy Rabbat Hardware And Software Conceets o f VLSl Van Norstrand Reinhold Co. Inc. New York 1983 W. R. Curtice, \"A MESFET Model for Use in the Design of GaAs IC\" IEEE transactions On Microwave and Techniaues May, l980 pp. 448-456 S. Lee and C. Crowel Optimiza!ion of HEMT in Ultra High Speed GaAs lntearated Circuits IEEE Transactions on Electron Devices, June,1983 p p 83-103. T. Cunningham, W Kent, P. Banerjee \"Fault Characterizations and Delay Faults Testing of GaAs Logic Circuits\" ,1987 IEEE International Test Conference, pp836-842. S. Koeppe , \"Modeling and Simulation of Delay Faults in Logic Circuits\" 1986 IEEE International Test Conference, pp. 530-535. S. K. Jain and V.D. Agrawal, \" Test generation for MOS Circuits Using D-Algorithm, Proceedinas of 20th Desian Automation Conference, Miami Beach, FI, June 1983 pp. 64-70 S. AI-Arian and D. Agrawel, \" Physical Failures and Fault Models of CMOS Circuits,\" IEEE Transactions on Circuits and Svstems, March 1 9 8 7 , Vol. CAS-34, pp. 269-279. T. Chen and M. Shur \" A Capacitor Model for GaAs MESFET's\" IEE Transactions on Electron Devices, Vol. ED-12, No. 5, May 1985. [ l o ] D. Widiger and J. Coleman, \" Two-Dimentionat Transient Simulation of an Ideal High Electron Mobility Transistor\" IEEE Transactions o n Electron Device, Vol. ED-23 No. 6, June, 1985 pp. 1092-1 099. Proceedings - 1989 Southeastcon" + ] + }, + { + "image_filename": "designv6_24_0000082_acc.2009.5160177-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000082_acc.2009.5160177-Figure1-1.png", + "caption": "Fig. 1. 2D illustration of a parabolic projection.", + "texts": [ + " It is supposed that the observed position on the planar surface is defined by y\u2217(t) = [y\u2217 1(t), y\u2217 2(t)]\u22a4 = [ X(t) \u2212 m Z(t) , Y (t) \u2212 n Z(t) ]\u22a4 , (10) where m and n are known constants. Parabolic Camera: The parabolic camera assumes parabolic projection, which refers to the projection induced by a parabolic mirror onto an image plane. The parabolic projection of a point P in space is the orthographic projection of the intersection of the line connecting the point P with the paraboloid\u2019s focus, and the paraboloid [25]. Consider a paraboloid placed in such a way that its symmetric axis is the Z-axis and its focus is at the origin, as shown in Fig. 1. Further, assume that the paraboloid\u2019s focal length f equals 1/2, without loss of generality. The function of the parabolic surface is 1 2 (x2 1 + x2 2 \u2212 1) = x3. (11) The projection of a point P = [X, Y, Z]\u22a4 onto the paraboloid surface can be described by [X, Y, Z] \u22a4 /L with L = \u00b1 \u221a X2 + Y 2 + Z2 \u2212 Z. Let P1 be the intersection point (as light ray enters the paraboloid), as shown in Fig. 1. Correspondingly, L = \u221a X2 + Y 2 + Z2\u2212Z [15], [16], [25]. Let x(t) = [x1(t), x2(t), x3(t), x4(t)] \u22a4, = 1 L(t) [X(t), Y (t), Z(t), 1]\u22a4, (12) where L(t) = \u221a X2(t) + Y 2(t) + Z2(t) \u2212 Z(t). (13) The system (1) with output observations (12) is equivalent to the following system: x\u03071(t) x\u03072(t) x\u03073(t) = b1 + \u21260(t)x1 b2 + \u21260(t)x2 b3 + \u21260(t)x3 x4 + ( 1 \u2212 x2 1 1 + x3 ) 3\u2211 j=1 a1jxj \u2212 x1x2 1 + x3 3\u2211 j=1 a1jxj \u2212 x1x3 1 + x3 3\u2211 j=1 a1jxj + \u2212 x1x2 1 + x3 3\u2211 j=1 a2jxj + x1 1 + x3 3\u2211 j=1 a3jxj ( 1 \u2212 x2 2 1 + x3 ) 3\u2211 j=1 a2jxj + x2 1 + x3 3\u2211 j=1 a3jxj \u2212 x2x3 1 + x3 3\u2211 j=1 a2jxj + (1 + x3 1 + x3 ) 3\u2211 j=1 a3jxj , x\u03074(t) = 3\u2211 j=1 a3jxj \u2212 1 1 + x3 3\u2211 i=1 3\u2211 j=1 aijxjxi x4 + \u21260(t) x2 4, (14) with output y(t) = [y1(t), y2(t), y3(t)] \u22a4 = 1 L(t) [X(t), Y (t), Z(t)]\u22a4, (15) where \u21260(t) in (14) is \u21260(t) = b3 \u2212 1 1 + x3 3\u2211 i=1 bixi" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003331_9781119546924.ch7-Figure7.7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003331_9781119546924.ch7-Figure7.7-1.png", + "caption": "Figure 7.7 q-axis and 0-sequence magnetic circuits.", + "texts": [ + " Note also that only one d-axis amortisseur circuit is shown in Figure 7.6. If more amortisseur circuits are included, the L\ud835\udcc1kd branch is replaced by an appropriate number of parallel branches of inductances L\ud835\udcc1id driven by fluxes \ud835\udf13 id, i = 1, 2, ..., nd, where nd is the number of individual d-axis circuits. The two per-unitized reciprocal q-axis flux-linkage equations are summarized as \ud835\udf13q = \u2212(L\ud835\udcc1 + Laq)iq + Laqikq (7.142) \ud835\udf13kq = \u2212Laqiq + Lkkqikq (7.143) These two equations can be combined into a single q-axis equivalent flux-linkage circuit as shown in Figure 7.7, where L\ud835\udcc1kq = Lkkq \u2212 Laq is the leakage inductance. Again note that only one q-axis amortisseur circuit is shown in Figure 7.7. In the subtransient synchronous machine model, the q-axis amortisseur winding is represented by two parallel circuits. Finally, the zero-sequence flux-linkage equation is \ud835\udf130 = \u2212L0i0 (7.144) which is in general not modeled. The flux-linkage equations involve instantaneous algebraic quantities. The voltage equations, which involve time derivatives of the flux linkages, represent dynamic equations used for time simulations. The three d-axis voltage equations in reciprocal per-unit bases are ed = d\ud835\udf13d dt \u2212 \ud835\udf14\ud835\udf13q \u2212 Raid = \u2212(L\ud835\udcc1 + Lad) did dt + Lad difd dt + Lad dikd dt \u2212 \ud835\udf14\ud835\udf13q \u2212 Raid (7" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003205_ecce.2014.6954109-Figure14-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003205_ecce.2014.6954109-Figure14-1.png", + "caption": "Fig. 14: Optimal machine design without ANN", + "texts": [ + " 13 that there is very good agreement between the predicted results using ANN and the coupled model data. The trained network has been subsequently built into the optimization program. An EM-only model combined with the predicted maximum current densities provided by the trained ANN is sufficient to replace the full coupled EM/thermal model for the remaining 1,200 designs after the initial 300 designs required for training. This updated optimization with the ANN has been run, and the resulting optimal design (Fig. 15) is compared with the optimal design from Section III (Fig. 14). The key parameters of these two machines are listed in Table IV. The close match between the two designs demonstrates the usefulness of the ANN technique for further accelerating the optimization process. As discussed in Section III, the total number of analyses required to accomplish the optimization without ANN is approx. 7,500 transient EM plus 7,500 static thermal FE analyses. Applying the ANN technique, the machine design parameters and their associated maximum current densities found from the first 10 generations \u2013 that is, the first 300 designs \u2013 can be used to train a network" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001784_92.250205-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001784_92.250205-Figure2-1.png", + "caption": "Fig. 2 . Diffusion to polysilicon input.", + "texts": [ + " Use of LI here eliminates the need for contacts. Also, in many routing programs, the preferred direction for metal 1 connections is horizontal since the polysilicon gates usually run vertically between the p- and n-diffusion lines. Since LI and metal 1 make no direct connection (a contact would be required), the conflicts between horizontal metal connections and p- to n-diffusion connections area avoided. A second potential usage of the LI layer is for directly connecting sourceldrain diffusion regions to polysilicon wires as shown in Fig. 2. As with the n-device/p-device connections, such connections are presently made in metal 1, and require a minimum of two contacts (or three if both n-device and p-device diffusions are involved). LI eliminates the need for any contacts. The third use for LI is aimed at increasing global routing flexibility by facilitating the creation of polysilicon Z/O ports at the top and bottom of the cell (see Fig. 3). Leading edge routing programs are capable of exploiting a large number of routing layers for making connections to the cells" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000406_tmag.2009.2023243-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000406_tmag.2009.2023243-Figure1-1.png", + "caption": "Fig. 1. Yoke composed by several GO shifted laminations.", + "texts": [ + " First, the new design is presented, then the experimental validation procedure is exposed and finally the test results are shown and discussed. The GO steel is highly anisotropic, it has the easy magnetization direction at [001] (0 ), and the hard one at [111] (55 ) [4]. Consequently, this particularity makes the use of such steels in ac electrical machines inadecuate because of the rotating magnetic field. The new structure put fordward in this paper, gives a solution to this problem. It uses shifted laminations to constitute the Stator Magnetic Circuit (SMC) as shown in Fig. 1. The SMC is assembled shifting each lamination from the previous one with a constant spatial angle . So the easy magnetization direction appears in different parts of the SMC. This principle forces the magnetic flux to pass from one lamination to another, along the axis, in order to satisfy the principle of energy minimization [5]. Several SMCs are built using this principle. The geometry of the laminations is shown in Fig. 2(a). Two different qualities of steel, provided by ThyssenKrupp E.S (TKES), are tested" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002482_9783527631506-Figure6.13-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002482_9783527631506-Figure6.13-1.png", + "caption": "Figure 6.13 Parasitic-insensitive (PI) inverting and noninverting integrators combined: OA input connected to the signal paths during (a) clock phase \u03c61 and (b) clock phase \u03c62.", + "texts": [ + "21) is a characteristic of the integration and appears in view of the sample-and-hold operation over a full period T by the integrating capacitor Co. Because of the time relationship between \u03c61, and \u03c62, we can write as before, V (2) out = z\u2212 1 2 V (1) out (6.22) If the clock signals in Figure 6.12b are interchanged, we can similarly obtain V (2) out = C2 Co z\u2212 1 2 1 \u2212 z\u22121 V (1) in (6.23) and V (1) out = z\u2212 1 2 V (2) out (6.24) 6.5.1.3 Inverting and Noninverting Lossless Integration Combined Consider Figure 6.13a, where signal v1 is processed for inverting integration, while signal v2 is processed for noninverting integration. Equations (6.17) and (6.21) were derived assuming the presence of only one signal at a time. When both v1 and v2 are operating simultaneously, one can invoke the superposition principle of linear time-invariant networks and write V (1) o = C2 Co z\u22121/2 1 \u2212 z\u22121 V (2) 2 \u2212 C1 Co 1 1 \u2212 z\u22121 V (1) 1 (6.25) In the above, V (y) x (x = 1, 2 and y = 1, 2) represent the z-transformed variables pertaining to the signals vx (x = 1, 2) due to the two clock signals \u03c61 and \u03c62", + " Equation (6.25) represents the z-transformed output signal, when two signals v1 and v2 are fed to the OA input. Notice that, although the two signals are processed by the clock signals in different ways, they are connected to the input of the OA at the same time (i.e., clock phase \u03c61). Since the signals are switched on to the OA input during clock \u03c61, the significant output signal Vo is labeled with (1) as the superscript. If the clock phases in the paths for v1 and v2 are switched around (see Figure 6.13b) with the OA input being connected to the signal paths during the clock phase \u03c62, we could similarly derive V (2) o = C2 Co z\u22121/2 1 \u2212 z\u22121 V (1) 2 \u2212 C1 Co 1 1 \u2212 z\u22121 V (2) 1 (6.26) Note that, since now the signals are switched to the input of the OA during the clock signal \u03c62, the significant signal at the output is recognized as V (2) o . The reader is advised to practice writing down similar equations with other combinations of clock phasing and signal positioning. It is obvious that the 6.5 Analysis of SC Networks Using PI-SC Integrators 181 above technique can be extended easily to cases with three or more input signal paths" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000587_0029-5493(65)90101-9-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000587_0029-5493(65)90101-9-Figure6-1.png", + "caption": "Fig. 6. Notation for spherical part of shell.", + "texts": [ + " The constants A and B can be evaluated using any suitable stress distribution, provided this assumed stress field nowhere exceeds any of the chosen yield surfaces of section 2, e.g. / a t x = l , Q = 0 , M x = M c , N o = N c , (12) a t x : 0 , Q : Q ' , M x : M ' c , NO : N c . T where Mc, M c, N c a r e the a p p r o p r i a t e va lues of the m o m e ~ s and t h ru s t on the y ie ld sur face . S i m i l a r l y for the s p h e r i c a l p a r t of the she l l , equ i l ib r ium equat ions with the notat ion of fig. 6 become\" N 0 sinq~ + Q cosq~= \u00bdPR sinq) , s i n e + N 0 sin~0 + d~(Q s i n g ) = p R s i n ~ , (13) N, dM\u00a2 de sin(p + (Mq) -Mo)cosq ) - Q R sinq) = 0 . In o r d e r to mee t the r e q u i r e m e n t s govern ing the use of the y ie ld su r face , MO must be e l i m i - nated f rom the t h i rd equation. Two poss ib l e ways of achieving th is may be dev i sed by set t ing: a) M O = 0 , (14) b) MO = M~o \u2022 (15) In t roduct ion of the condi t ion (14) in the equ i - l i b r i um equat ions (13) and in tegra t ing l eads to: Q = (\u00bdPR -No)cp + C , =R(\u00bdPR " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000810_0470871199.ch11-Figure11.12-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000810_0470871199.ch11-Figure11.12-1.png", + "caption": "Fig. 11.12 Solid angle d\u2126 = sin(\u03d1)d\u03d1d\u03d5.", + "texts": [ + "20) In a similar way we find for the radiated magnetic field of the fully excited array Ha (\u03d10) = Ve \u03b7 Se (\u03d10) K\u2211 k=1 [1 \u2212 \u0393k (\u03d10)] e\u2212jk0r r . (11.21) In the chapter on antenna parameters we have seen that the gain function G(\u03d1) is given by G(\u03d1) = P (\u03d1) Pin/4\u03c0 , (11.22) where P (\u03d1) is the radiated power in the direction \u03d1 and Pin is the totally incident available power. The time-average power density S (watts per square metre) of the radiated fields can be calculated as [2] S(r) = 1 2 {Ea(r) \u00d7 H\u2217 a(r)} = 1 2 {EaH\u2217 a}ur. (11.23) The radiated power, P (\u03d1), per solid angle unit d\u2126 = sin(\u03d1)d\u03d1d\u03d5, see figure 11.12, is then given by P (r) = P (\u03d1) = \u2223\u2223r2S(r) \u2223\u2223 . (11.24) The radiated power into the direction \u03d10 thus becomes P (\u03d10) = V 2 e S 2 e (\u03d10) 1 2\u03b7 { K\u2211 k=1 [1 + \u0393k (\u03d10)] }{ K\u2211 k=1 [1 \u2212 \u0393k (\u03d10)] } . (11.25) The incident available power is, since we assumed a normalised characteristic impedance, Pin = K V 2 e 2 , (11.26) so that we find for the gain of the fully excited array in the direction \u03d10 Ga (\u03d10) = 4\u03c0S2 e (\u03d10) K\u03b7 K2 \u2212 \u2223\u2223\u2223\u2223\u2223 K\u2211 k=1 \u0393k (\u03d10) \u2223\u2223\u2223\u2223\u2223 2 . (11.27) So, based on the scan reflection coefficient, \u0393k, that can be obtained from pair-wise mutual coupling measurements between the array elements, and the isolated element pattern - that in general is relatively easy to obtain - we may find the gain of the complete array antenna in the direction of the scanned beam" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001942_aim.2011.6027029-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001942_aim.2011.6027029-Figure5-1.png", + "caption": "Fig. 5. Potier method", + "texts": [ + " Identification methods with blocked rotor Methods that use blocked rotor are: Rayleigh-Kapp [12] and Potier [12] methods. For both methods, two motors are coupled. In the first method, one source is necessary, and for the second method two sources are used. Equation 9 is used for calculating necessary data for the Rayleigh-Kapp method (figure 4) where Em and Eg are back EMF1 of motor and generator, respectively. Im and Ig are currents which exist in motor and generator. Equation 10 is used for the Potier method (figure 5). 1ElectroMotive Force Ur = Em +R \u00b7 Im = Eg \u2212R \u00b7 Ig (9) PM + PG = V \u00b7 Im (10) In the above mentioned methods, it is not practically possible to have identical motors, and as a result, a very small rotation will appear. Oscillation of voltage will occur, what is not desirable for calculating the circuits. The best solution is to block the rotor to get thermal resistances more easily. The motor bench is shown on figure 6. The DC motor used in the experiment is a Maxon RE 25 [11], 20 Watts, equipped with graphite brushes" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002250_tmtt.2006.877424-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002250_tmtt.2006.877424-Figure6-1.png", + "caption": "Fig. 6. Overlap effect at source node [45].", + "texts": [ + " Such equalization is relatively simple because the signal is known. The phenomenon of overlap may also degrade performance [44], [45]. Depending on the link distance, a busy-signal pulse may overlap a data pulse in time at either the source node or destination node. For clarity, we allow the destination to time its busy-signal transmission to avoid any overlap with the received data signal. Thus, a source node may lose a portion of the busy signalwhen it transmitsadatapulsewith itsPAenabled on and its LNA disabled off . Fig. 6 illustrates overlap at a source node. At Time 1, the source node transmits a pulse, which arrives one propagation time , later at the destination node at Time 2. At Time 3, the destination sends a busy-signal pulse exactly s after the arrival of the first data pulse. Finally, at Time 4, the source node receives the busy-signal pulse. In Fig. 6(a), the link distance is m so the round-trip propagation time is s. Therefore, the busy-signal pulse arrives s after the corresponding data pulse. In Fig. 6(b), the link distance is m, the round-trip propagation time is s, and the busy-signal pulse arrives s after the corresponding data pulse. Since the source node is transmitting, it loses energy from the busy signal. Note that Fig. 6 shows only the first multipath of a busy signal; in reality, a receiver could detect some portion of the multipath energy. A source node should mitigate overlap such that ideally (8) To completely avoid overlap, both the source and destination nodes may wait for the maximum multipath delay spread of between receiving a busy-signal (data) pulse and transmitting a data (busy signal) pulse. Thus, for a maximum link distance of , a PRI can satisfy (8) if [45] (9) At shorter PRIs, the source node may lose up to 2 ns of the busy-signal energy from overlap and from the enable/disable timing resulting from (6)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001536_0167-9260(96)00003-x-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001536_0167-9260(96)00003-x-Figure6-1.png", + "caption": "Fig. 6. Example of the effects of pipelining and parallelism on power consumption.", + "texts": [ + " In [24], the authors point out that a reduction of supply voltage does not necessarily translate into slower computation time and throughput rate, if accompanied by pipelining techniques and increased parallelism. The latter is accomplished by duplicating the hardware, thereby introducing redundancy into the system. Here, we wish to show that hardware redundancy can be avoided, if the algorithm's inherent parallelism is exploited. Mapping techniques like the one presented in the present paper do indeed make full use of the algorithm's concurrency and pipelineability and can therefore be used for power-efficient design. With reference to Fig. 6, consider the following example. Let M be a 2 x n matrix and x an n-vector. Also, let mk, k = 1,2, be the kth row of M. It is required to perform the multiplication M.x. The index space is 5 P = {O<~i<~n, 0~ KINEMATICS OF CHASSIS: The chassis kinematics is formulated in order to evaluate the instantaneous positions of the suspension joints on the chassis under chassis rotation or vertical motions. The suspension kinematic responses are subsequently determined from the coordinates of the linkage joints. A fixed coordinate system is assumed with its origin located at the mass center (cg) of the vehicle sprung mass in its static equilibrium. The sprung mass vertical and rotational displacements are considered about the kinematic roll center G of the vehicle body. The initial (Gy0, Gz0) and SAE Int. J. Mater. Manuf. | Volume 3 | Issue 1 306 instantaneous (Gy, Gz) coordinates of the roll center are related through the general displacement matrix Dc, formulated under a finite displacement of the chassis, given by [21]: (1) where , , and \u03d5s is the vehicle body rotation about the roll center. The y- and z- coordinates of chassis-suspension joints, Mr, Or, Ml, and Ol, shown in the Fig 1, can be determined using the displacement matrix Dc, such that: > (2) The subscripts \u2018r\u2019 and \u2018l\u2019 in Eq (2) refer to right or left suspension joints, respectively, while the subscripts \u2018y\u2019 and \u2018z\u2019 represent the lateral and vertical axes, respectively. The subscript \u20180\u2019 represents the initial coordinate of the joint. The expansion of the Eq (2) yields the expressions for the instantaneous coordinates of the suspension joints at the chassis, such that: (3) where Gy = Gy0; and Gz = Gz0 + zs. The above equation can be solved to obtain instantaneous coordinates of the chassislinkage joints for given chassis rotation \u03d5s about the roll center and/or a vertical displacement of the chassis, zs", + " KINEMATIC RESPONSES: The solution of Eqs (3) and (6) yield the camber angle responses (\u03d5ur and \u03d5ul) of the left and right wheels for given sprung mass and unsprung mass displacements either individually or simulataneously. The variations in the track width have been generally evaluated from the changes in the lateral coordinates of the wheel centres (Cr and Cl) [17], which may not truly define the track width due to lateral tire compliance and camber effects. In this study, the lateral displacement of the tire-ground contact points, Tr and Tl, as shown in Fig 1, during the wheel motions are considered as the variations in the wheel track width [5]. Thus, the wheel track variations yield wheel center lateral displacement coupled with the wheel camber angle variation. The rigid body assumption of the tire leads to expressions for the y- and z- coordinates of the tire-road contact points, Tr and Tl, as: (7) The y- coordinates of the tire-ground contact points in Eq (7) determine the wheel track width variations, while the kinematic roll center of the vehicle in the roll plane is estimated using the instantaneous centers of rotations of the wheel knuckles [5,6] as illustrated in Fig. 1. The formulations derived in the previous section are solved using Newton-Raphson method to obtain kinematic responses of the half-car model under either vertical motions of the wheels or roll motion of the chassis or a combination of sprung and unsprung masses motions. A sensitivity analysis is then performed to evaluate the relative contributions of variations in different joint coordinates. A composite performance index of camber angles and track width measures under wheel vertical displacement and chassis roll excitations is subsequently formulated and solved with constraints imposed on the roll center height and the suspension lateral packaging space to seek optimal joint coordinates. The analyses are performed using the known geometry of a double-wishbone suspension. The coordinates of right suspension linkage joints, shown in Fig. 1, are taken in mm as: Mr(430, 818), Nr(644, 852), Or(365, 360), Pr(743, 347), Cr(787, 452), Ar(660, 349) and Br(615, 920), respectively [4]. The left suspension is considered to be symmetric to the right suspension about a vertical line through the mass center of the vehicle body. Initial camber angles of the wheels are assumed to be zero, since the analyses are concerned with the variations in the responses alone. KINEMATIC ANALYSIS: The kinematic responses are evaluated in terms of: (i) variations in the roll center height, the wheel track width variations and the bump camber angles under vertical displacement inputs at the wheel centers with fixed chassis; (ii) roll camber response to chassis roll input; and (iii) variations in the bump/roll camber angles under simultaneously applied wheel centers displacements and the chassis roll" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001756_8.366381-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001756_8.366381-Figure2-1.png", + "caption": "Fig. 2. Patch geometries considered for illustration (All dimensions in mm) f, = 2..>. t m h = O.II01S.h = 1.j2.Patch surface resistivity used in computations = 3 x lO-\u2019I>m. (a) Circular patch with trim tabs, coaxial probe at (--S.783.0). -1.1. = -I!/ = 1.255. (b) Annular ring, coaxial probe at (-20.35;. 0 ) . 1.r = At/ = 0.885. (c ) Annular sector, coaxial probe at (-10.i01.-2.283).1.r. = A!/ = 0.457.", + "texts": [ + " Further, we do not calculate the near fields as seen by an infinitesimal dipole, but the near field value tested by a rooftop function. Also, the computation is made at those . r , IJ coordinates where the boundary condition was enforced on the patch. This simply requires a re-computation of (12)-(13) and (19)-(20) with z = h+d using the converged current distribution obtained from the iterative solution, and summing them up vectorially. The finite dimension of the measurement probe is then included by a convolution operation as explained below. 111. RESULTS The geometries considered for illustration are shown in Fig. 2. Also shown is the cell size used in the computations. The generalized annular sector family of patch antennas [28], [ 291 have several interesting characteristics, one of which is size reduction. All three patches were designed so that the dominant mode resonance is at more or less the same frequency. Since the fields of a circular patch are known rather well, a small complexity was introduced by adding trim tabs. The patches were excited for linear polarization so as to clearly distinguish the two near field components", + " It was therefore not possible to obtain reliable estimates of the near fields of this geometry by measurement. Note that, in academic terms, the insensitivity of retum loss to the presence of the measurement probe is necessary but not sufficient to guarantee unperturbed near fields. Geometry B which exhibits a high impedance level was matched with a coaxial triple stub tuner for the near field measurements. Measurements were made with a step size of 2 mm between successive probe positions. The step size used in computations is simply the cell size given in Fig. 2. Near field results are shown in Figs. 5-7. Both computations and measurements are made at frequencies where the impedance match is good. The scales have been rounded and the outline of the patch geometry and the position of the coaxial excitation have been superimposed. The results of both the computed and measured amplitudes have been smoothly interpolated using an option available in the graphics display program (Spyglass). The scales and the color coding scheme were selected for best visual representation of the amplitude" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003231_tap.1980.1142339-Figure10-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003231_tap.1980.1142339-Figure10-1.png", + "caption": "Fig. 10. Center-fed inverted dipole over lossy half-space.", + "texts": [ + " The vertical dipole shown in Fig. 8 is the next geometry considered. Unfortunately, as was the case for the horizontal antenna, the total [2 'mp] matrix is not in a Toeplitz form. However, the computer program is designed to take maximum advantage of the- available symmetry. As an example, Fig. 9 is included to show the radiation pattern of a 2L = 10 m, h = 8 m, center-fed vertical dipole at resonance (f = 15 MHz) located over various lossy grounds. Finally, as 2 complicated example, the inverted V dipole of Fig. 10 is considered. Again, as in the two previous sections, symmetry is used in the computer program in constructing the [ZimP] matrix. The program is tested for an inverted V dipole structure having L = 7.5 m, h = 10 m, and $ = 90\u00b0; Fig. 11 demonstrates the radiation pattern of this structure at 10 MHz and for various lossy grounds. In all three of these examples care has been taken not to 0 10 20 30 40 50 60 70 FREQUENCY (MHz) Fig. 5. Input resistance of center-fed horizontal dipole antenna as function of frequency and ground parameters" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002482_9783527631506-Figure8.22-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002482_9783527631506-Figure8.22-1.png", + "caption": "Figure 8.22 (a) A universal VM filter using a DISO-OTA, (b) its transpose, the CM counterpart of the circuit given in (a).", + "texts": [ + "30) Thus the transpose of a DISO-OTA is a single-input dual-output (SIDO) OTA; it is symbolically shown as in Figure 8.21b. A SIDO-OTA can be obtained by simply 278 8 Current-Mode Filters applying a current replica in a standard OTA. A simple realization using two single-ended OTAs is shown in Figure 8.21c. Using the transpose of a DISO-OTA, we can easily derive CM circuits from VM circuits that employ such OTAs. We illustrate this with two examples. Example 8.1. SIDO-OTA universal CM filter: Consider the VM DISO-OTA universal filter shown in Figure 8.22a (Deliyanis, Sun, and Fidler, 1999). It can be shown that the VTF of this circuit is given by T\u03bd (s) = Vo Vi = gmY2Y4 D(s) (8.31a) where D(s) = (Y1Y2 + Y2Y3 + Y3Y1) (Y4 + Y5) + (Y1Y5 + Y2Y5 + gmY2 ) Y4 (8.31b) By replacing the DISO-OTA by its transpose, we get the CM filter shown in Figure 8.22b, proposed by Al-Hashimi and Fidler (1988). It can easily be shown that the CTF Ti (s) = Io Ii is given by Ti (s) = Io Ii = gmY2Y4 D (s) which is the same as the T\u03bd (s) of the original network. Various filters such as LP, BP, and HP second-order filters can be obtained by appropriately choosing the values of the various admittances. Example 8.2. Leapfrog structure: As a second example, we consider the derivation of the CM leapfrog structure from that of a VM leapfrog structure. Consider the ladder network of Figure 7" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003203_jas.2018.7511072-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003203_jas.2018.7511072-Figure1-1.png", + "caption": "Fig. 1. Model of 2D overhead crane systems.", + "texts": [ + " 4) It needs no knowledge of the system parameter associated with a standard SMC. The rest of this paper is outlined as follows. In Section II, the model of 2D overhead crane systems is described. In Section III, the main results, including the APD-SMC law design and closed-loop stability analysis, are given. To verify the superior performance of the proposed method, some experimental results are given in Section IV. In Section V, we draw the conclusion of this paper. II. 2D OVERHEAD CRANE SYSTEM MODEL The dynamic equations of a 2D overhead crane system (as shown in Fig. 1) can be described as follows [21], [23]: (mx + mp)x\u0308 + mpl\u03b8\u0308 cos \u03b8 \u2212mpl\u03b8\u0307 2 sin \u03b8 = F \u2212 frx (1) mpl 2\u03b8\u0308 + mplx\u0308 cos \u03b8 + mpgl sin \u03b8 = 0 (2) where mx and mp represent the trolley mass and the payload mass, respectively, l and g stand for the cable length and the gravitational constant, respectively, x (t) and \u03b8 (t) are the trolley displacement and the payload swing, respectively, F is the control input, and frx denotes the friction, which is of the following form [5], [15]\u2212[16]: frx = f0rx tanh ( x\u0307 \u03b5 ) \u2212 krx |x\u0307| x\u0307 (3) with f0rx, \u03b5, krx \u2208 R1 being the friction-related parameters" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001388_s1644-9665(12)60163-0-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001388_s1644-9665(12)60163-0-Figure1-1.png", + "caption": "Fig. 1. The hydroforming principles: a \u2013 tool setup, b \u2013 initial tube, c \u2013 final product (T-joint)", + "texts": [ + " There are also presented some examples on computer modelling of these processes and limiting phenomena. Keywords: hydroforming, THF, SHF, FEM Hydroforming uses fluid pressure in place of the punch as comparing with a conventional tool set to form the component into the desired shape of the die. Generally, hydroforming processes would be classified as tube or sheet hydroforming depending on the initial shape of workpiece. In the tube hydroforming process (THP), the initial workpiece is placed into a die cavity, which corresponds to the final shape of the component, Figure 1. Next, the dies are closed under the force and the tube is internally pressurized by a liquid medium to effect the expansion of the component (internal pressure, pi) and axially compressed by sealing punches to force material into the die cavity (axial force, 2). Hence the component is formed under the simultaneously controlled action of pi and axial force. The process should be controlled to avoid failures such as buckling, wrinkling and bursting. Appropriate fundamentals to determine process controls were developed by experimental approaches as well as by means of FE simulations, e", + " Then a liquid medium can be filled between the blanks, and pressurization can be effected by a hydraulic system. It is very difficult to realize radial feeding using this method, as it is essentially a pure bulging deformation. To some extent, this technology is similar to tube hydroforming. Grey et al. in 1939 [33] formed a T-joint using a seamless copper tube and applied for and achieved a US patent in the 1940s, which gave an indication of the coming period of tube hydroforming. Until now, the forming of T-shape used for a joint is still a problem in hydroforming. Figure 1 shows one of the schemes of tube hydroforming. When using seamless tube or welded tube, the blank can be formed into the shape of the die cavity by internal pressure and when the side punches move in. Tube hydroforming has many advantages such as part consolidation, weight reduction, improved part strength and stiffness, highly accurate dimensions and low springback, lower tooling cost and fewer integrated processes, etc. which all promote rapid spreading of this technology in the automotive, household and aerospace industries [2, 34]", + " The conventional bumper structure is assembled from several parts (Figure 13a), so several manufacturing processing steps are needed, and the structure is somewhat complex. Most research work on bumper stays has focused on using reinforcing members that have complicated shapes [44\u201345]. Hydroformed bumper stay (Figure 13c,d) is rather simple in the shape but its ability to absorb energy through plastic deformation is relatively high. In most of the tube hydroforming processes, the decrease in wall thickness is prevented by compressing the tube in the axial direction simultaneously with the action of the internal pressure (see example in Figure 1). If the internal pressure is too small, the axial compression causes the wrinkling of the tube wall. Hence the paths of internal pressure and axial compression in the tube hydroforming are keys to prevent the occurrence of these defects. The finite element simulation has been employed to determine the pressure paths [16, 47\u201350]. A pulsating hydroforming process of tubes has been developed for the forming of hollow products with a complex shape [51]. An improvement of the formability by means of the pulsating hydroforming have been investigated [52\u201353] and simulated by the finite element method [52]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003984_aim.2009.5229822-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003984_aim.2009.5229822-Figure2-1.png", + "caption": "Fig. 2. Region division of rotor and stator space", + "texts": [ + " In this design, the coils are mounted on the stator to facilitate the power supply and the system motion control in our spherical actuator design. The basic structure of the proposed PM spherical actuator is illustrated in Fig. 1. PM poles are mounted on the rotor equator to generate a three-dimensional (3D) magnetic field in surrounding space. Two layers of coils are assembled on the stator to interact with the magnetic field of the PM poles, and thus to produce the spherical motion of the rotor in 3-DOF. 1) PM-pole Parameters: Figure 2 presents the shape of a single rotor pole - an approximated dihedral cone enclosed by ABCD and abcd. The dihedral cone can be specified by four parameters: longitudinal angle \u03b1, latitudinal angle \u03b2, rotor radius Rr and rotor core radius Rb. 2) Coil Parameters: Conical-shaped coil is utilized in this PM spherical actuator as shown in Fig. 3. The sectional area of coil can be specified by four parameters, i.e., R0-the distance from the rotor center to the top surface of the coil, R1-the distance from the rotor center to the bottom surface of the coil, \u03b60-the inner surface angle of the coil and \u03b61-the outer surface angle of the coil" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002177_tap.2007.915427-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002177_tap.2007.915427-Figure7-1.png", + "caption": "Fig. 7. Geometry of the proposed CP dual L-probe circular patch antenna.", + "texts": [ + "5 dB)], and consistent 90 output ports phase difference over a significantly wide band of 57.5%, from 1.3 to 2.35 GHz; hence we term it a \u201cbroadband\u201d balun. The conventional 90 hybrid coupler delivered low input port return loss , balanced output ports power distribution [ ( 0.5 dB)], and consistent 90 output ports phase difference over a much narrower band of 14%, from 1.66 to 1.91 GHz; inherently limited by its output port power distribution. The geometry of the CP dual L-probe circular patch antenna is shown in Fig. 7. The circular copper patch, of diameter , has an air substrate height above a grounded Rogers RO4003 dielectric substrate of thickness and dielectric constant . The feed network, comprising the proposed 90 broadband balun, was printed on the RO4003 substrate. The two L-probe feeds, each of diameter , vertical length , and horizontal length , were orthogonally oriented and positioned a distance away from the circumference of the patch, and soldered to the respective output ports of the feed network", + " These results reveal significant enhancements in the impedance and axial ratio bandwidths over the dual L-probe antenna presented in [12]. In terms of the common frequency coverage of , axial ratio 3 dB, and 3-dB gain (gain 5.53 dBi), the proposed CP antenna exhibits a measured CP bandwidth of 28.04% from 1.38 to 1.83 GHz. The geometry of the CP quadruple L-probe circular patch antenna is shown in Fig. 13. The quadruple L-probe antenna shares the same antenna parameters with the dual L-probe antenna shown in Fig. 7. The feed network, comprising a pair of the proposed 90 broadband baluns connected by a 180 transformer, was printed on the RO4003 substrate. To provide 180 phase shifting, the lengths of the microstrip branches must differ by , where refers to the guide wavelength at the center operating frequency, say, 1.8 GHz, in this work. The input transmission line is connected to the two microstrip branches by a quarter-wavelength transformer with characteristic impedance given by . The four L-probe feeds were soldered to the respective output ports of the balun pair, orthogonally orientated, and provided equal amplitude power with relative excitation phases of 0 , 90 , 180 and 270 " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002224_3527608117-Figure10.1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002224_3527608117-Figure10.1-1.png", + "caption": "Fig. 10.1 Schematic presentation of the gas atomization of metallic melts to produce metallic powder and semi-finished product, produced by spray forming.", + "texts": [ + " In the case of the atomization of molten baths with gases or liquids several advantages result [21, 36]: \u2022 homogeneous particle size distribution \u2022 homogeneous distribution of alloying elements \u2022 supersaturation at alloying elements also in the manufactured powder \u2022 isotropic material properties \u2022 cooling rates within the range of 10\u20132\u201310\u20134 K s\u20131 \u2022 production of large quantities \u2022 high economic efficiency Atomization procedures are wide ranging, since they are economically very interesting, and permit the manufacture of large quantities of metal and alloy powders with accurately defined compositions. This concerns the use of water or gases such as argon, nitrogen and possibly air as the atomization medium. Impurities also are hardly a problem. Such powders can undergo, with suitable modifications, the usual procedures for subsequent treatment, as described in the following. Small modifications enable such plants to be used in the area of spray forming (Fig. 10.1). The modifications concern mainly a substrate carrier plate, onto which the sputtered metal particles land. The particles are thus either already solid, in the part-liquid condition or still completely liquid. By suitable process control and adjustment of the atomizing conditions to the melt, semi-finished material in the form of bands, pipes or pins can be manufactured (Fig. 10.2). These then, for ex- 246 10 Powder Metallurgically Manufactured Metal Matrix Composites ample, can be processed by extrusion, forging, hot or cold isostatic pressing to the final product" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000406_tmag.2009.2023243-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000406_tmag.2009.2023243-Figure2-1.png", + "caption": "Fig. 2. (a) Lamination geometry (b) tested GO magnetic circuit.", + "texts": [ + " It uses shifted laminations to constitute the Stator Magnetic Circuit (SMC) as shown in Fig. 1. The SMC is assembled shifting each lamination from the previous one with a constant spatial angle . So the easy magnetization direction appears in different parts of the SMC. This principle forces the magnetic flux to pass from one lamination to another, along the axis, in order to satisfy the principle of energy minimization [5]. Several SMCs are built using this principle. The geometry of the laminations is shown in Fig. 2(a). Two different qualities of steel, provided by ThyssenKrupp E.S (TKES), are tested. The main characteristics of these materials are exposed in Table I, where is the peak flux density value and is the core loss amount. Manuscript received March 06, 2009; revised April 16, 2009. Current version published September 18, 2009. Corresponding author: S. Lopez (e-mail: samuellopezruiz@gmail.com). Digital Object Identifier 10.1109/TMAG.2009.2023243 Each device has a primary 100 turn winding , supplied with a sinusoidal voltage . A secondary 100 turn winding , permits to appreciate the magnetic behavior of the magnetic circuit by means of the measurement of the induced voltage (Fig. 2(b)). The SMCs comprise, respectively, 150 and 100 laminations for the GO35 (0.35 mm thickness) and the NO50 (0.5mm thickness), for a total core height of 5 cm. The GO35 is not compared with TKES NO35 (0.35 mm thickness) because this material is highly anisotropic [6], and the aim of this study is to compare the new structure, built with an anisotropic material, with another made with an quasi-isotropic one. Therefore the 0018-9464/$26.00 \u00a9 2009 IEEE results obtained with the NO50 are adapted to enable the estimation of the NO35 isotropic characteristics" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000948_09544070jauto916-Figure13-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000948_09544070jauto916-Figure13-1.png", + "caption": "Fig. 13 FE model for the optimization", + "texts": [ + " The critical location was point A, which still has the shortest life: 2.796104 km. The result is much worse than the previous result, and is more in keeping with the experimental result. There are many factors that may lead to low fatigue life, and the corresponding methods of improvement are also very different (see Table 4). For the transverse beam, optimization of the location of the spot welds can improve the fatigue life of the critical region. The FE model for the optimization is shown in Fig. 13, which includes the transverse beam and the components jointed directly with it. For modelling of the practical conditions, the locations where the transverse beam is jointed to the other components by spot welds were constrained, and unit displacement loads were applied on the foundation of the spring\u2013damper model, as shown in Fig. 13. The optimization objects are the spot welds in regions A and B (see Fig. 13). In the present study, topology optimization by the homogenization method was carried out. The microstructure of the cell with the cavity is introduced into the topology structure in the homogenization method. The initial material density of the cell is set as 1, and then the density of the cell with the cavity, e, can be written as e~1{ 1{a\u00f0 \u00de| 1{b\u00f0 \u00de| 1{c\u00f0 \u00de \u00f07\u00de where a, b, and c are the dimensions of the cavity, as shown in Fig. 14. Obviously, e 5 0 when a 5 b 5 c 5 0, i.e. the cell is empty; e 5 1 when a 5 b 5 c 5 1, i" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003480_icma.2017.8015926-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003480_icma.2017.8015926-Figure3-1.png", + "caption": "Fig. 3 Schematic of a two-phase hybrid stepper motor.", + "texts": [ + " The test result shows that a stepper motor has low dynamic position stiffness, which makes the relatively high linearity meaningless for many industry applications. Another problem for stepper motor is low efficiency[6]. Since maximum phase currents are always applied in open loop operation to avoid losing step, stepper motor has low efficiency and its related problems, e.g. thermal issue. This paper investigates these problems and proposes to solve them from control approach. Fig. 1 Stepper motor linearity test result. Fig. 2 Position error when an external torque applied to the motor. Schematic of a two-phase hybrid stepper motor is shown in Fig. 3. The hybrid stepper motor is made up of certain distinguishable components arranged in a specific layout. The outer layer of the stepper typically contains eight electromagnets spread out evenly around the central rotor wheel. The central rotor is a solid metal piece with several teeth. There are usually 50 teeth in total. The motor works by attracting and repulsing teeth using the electromagnets. In a hybrid stepper motor, the rotor is a permanent magnet and is moved by exciting a single electromagnet (one phase) or a pair of electromagnets (two phase) in turn. The effect is complemented by a minimal reluctance effect, where the rotor is attracted to a position where the space between the teeth and the electromagnet is minimized. The rotor does not line up against all of the electromagnets at the same time, as is shown in Fig. 3. By varying the magnitude and direction of the winding currents, the rotor is continuously attracted in the desired direction. A \"step\" occurs whenever a rotor tooth moves slightly to align itself to an electromagnet tooth. It is possible to decrease the step size of a hybrid stepper motor by using a control logic called micro stepping[7]. Micro stepping involves transitioning between each phase shift. That is, the current references are defined by sinusoidal signals displaced 90 electrical degrees from each other" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000957_1.1559581-Figure13-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000957_1.1559581-Figure13-1.png", + "caption": "Fig. 13 Binding force on a surface", + "texts": [ + " (10) where k is a scalar constant, R(t) is the equation of the curve expressed by parameter t, t i is the parameter of the previous point to which the cursor was attracted, P is a position vector of the cursor and Dt is the difference of parameter t between the new closest point and the previous one Pi . To calculate Dt efficiently, we approximate it using the tangent vector T5R\u0307(t i): Dt5~P2Pi!\"T/uTu2 (11) 3.3 Binding Force on a Surface. The third type of haptic navigation is to keep a cursor on a surface in order to help designers to trace the shape of the surface with a haptic device. The force is represented by the following equation ~see Fig. 13!: F5k~S~ui1Du ,v j1Dv !2P! (12) where k is a scalar constant, S(u ,v) is the equation of the surface expressed in parameter u, and v , ui , and v j are the parameters of the previous closest point, P is a position vector of the cursor, and Du and Dv are the differences of parameters u, v between the new closest point and the previous one. To calculate Du , Dv efficiently, we use a similar approximation method to that used to bind a cursor on a curve. In this case, we have to project the vector P2Pi j onto the tangential vectors Su and Sv at Pi j , and divide them by their norms: Du5~P2Pi j" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001801_j.triboint.2016.11.027-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001801_j.triboint.2016.11.027-Figure3-1.png", + "caption": "Fig. 3. Coordinate system of NSFBs for theoretical model.", + "texts": [ + " Based on the assumption that the excitation motion is a sine function with amplitude \u0394U and excitation frequency v, the energy dissipation of one spring coil caused by Coulomb friction in one periodic can be expressed as \u0394W A f \u0394U= 4ci i i (3) The energy dissipation of one spring coil caused by viscous damping in one period can be expressed as \u0394W b \u03c0v\u0394U=vi i 2 (4) where bi is the equivalent viscous damping coefficient. Thus, the equivalent viscous damping coefficient can be calculated by equating Eq. (3) and Eq. (4), which is given as b \u03c0v A f \u0394U = 4 i i i (5) The dynamic equilibrium equation for one spring coil can be given Fig. 3 shows a schematic view of an NSFB and a rotor with its relevant nomenclature. Ob and Oj represent the bearing center and journal center, respectively. The pressure distribution of the gas film, which is described as an isothermal, isoviscous, inertialess, and compressible flow, can be obtained by solving the dimensionless Reynolds equation: \u239b \u239d\u239c \u239e \u23a0\u239f \u239b \u239d\u239c \u239e \u23a0\u239f\u03b8 p h p \u03b8 z p h p z \u039b p h \u03b8 \u039b p h t \u2202 \u2202 \u2202 \u2202 + \u2202 \u2202 \u2202 \u2202 = \u2202( ) \u2202 + 2 \u03d2 \u2202( ) \u2202 3 3 (7) where \u239b \u239d\u239c \u239e \u23a0\u239fp p p h h C z z R \u039b \u03bc\u03c9 p R C \u03c5 \u03c9 t \u03c5t= , = , = , = 6 , \u03d2 = , = a a 2 (8) where p is the gas film, h is the film thickness, \u03b8 is the circumferential coordinate, z is the axial coordinate, \u039b is the bearing number, and \u03c5 is the excitation frequency" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000231_s00170-017-0346-6-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000231_s00170-017-0346-6-Figure1-1.png", + "caption": "Fig. 1 Bending types. a Conventional or small radius bending and b large radius bending", + "texts": [ + "eywords Bending . Sheet metal . Springback . High-strength steel . Multi-breakage phenomenon Air bending of metal plates remains one of the most popular forming techniques within the sheet metal domain. This forming process is known for its flexibility, since one pair of tooling suffices for different forming angles. The traditional approach of conventional or small radius air bending is described by a three-point bending model (see Fig. 1a). In this model, the punch radius is typically less than the thickness of the formed plate. However, it is traditionally considered that when the forming of a plate occurs with a punch with a larger radius (see Fig. 1b), the progression of the bending mechanism is significantly dissimilar to the traditional three-point bending scheme and the process is referred to as large radius bending. Moreover, if the punch radius is several times larger than the sheet thickness, the effect of the large radius becomes more pronounced. Large radius bending brings about constant changing of the loading scheme during the vertical displacement of the punch, and it is best described by a four-point bending model that originates from the fact that the resulting bend radius at the top is smaller than the punch radius" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002295_b978-0-08-011216-9.50010-9-Figure7.3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002295_b978-0-08-011216-9.50010-9-Figure7.3-1.png", + "caption": "FIG. 7.3", + "texts": [ + " Transverse shear stress at A due to bending is zero since the complementary longitudinal shear is zero at the free surface Resultant direct stress at A = -2865 -99-5 = -2965 lb/in2. Therefore, principal stresses at A are ff1>2 = - ^\u03c8- \u00b1 -^{(-2965)2 + 4 x (1194)2}1/* = -1483 \u00b1 1905, at = +422 lb/in2, \u03c32 = -3388 lb/in2. With the third principal stress zero the maximum shear stress is Tmax = ^ p = 422 - ( -3388) = 1905 lb/in2. If a curved bar is subjected to loading out of the plane of curvature of the bar, then both bending and twisting will occur. The centre-line of the slender curved member of circular cross section shown in Fig. 7.3 is an arc of a circle. It is fixed at one end and carries a concentrated load acting perpendicular to the plane of curvature at the free end. The moments on a section such as B referred to the coordinate axes are Mx = WR sin (\u03c6 - \u0398) (Bending), (7.6) My = 0, (7.7) M2 = T = WR {1 - cos(0 - 0)} (Torsion). (7.8) It is seen from the above equations that both bending moment and torque are functions of the distance along the bar. The direct and shear stresses are determined from the bending and torsion relationships, equations (7" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003415_iros.2014.6942881-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003415_iros.2014.6942881-Figure6-1.png", + "caption": "Fig. 6. Displacement of the movable magnet", + "texts": [ + " Here, to evaluate the performance of the HMPM the hopping velocity of the rover generated by the HMPM just after impact is focused on. The hopping velocity is calculated by using (1). In advance of the calculation, the parameters of the HMPM are defined as follows (see Fig. 5). the distance between the stationary magnets, the inside length of the slide bracket and the thickness of the bracket are denoted by lS, lw and la, respectively. First, we consider the state in which the center of the movable magnet is located on the central position between the stationary magnets, as shown in Fig. 6. The origin of the calculation is set to the upper side of the movable magnet in the initial state, and the displacement of the movable magnet from the origin is denoted by x. The total magnetic force FM(x) acting on the movable magnet is obtained from the following equation, FM(x) = Fa ( lS 2 + lMs 2 \u2212 x ) \u2212Fa ( lS 2 + lMs 2 + x ) . (3) Then, the hopping velocity is derived by following procedures. [procedure 1] In the 3rd state mentioned in the section 2, only the movable magnet is in motion until the magnet reaches to the upper side of the slide bracket" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000940_s1474-6670(17)66076-2-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000940_s1474-6670(17)66076-2-Figure1-1.png", + "caption": "Fig. 1 Int egrated Spacecraft Vertical", + "texts": [ + " This passive control was maintained for the remainder of the 26 months of operation except for brief use of the mass expulsion system for recovery from two anomalous condi tions that developed during the mission. Flight Experiments The RTD-806 Flexible Roll-Up Solar Array, FRUSA, was the most significant experiment affecting the stabilization system. 2 The deployed system included two axes of servo controlled rotation between the arrays and the vehicle. The drum axis would rotate the entire configuration about the long axis of the Agena and the tracking axis could rotate the array at right angles to the drum. A sketch of the fully deployed vehicle is shown in Fig. 1. Mass distribution of the deployed array was such as to offset the principal axis of the system up to 15 degrees from the long axis of the Agena. The primary objective of this experiment was to demonstrate deployment extension and retraction along with its abil ity of sun acquisition, lock on, and track. The experiment would further demonstrate the power capability of the large array and was to demonstrate operation for at least six months. A secondary objective was to demonstrate a 1214 J. J . Rodd en calibrated solar cell and module performance" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002367_iros.2009.5354127-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002367_iros.2009.5354127-Figure5-1.png", + "caption": "Fig. 5 When the Zm axis motor is rotated", + "texts": [ + " With this type of motor bearing, the ends are fixed to the rotation axes, and there is axial support in such a way that rotation is possible at the support-end located on the Z axis of the clamp. There are two bearings installed internally in the support-end, and by aligning the two bearings in the axial direction of the rotation axis, even when an external force is applied to tilt the rotation axis the two bearings act together to support the rotation axis and prevent it from tilting. According to the structure stated above, among the other 2 orthogonal axes, the motor bearing, i.e., the Xm axis motor, can rotate about the Z0 axis (see Fig.5). We focused on and explained the motor bearing supporting the Xm axis motor, but the explanation regarding the motor bearings supporting the Ym axis and Zm axis motors is the same. According to the structure stated thus far, the device has the functionality of a 3-DOF active rotational joint between a pair of links. According to these structural principles, if the Xm, Ym, and Zm axis motors are rotated simultaneously and independently, rotational motion in an arbitrary direction is possible as a composition of motion in the direction of each DOF" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003247_9781119258827.ch7-Figure7.36-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003247_9781119258827.ch7-Figure7.36-1.png", + "caption": "Figure 7.36 Controlled differential. Source: Courtesy of R Ogorkiewicz.", + "texts": [ + " The rear tyres are also subject to relatively high sideslip angles which will increase wear yet further. A number of high\u2010performance road cars use various forms of controllable differential to modify the steering behaviour of the vehicles. Used in conjunction with an electronic control system, the technology is usually referred to as torque vectoring [7.23]. The differentials used are similar in principle to the controlled differentials that have been widely used to steer various military tracked vehicles (see Figure 7.36). Here the driver uses two levers to operate a pair of brakes for left\u2010 and right\u2010hand turns. Like the double differential, they can transfer torque from the slower\u2010 to the faster\u2010running shaft. Effectively, a fixed ratio drive is clutched between an axle shaft and the differential carrier. When the brakes are fully engaged, the system provides just one radius of turn. For a skid steered vehicle this radius is set as a compromise between the minimum tightness of turn and excessive slipping of the brakes that is required for larger radii of turn" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001905_2010-01-0530-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001905_2010-01-0530-Figure5-1.png", + "caption": "Figure 5. Vehicle Lab Model (VLM). (a) Assembled VLM. (b) Elements of VLM.", + "texts": [ + " Figure 2 shows the variation of the sprungmass and the unsprungmass bounce with frequency. The sprungmass bounce natural frequency is 5.2 Hz, while the sprungmass roll natural frequency is about 2 Hz (Figure 3). Figure 4 shows that the sprungmass lateral displacement natural frequency is about 5.25 Hz. There is a noticeable coupling between the roll mode and the lateral displacement mode. The model is designed in accordance to the design parameters of a selected sport utility vehicle. The vehicle lab model has geometry 20.5 cm length, 8.4 cm width, 10 cm height (Figure 5a) and a total mass of 0.73 kg. The model consists of 18 parts (Figure 5b). The ratio between the lab model lengths and the real one is 1:22, while the mass ratio is \u22431:2000. Table in Appendix 2 shows a comparison between the main design parameters of the selected real vehicle and the Vehicle Lab Model (VLM). The model testing includes the impulse force test and turntable test to study the effect of design parameters mentioned in Appendix 2. This test is developed to measure the response of the vehicle model due to an impact force. In this test a new digital analysis was developed to separate the roll and bounce modes in lab" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003316_eej.4391110212-Figure12-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003316_eej.4391110212-Figure12-1.png", + "caption": "Fig. 12. (a) Voltage i n t e r f e r e n c e due t o the s a t u r a t i o n of i n v e r t e r ou tpu t by gene r a l method; (b) t ak ing vd = cons t . t o", + "texts": [], + "surrounding_texts": [ + "4 . Some P r a c t i c a l Problems and Treatments\n4.1 Compensation of v o l t a g e d e v i a t i o n due t o t h e switching dead t i m e\nTo avoid t h e s h o r t c i r c u i t of t h e i n - v e r t e r l e g s , i t is necessary to insert a t i m e delay between t h e off-and-on o p e r a t i o n s of two power dev ices i n one l e g . However, t h i s dead t i m e may no t only induce d e v i a t i o n of t he real ou tpu t v o l t a g e from t h e r e f e r e n c e vo l t age command, bu t a l s o produce some lower harmonic c u r r e n t components of lower o r d e r . Compensation of t h e v o l t a g e d e v i a t i o n is achieved by adding a compensating v o l t a g e vec to r t o t h e r e f e r e n c e v o l t a g e v e c t o r .\nAssuming t h e case where i n v e r t e r ou tpu t c u r r e n t s f low i n t h e d i r e c t i o n shown i n Fig. 9 ( a ) , t h e s p a t i a l c u r r e n t v e c t o r i(lO0) i s i n the d i r e c t i o n shown i n Fig. 9(b) . Therefore , t h e dev ia t ion v e c t o r AV due t o t h e dead t i m e\nATmust be i n t h e d i r e c t i o n oppos i t e t o i(lO0) and has t h e mean v a l u e which is p ropor t iona l t o Emax*ATITSW. Adding t h e v e c t o r -AV t o\nt h e r e fe rence vec to r V and sending t h e re-\ns u l t i n g vec to r t o t h e i n v e r t e r , t h e ou tpu t vo l t age v e c t o r w i l l b e equa l t o V Th i s\nprocedure can b e accomplished by adding AT t o each t i m e d u r a t i o n of nonzero components shown i n Eq. (5) . Therefore , t h e dev ia t ion due t o t h e dead t i m e can be compensated e a s i l y by so f tware , and t h e whole process is not a f f e c t e d by t h e d e t e c t i n g e r r o r of t h e curr e n t , a s r epor t ed by some o t h e r i n v e s t i g a t o r s using hardware compensation [ 1 4 ] .\nE\nE\nREF\nREF'\nFigure 10 shows t h e e f f e c t of compensat i o n , where switching per iod T is 100 us. sw One can see from t h e spectrum t h a t t h e f i f t h - o rde r harmonics of t h e c u r r e n t is reduced from 2.5 t o 0.56 pe rcen t by compensation. The 7th-, llth- and 13th-order harmonics a l s o are reduced.\n4.2 Method of sampling s t a t o r c u r r e n t s\nFor high-performance v e c t o r c o n t r o l , t h e in s t an taneous va lue of t h e motor c u r r e n t s is important. The in s t an taneous measurement of t h e average component of t h e c u r r e n t , which i s hidden behind t h e zigzag waveform of motor c u r r e n t , is one of t h e d i f f i c u l t i e s t o be overcome. The method us ing a d i f f e r e n t i a l equat ion proposed i n [15] h a s high p r e c i s i o n , but it r e q u i r e s complex hardware and h a s a t i m e delay of one sample time.\nIt h a s been r epor t ed [ 6 ] t h a t t h e \"instantaneous\" average v a l u e of t h e c u r r e n t can be sampled a t t h e i n s t a n t when t h e carrier\ncompensating dead t i m e .\nt r i a n g l e wave r eaches i t s peak i n t h e case of t h e subharmonic c o n t r o l method. According t o our s imula t ions , t h e c u r r e n t va lue a t t h e i n s t a n t i nd ica t ed by t h e b l ack t r i a n g l e i n Fig. 3(b) is only equal t o t h e \"instantaneous\" average v a l u e of t h e c u r r e n t , t h a t i s , t h e average c u r r e n t va lue during t h e switching per iod. modulation index of PWM and t h e source f r e - quency are changed widely. This r e l a t i o n ho lds even though t h e\nFigure 11 shows t h e r e s u l t of t h e sampli n g c u r r e n t method we proposed. form on t h e lef t -hand s i d e of Fig. l l ( a ) i s the l o c u s of t h e s p a t i a l c u r r e n t v e c t o r , wh i l e t h e motor is d r iven c o n s t a n t l y by 40-Hz c u r r e n t s w i th switching frequency of 2.5 kHz. Th i s c u r r e n t i s sampled a t the forementioned i n s t a n t and taken i n t o DSP. Then, i t is soon The wave-", + "lOdB lOdB\nb t h Harmonic Current ]-ll::i Harmonic C u r E t 1 120Hzidiv - lOdB - 20dB 1ZOHzldiv -20dB\nr o t a t i n g frequency = 40 Hz).\navoid t h e i n t e r f e r e n c e .\noutput again through a D / A conver t e r . The r e s u l t i n g locus is shown on t h e right-hand s i d e of Fig. l l ( a ) . The D / A process induces a c u r r e n t e r r o r as i n Eq. ( 1 2 ) .\nTo show t h e accuracy of t h e method, t h e d i f f e r e n c e between t h e s e two c u r r e n t s is produced through a l i n e a r adder and then analyzed by a spectrum ana lyze r . Figure l l ( b ) shows t h e spectrum which c o n s i s t s of fundamental frequency component of 40 Hz and higher-order components.\nThe fundamental component induced mainly by t h e D f A process , h a s almost t h e same a m - p l i t u d e as t h a t expected t h e o r e t i c a l l y . Figure l l ( c ) p r e s e n t s t h e comparison of t h e experimental e r r o r va lue with t h e t h e o r e t i c a l one. They are s u b s t a n t i a l l y equal as t h e d r i v i n g frequency i s swept from 10 t o 50 Hz, which confirms t h a t t h e d e t e c t i n g p rec i s ion of t h e method i s s u f f i c i e n t f o r a d i g i t a l c o n t r o l :\n4 . 3 Decoupling d-q i n t e r f e r e n c e due t o t h e vo l t age s a t u r a t i o n\nI n v e c t o r c o n t r o l of an induct ion motor, i n most cases t h e output v o l t a g e of t h e VSI s a t u r a t e s when a comparatively l a r g e s t e p i n torque i s commanded. I n gene ra l , t h e s a t u r a - t i o n i s taken i n t o t h e PWM p a t t e r n by hardware-at t h e s t a t i o n a r y frame fol lowing E q . (13). nously r o t a t i n g frame (d-q a x i s ) , a s shown i n F ig . 12(a) , i t w i l l decrease both V This c o n t r a d i c t s t h e expec ta t ion t h a t vd must\nbe sus t a ined c o n s t a n t l y even i n t h e t r a n s i e n t f o r v e c t o r c o n t r o l , and i t s in f luence becomes even more e x p l i c i t i n t h e lower revolving ranges. a x i s :\nAnalyzing t h e process i n t h e synchro-\nand Vd. 9\nThis is c a l l e d coupling between d-q", + "To decouple the quantities in the d-q axis in the case of saturation, we make the magnitude of V constant and saturate voltage V only, as shown in Fig. 1 2 ( b ) . The process can be implemented easily by software. d 4\n5. Conclusions\nTwo kinds of variable structure control methods for current control are discussed and the following conclusions are drawn from the research herein.\n(1) The overall control procedure is undertaken with a digital signal processor (TMS32010) within a time interval of about 100 u s , which includes an algorithm of the vector control, variable structure control, PWM pattern and others.\n( 2 ) To obtain a fast dynamic response as well as an accurate response in the steady state of the stator currents, the Discontinuous Damping System method is suitable when the switching frequency of the inverter i s higher. In the case of the low switching frequency, a method of combining the software PI controller and hardware hysteresis comparator is better.\n(3) Software implementation of the optimum PWM pattern makes it simpler to compensate the switching dead-time, to decouple d-q axis quantities when the output voltage saturates, and to accurately sample the average value of stator current.\nREFERENCES\n1.\n2.\n3. T. H. Chin, et al. Improvement of current control characteristics using DSP for inverter-induction motor system. 1988 Conf. Rec. Ann. Meeting IAS SOC., I.E.E., Japan, pp. 367-372. Kohlmeier, et al. Highly dynamic fourquadrant ac motor drive with improved power factor and on-line optimized pulse pattern with PROMC. I.E.E.E. Trans. Ind. Applic., Vol. IA-23, p. 1001, 1987. K. Asano, et al. Evaluation and improvement of the current control system with dc brushless motor. Trans. I.E.E., Japan, Vol. 108-D, No. 11, pp. 1033-1040, 1988.\n4 .\n5.\n6.\n7.\n8.\n9.\n10.\n11.\n12.\n13.\n14.\n15.\nT. H. Chin, et al. Improvement of current control for vector-controlled inverter-induction motor system. 1989 Conf. Rec. Ann. Meeting of I.E.E. Japan, No. 1650. S. Ogasawara, et al. Current-controlled PWM inverters having high-speed current response and low harmonic currents! Trans. I.E.E., Japan, Vol. 106-B, No. 2,\nMorita et al. A vectorial control method of an inductor to reduce torque pulsations, Trans. on Semiconductor Transformation, IEE, Japan, Vol. SPC-87-55, p. 11, 1987. H. Nagase, et al. A design method for current control loop on vector control of induction motors. Trans. I.E.E., Japan, Vol. 107-D, No. 12, pp. 1491- 1498, 1987. M. Terashima, et al. Comparison of practical performances between controlled current source and controlled voltage source vector control systems.\n1987. M. Iwasaki, et al. DSP-based vector controlled IM drive system with identification of secondary time constant.\n1987. G. C. Cheng, et al. New method for analyzing inverter-induction motor system and its application to optimization of PWM pattern. Ibid., Vol. 108-D,\nH. W. Vander Broeck, et al. Analysis and realization of a pulsewidth modulator based on voltage space vectors. I.E.E.E. Trans. Ind. Appl., Vol. IA-24,\nM. Ito. Optimization of nonlinearly damped servomechanisms. Trans. I.E.E., Japan, 78-836, pp. 617-622, 1958. D. M. Brod, et al. Current control of VSI-PWM inverters. I.E.E.E. Trans. Ind. Appl., Vol. IA-21, No. 4 , pp. 562- 570, 1985. T. Sukegawa, et al. Fully digital, vector-controlled PWM-VSI-fed ac drives with an inverter dead-time compensation strategy. Conf. Record of 1988 I.E.E.E. Ind. Appl. Annual Meeting, pp. 463- 469. M. T. Okuyama, et al. A high accuracy current component detection method for fully digital vector-controlled PWM ac drives. PESC '88 RECORD, pp. 877-884, April 1988.\npp. 89-96, 1986.\nIbid., Val. 107-D, NO. 2, pp. 183-190,\nIbid., Vol. 107-D, NO. 9, pp. 845-852,\nNO. 11, pp. 1041-1046, 1988.\nNO. 1, pp. 142-150, 1988." + ] + }, + { + "image_filename": "designv6_24_0000810_0470871199.ch11-Figure11.4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000810_0470871199.ch11-Figure11.4-1.png", + "caption": "Fig. 11.4 Measurement of a radiation pattern in a \u03d5 = constant plane. a. Probing a cut in a three-dimensional radiation pattern. b. Practical two-dimensional radiation pattern measurement.", + "texts": [ + " One of the advantages of such a description is that the elements of the scattering matrix can be obtained directly by measurement, using a VNA. How the obtained scattering matrix helps in obtaining the array radiation pattern will be explained after we have discussed the measurement of antenna radiation patterns in general. The radiation power or field pattern of an antenna is a graphical representation of the radiated power or field amplitude of that antenna as a function of direction angles \u03d1 and \u03d5, see figure 11.4. Normally, we are interested in two-dimensional cuts taken from the threedimensional pattern and most often these cuts take the form of the radiated power (field amplitude) as a function of \u03d1 for a constant angle \u03d5 as shown in figure 11.4a for one-eight of a sphere of radius r. Here we assume that the antenna to be evaluated, i.e. the antenna under test (AUT) is placed in the origin of the coordinate system, that this antenna is acting as a transmitting antenna and that on a sphere with radius r, a probe is moved that receives the signal transmitted by the AUT. This received signal is plotted as a function of direction and thus gives the radiation pattern.1 2 A practical implementation of measuring two-dimensional cuts of the threedimensional radiation pattern is shown in figure 11.4b. The AUT is placed on a pedestal and rotated around its axis, while a standard gain antenna (SGA) is placed on a distance r from the AUT and receives the signal transmitted by the AUT while being kept in a fixed position. An SGA is an antenna with a known gain as a function of direction and frequency. Provided that the distance r satisfies the far-field condition that we stated in chapter 3, r \u2265 2D2 \u03bb , (11.5) where D is the largest dimension of AUT and SGA and \u03bb is the used wavelength, the radio equation (chapter 3) gives us the possibility to calculate the gain function from the received power as a function of direction", + " When the far-field condition is met exactly, the spherical wavefront (transmitted by the SGA, assuming that the SGA is smaller than the AUT) deviates from a planar wavefront (over the aperture of the AUT) maximally 22.5\u25e6 [2]. For most measurements, this deviation from a plane wave is acceptable. In the following, we will very briefly outline the most common ways in which antenna radiation patterns are measured nowadays. The obvious way to perform an antenna radiation pattern measurement is to build a set-up as shown in figure 11.4b and make sure that this set-up is constructed in such a way that possible sources of error are reduced to an acceptable level. Since equation (11.5) tells us that for large antennas (large in terms of wavelengths) the far field distance may become considerable, an outdoor antenna range seems to be a good solution. The most important error source is formed then by reflections from ground and surrounding objects. The influence of these error sources can be reduced by elevating the antennas above the ground and possibly above the reflecting surrounding at ground level" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003858_tmag.2006.872493-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003858_tmag.2006.872493-Figure1-1.png", + "caption": "Fig. 1. Small analysis model. (a) General view. (b) Analyzed region (x\u2013z plane).", + "texts": [ + " The analyzed region is classified into two parts, one part near the surface of lamination is subdivided into a fine mesh, and the inner part is modeled by a bulk core having anisotropic conductivity. The effect of shape and gap on the error of eddy current loss is examined systematically. A criterion how to divide into two parts is examined. The effectiveness of the technique is illustrated by comparing the CPU time of the conventional method and the new method. Digital Object Identifier 10.1109/TMAG.2006.872493 Fig. 1 shows a simple model of a core block. The core block is laminated in the -direction. The lamination is composed of 20 sheets. The core is made of grain-oriented silicon steel 35G165 (thickness: 0.35 mm, coating: 0.005 mm, iron loss: 1.65 W/kg at 1.5T, 50 Hz). The conductivity of silicon steel is assumed to be S/m. The average flux density of core is 1.0 T and the frequency is 60 Hz. In order to impress the flux in the -direction, the Dirichlet boundary condition is imposed on the outer surfaces ( mm and mm) of analyzed region shown in Fig. 1. region [gray region in Fig. 1(a)] is analyzed using the 3-D edge-based hexahedral edge element ( method). In order to obtain a steady state periodic result faster, the result of j method is used as an initial value [5]. Almost steady state result can be obtained 48 steps calculation (2 periods). Silicon steel is assumed to be isotropic; only the \u2013 curve in the rolling direction is used. The skin depth is equal to 0.202 mm (at relative permeability Hz). It is required to subdivide three layers mesh within the skin depth region in order to obtain an accurate result" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002203_0010-4655(91)90210-c-Figure18-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002203_0010-4655(91)90210-c-Figure18-1.png", + "caption": "Fig. 18. Effects of substrate thickness on the active impedance locus of a shunt-loaded array (a = 0.52 m, W = 0.26 m, b = 0.39 m, Br = 2.5, XL =0.46 m, h = 0.01 m), d= 0.03 m, f (MHz): [270, 280, 290, 300, 310, 320, 330, 340, 350]; d = 0.05 m, f (MHz): [250, 260, 270, 280, 290, 300, 310, 320, 330, 340, 350]; d = 0.20 m, f (MHz): [90,120, 150, 165, 185, 200, 210, 220, 230, 240, 250, 260, 300,", + "texts": [ + " [3] demon- possible explanation is the following: the structure strate the accuracy of the method even though the is being operated at a frequency far from resoaddition of the shunt creates rapid variations of nance, thus the current on the strip does not the current at the connection points. Figure 15 effectively couple to any of the dominant modes. shows a plot of the strip current for three sub- Consequently, the current decays exponentially to strate thicknesses. Observe, in comparison to the vanish at the end of the strip. This explanation is series-loaded cases shown in fig. 16, the sharp further substantiated by the impedance locus discontinuity occurring at the shunt location. This which is presented in fig. 18. discontinuity is caused by the injection of current The accuracy of the method is also evident in from the shunt, a direct result of the continuity of the calculation of the array active impedance. For current enforced in the present method. For the example, for don 0.05 A and d = 0.20 A, the active series-loaded array, the currents are continuous impedances are respectively 47.7 + 3.llj ~ and 8.4 functions, and for don 0.03A and don 0.05A, they + 51.75j ~l as compared to 48 + 3.8j ~2and 7 + 60j exhibit the typical dominant sinusoidal behavior", + " Figures 17 and 18 show the fig. 17. However, these models are no longer accuimpedance loci for the series-loaded array and for rate for antennas with a substrate as thick as the shunt-loaded array as a function of thickness, don 0.20 A. It was suggested earlier that this To understand these results, it is helpful to review shunt-loaded array behaves as an array of topsome common models assumed when dealing with loaded monopoles. Thus, one would expect (as printed dipoles and microstrip antennas. Usually, seen in fig. 18) the impedance locus to follow the a microstrip antenna with a thin substrate is mod- trend of a dipole rather than that of a typical elled using a parallel RLC circuit. This determines microstrip antenna. Lastly, observe for both that the impedance locus goes from being induc- series-loaded and shunt-loaded arrays the increase tive to being capacitive as frequency is increased, in impedance bandwidth associated with thicker This trend is observed in fig. 18 for the don 0.03 A substrate. 320, 360, 400]. f-P.R. Bayar4 D.H. Schaubert / Infinite arrays of 2-D microstrip structures 383 In fig. 19, the impedance locus is calculated for In the next set of results, a study of input a shunt-loaded array using three values of sub- impedance, strip current and scan variation versus strate permittivity. In general, a higher value of the shunt position is presented. These quantities the substrate\u2019s permittivity produces a lowering of are calculated for various values of XF, with XF the resonant frequency" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003385_tmag.2013.2238897-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003385_tmag.2013.2238897-Figure1-1.png", + "caption": "Fig. 1. 3-D cutaway view of LRPMA.", + "texts": [ + " The magnetic field distributions of the columnar LRPMA, the linear electromagnetic force as well as the rotary electromagnetic torque are all derived and verified by FEM and experiments. Manuscript received October 29, 2012; accepted January 05, 2013. Date of current version July 15, 2013. Corresponding author: P. Jin (e-mail: seueelab_jp@163.com). Color versions of one or more of the figures in this paper are available online at http://ieeexplore.ieee.org. Digital Object Identifier 10.1109/TMAG.2013.2238897 Fig. 1 shows the structure of a LRPMA prototype. It consists of a mover and a stator housing the windings. On the tubular mover, there are forty-eight alternately polarized permanent magnets (PMs), with six stacks in the -direction and eight in the -direction. There are eighteen coils, three stacks in the -direction and six in the -direction. The windings are housed in either an air-cored or iron-cored stator as shown in Fig. 1. A knowledge of the magnetic field distribution produced by the tubular PM mover is fundamental to establishing an accurate model of the LRPMA for design optimization and dynamic modeling. Without loss of generality, an air-cored LRPMA is considered in this paper. Fig. 2 shows the cutaways of the aircored actuator along the linear and rotational directions. Region I is air/winding region and region II is the PM region. In the 3-D model, the soft-magnetic parts are considered to be infinitely permeable, 0018-9464/$31" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002095_cobep.2013.6785220-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002095_cobep.2013.6785220-Figure5-1.png", + "caption": "Fig. 5 \u2013 Flux assignment to stator teeth.", + "texts": [ + " Similarly equations (5) e (6), for an arbitrary rotor excitation fi , the flux in each tooth may be obtained by: ),()( RR base f rj j i i \u03b8\u03b8 \u03a6\u00d7=\u03a6 (8) Where )( rj \u03b8\u03a6 represents the flux in the j stator tooth when the rotor is in the r\u03b8 position, for the arbitrary current applied. Afterwards, the flux in teeth linking phase A windings are added, resulting: \u2211 = \u03a6= m j rjjA NP 1 )( 2 \u03b8\u03bb (9) in which P is the number of poles, N denotes the number of conductors in each j-th stator slot and m represent the number of stator slots forming one full pole pitch. Figure 5 depicts the flux corresponding to the first pole which passes through the phase A frame (linking A1-A1', A2-A2' and A3-A3') for the salient pole synchronous machine. Following the same procedure: \u2211 += \u03a6= m mj rjjB NP 2 1 )( 2 \u03b8\u03bb (10) \u2211 += \u03a6= m mj rjjC NP 3 12 )( 2 \u03b8\u03bb (11) When operating as generator, just the field winding is fed and an external prime mover is used to put the rotor at synchronous speed, creating a rotating magnetic field. During this motion, flux lines linkage the stator coils inducing voltage" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000309_910213-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000309_910213-Figure1-1.png", + "caption": "Figure 1 Variable displacement compressor", + "texts": [], + "surrounding_texts": [ + "Nobuhiro Hariu and Kenji Nakayama Zexel Corp. ABSTRACT Today, time available for production development is shorter and good quality development is required. Using the finite element method is a mean of meeting this requirement, This method of analysis is very adaptable to many fields and can predict static stress, strength, endurance, resonance, dynamic stress, thermal conductivity, convection, radiation, the optimum design, sound, fluid behavior, and electromagnetics in a short period. Furthermore, finite element methods are becoming more practical because of increased speed by computer vector operations and the main memory expansion of computers. This paper outlines an analytical model for studying vibration characteristics including amplitude, modal shapes and stress while a variable displacement compressor is operating. Finite element analysis reveals the dynamic behavior of the mounts and body of this compressor. This dynamic behavior analysis is based on the shaft bearing load which is subjected to unsteady forces, calculated beforehand by dynamic behavior simulation. Based on these loads, eigenvalues and frequency responses, transient dynamic analysis is conducted using simultaneous inverse iteration and the HouseholderIQR methods. (See flow chart.) The following results are obtained from the analysis mentioned below: DYNAMIC BEHAVIOR SIMULATION A numerical method to get the motion (acceleration, velocity, force, moment, etc.) of compressorrotatingparts by considering moment equilibrium equation. 1. Unsteady bearing loads are applied to the shaft area at certain compressor speed, discharge pressure, and suction pressure. 2. Unbalanced forces caused by the rotating parts such as the drive hub and thrust flange are determined. Other forces and moments created by the sliding parts are also determined. William R. Hill Zexel Illinois, Inc. DYNAMIC ANALYSIS FEM model representing the compressor body structure and the simulated dynamic bearing load as an input to the structure yield many significant engineering information. 1. The natural frequency and mode shape of the compressor body a t each vibration mode are determined. 2. The distribution of total energy and the strain energy of the compressor body a t each mode frequency are determined. 3. More precise values of stress, displacement, velocity, acceleration, frequency response, and transient vibration analysis were performed where the total energy and strain energy level are high in contour. Based on the above results, the vibration transmission path and vibration state were reviewed. Through this review, stiffness as well as design optimization of the dynamic behavior of the compressor were obtained. BRIEF DESCRIPTION ABOUT VARIABLE DISPLACEMENT COMPRESSOR Following describes the mechanism of wobble plate type variable displacement compressor. Compressor driving torque from engine is transmitted to the drive shaft by engaged magnetic clutch. The rotating assembly consists of a thrust flange, press fitted on the drive shaft, and a drive hub, hinged to the thrust flange with the link. The thrust bearing (2) in Fig. 2 converts the rotational motion of the drive hub into the reciprocating motion of the wobble plate. This reciprocating motion drives pistons, intake refrigerant gas, and compresses the gas. Displacement volume depends on the angle of the wobble plate. Since the angle of the wobble plate varies the position of piston's bottom dead center, it changes the piston stroke which results in change of displacement volume. The angle of the wobble plate is determined by balance of wobble chamber pressure Pw, cylinder chamber pressure Pd, stroking spring force, destroking spring force, and drive hub's centrifugal force. Eventually, displacement is determined by the wobble plate angle. In case of minimum stroke, the angle of the wobble plate is approximately perpendicular to the shaft, minimizing displacement. Conversely, in the case of full stroke, angle becomes acute and allows longer piston stroke (Fig. 2). Thus, a variable displacement compressor, as the name implies, varies the displacement volume of the compressor. Due to this capability, the variable displacement type compressor has the following advantages compared to the conventional fixed displacement type compressors; 1. Power loss reduction a. Operation under efficent compression ratio b. Reductin of power consumption c. The least influence to the vehicle acceleration 2. Drivability improvement a. Eliminates shock resulted from the clutch onloff operation b. Constant off-grille air temperature in passenger compartment METHOD OF ANALYSIS AND MODELING Figure 3 shows a schematic flow diagram of the analysis. Meshes are generated to prepare for the vibration analysis using FEM. These mesh data are Cviicvdor black' Redinr$oari\"~(21 Tcnck Sllppar Slrell Figure 2 A summary and a designation o f variable displacement comprSessor shown in Fig. 6. By taking the center ofhinge ball as the cos o( sin4 cos U t origin of the coordinate system, the slipper's reciprocat- ,o,ing motion, which is resulted from the drive hub's rota- cos2o(+ sinLo(cos2w t cos+o(+ sinad co; tional motion, is expressed in a unit vector as in equation (1). - - - / -.- F R I C T I C Y / :C;:4-?1:3 I CCEFPIC;~;~; / UET2drl i FIGURE 5 A SUMMARY OF INPUT DRTR RND ANALYI'ZCRL YODEL where o(= wobble angle R= pitch circle radius of wobble ball t,$ t= angle of rotation CL) = angular velocity (Step 2 in Fig. 4) The following equations can be established from the definition of scalar product and vector product if; is defined as a unit vector perpendicular to the wobble plate and the unit vector of each wobble ball center as Wi (subscript i denotes cylinder number). - - S . m=O - - ( 2) m 0-Wi=O - (3) s wi_=cos Xi - (4) Wi x S = E sin fi (5) - - where i is the angle formed by vector Wi and S. i can be expressed as: Coordinate of Wi was obtained by solving the above simultaneous equations (1) -( 5). Furthermore, the coordinate of the piston can be defined by considering the stem (connecting rod) length (Step 2 in Fig. 4). Velocity and acceleration are calculated by differentiating the coordinates with respect to t (Step 3 in Fig. 4). Based on D'Alembert principle, the equilibrium equation of force and moment is obtained by considering forces acting on the cylinder and wobble chamber due to gas pressure, inertia of compressor parts such as piston stems, wobble ball, wobble plate, thrust flange, drive hub, and slipper, and spring forces for the destroking spring and stroking spring (Step 4 in Fig. 4). Unsteady bearing load is obtained by rewriting the system of equations in matrix form. In this paper, calculation was executed under following conditions: Compressor speed = lOOOrpm Discharge pressure (Pd) = 20kgf7cmZ Suction pressure (Ps) = 2.5kgUcmZ 2 Wobble chamber pressure (Pw) = 2.6kgUcm B. Method and procedure of dynamic analysis Mesh data are required to start the vibration analysis with FEM. The mesh data used in this analysis is shown in Fig. 7. These mesh data are composed of cylinder head, valve plate, cylinder block, shell, and shaft. Total 7277 of three dimensional shell and solid elements (6 degree of freedom) are used in the analysis. In addition to the mesh data, material properties of parts (Young's modulus, Poisson's ratio, and density) and experimental constants were provided as the input data. Boundary conditions are given a t the mounting ear (a) - (d) in Fig. 7 since these holes are the attachment point to the engine with the bracket. In this paper, the vibration amplitude of location (e) a t stay was analyzed to see the correlation with the experimental data. The stay is the high stress area when the compressor is subject to vibration. The eigenvalue analysis is performed based on the input data described above. Generally, there are two types of matrices used in finite element eigenvalue analysis. O C l l f COMPRESSOR ASSEMBLED STRUCTURE U S I N G FEY Figure 7 Compressor assembled structure using FEM (mesh data) KX = A q t y p e method for getting eigenvalue and eigenvector by pressure change in cylinder. The figure shows that using stiffness matrix K and mass matrix M.) amplitude on the shell side is higher than that of head .. side since the center of gravity of the drive hub, wobble M2+Cg+KZ,=Otype plate and thrust flange are located closer to the shell side. (A method for getting complex eigenvalue and eigen- By using this result, transient vibration and frequency vector by using stiffness matrix K, mass matrix M, and response were analyzed by way of the finite element damping matrix C) method. In this analysis, the former type of matrix was solved by simultaneous inverse iteration and Householder - QR method. Therefore, calculated eigenvalues are the angular frequency of free vibration. And the energy distribution and energy density represent maximum energy of free vibration. Even though damping term C was not considered in the analysis, quantitative distribution of energy and energy density have significant meanings to obtain optimized stiffness and weight reduction. The compressor vibration is combined with various forces such as gas pulsation, unbalance of thrust flange and drive hub, bearing load, etc., during the operation. And the analysis becomes very complicated. In order to understand this kind of complex vibration, transient dynamic analysis and frequency response analysis are conducted. The transient dynamic analysis was done by adding a damping and forced vibration term to the equation obtained from the eigenvalue analysis. Structural damping was assumed for the damping terms. And forced vibration terms were provided for the bearing load as rotational and reciprocating motion inside of the compressor are transmitted to the compressor housing by way of the bearing. As mentioned earlier, this dynamic bearing force is the resultant force of the force acting on the cylinder chamber and wobble plate chamber due to gas pressure, frictional forces of the piston, inertia of compressor internal parts, and the spring forces of destroking spring and stroking spring. Therefore, the bearing load is regarded as the forced vibration term of the compressors moving parts. This force is transmitted to the compressor outer shell through the bearing. B. Dynamic Analysis Results of eigenvalue analysis and the vibration mode obtained by way of the finite element analysis are shown in Fig. 9 (Mode I), Fig. 10 (Mode 2), and Fig. 11 (Mode 3). The solid and dashed lines in the figures represent the before and after deformation. Fig. 9 shows natural frequency a t mode 1 is 1.46 KHz, and the mode shape exists in the Z direction (aong with the shaft). Note that all parts such as cylinder head, valve plate, cylinder block, shaft, and shell are vibrating along the shaft. From Fig. 10, the natural frequency of mode 2 is 1.63 KHz. This mode shape exists in the Y direction (vertical direction as shown as the arrow in Fig. 10). Each individual part vibrates in the same manner. Fig. 11 shows the nateral frequency (2.76 KHz) of mode 3, and the mode shape vibrates along the X axis (horizontal direction as shown as the arrow in Fig. 11). From the above, vibrational characteristics of the compressor have X, Y, and Z directional components. And these natural frequencies are 1.46, 1.63, and 2.76 KHz. This result agrees with the experimental data. Compressor vibration contributes 0 to 2 KHz of passenger compartment vibration. It is necessary to be cautious on natural frequency and vibrational direction of the engine connected with compressor. Figures 12 to 14 show strain energy distribution a t each vibration mode. Numbers in the figures represent the ratio of strain energy to the maximum strain energy. The strain energy which is energy stored internally is closely related to the status of stress a t each location. In others words, strain energy is a part of vibration energy and stored inside since: RESULT AND DISCUSSION vibration energy = strain energy + kinetic energy A. Dynamic Behavior Simulation The result is shown in Fig. 8 (bearing load). X axis in the figure represents a rotational angle (degree) and Y axis represents a load (kgf). X, Y, and Z component of bearing load is illustrated in the bottom of the figure. These X, Y, and Z components of dynamic forces show that phases are shifted approximately 180 degree. This phase shift can be explained by the inertia force of center of gravity of the thrust flange and drive hub which are also rotating with regard to shaft center. The Z directional force of bearing load exhibits the influence of gas The stain energy is proportional to the stress because of Hook's law. Therefore, the compressor must be stiffened where the strain energy is high. ~ i g . 12 shows strain energy contour of mode 1 (vibration along the shaft). From the figure, high stress energy area is stay (a) and (b), rib (c), and bottom of the shell (dl. Fig. 13 shows strain energy contour of mode 2 (horizontal vibration). The high strain energy area in this case is stay (a). Strain energy contour of mode 3 is shown in Fig. 14 (vertical vibration). In this figure, stay (a) and stay (b) show high strain energy distribution. - 1 CYCLE BEARING LOAD ON SIDE OF SHELL I X-DIRECTION 1 BEARING LOAD ON SIDE OF SHELL (Y-DIRECTION) BEARING LOAD ON SIDE OF HEAD ( X-OIRECTIQN 1 I CYCLE [i BEARING LOAD ON SIDE OF HEAD (Y-DIRECTION1 Y DIRECTION UOBBLE BBLL PISTON ' I CYCLE i i loo zoo 300 I-, BEARING LOAD ON SIDE OF SHELL (Z-DIRECTION) 1000 RPM Pd = 20 kgf/cmt Ps = 25 kgftcrn* Pn * 2 . 6 k g f /cmt Y DIRECTION t Z DIRECTION HINGE BRLC [>(I BEGRING L W D BEARING LOAD ON S I D E OF LERD ON SIDE OF SHAFT / SLIPPER FIGURE 8 RESULTS OF DYNQMIC BEHRVIOR SIMULATION Downloaded from SAE International by University of Edinburgh, Saturday, August 25, 2018 The following summarizes all the modes from the above. The compressor should have sufficient stiffnessin these areas: 1. Stay (a) area on the cylinder block (mode 1 to 3) 2. Stay (b) area on the cylinder block (mode 2) 3. Bottom of shell (dl (mode 1) 4. Rib (c) (mode 1) By stiffening area 1 to 4, vibration reduction can be achieved. In fact,vibration reductionhas been achieved by stiffening stay (a). The analyzed strain energy distribution value is important since it gives the strain energy contour. However, strain energy density (strain energy per volume) should be examined after reviewing the strain energy distribution. The strain energy density was reviewed in this analysis. The compressor shell can be made thinner where the strain energy and strain energy density are lower. In this way, weight reduction can be achieved. Fig. 15 shows the result of transient vibration analysis. The X axis is time and Y axis is acceleration in the figure. Fig. 15-a shows time history plot of acceleration in X direction, Fig. 15-b shows time history plot of acceleration in Y direction, Fig. 15-c shows time history plot of acceleration in Z direction, and Fig. 15-d shows time history plot of acceleration in X direction data obtained from experiment. From Fig. 15-a, the amplitude of acceleration _+ 3200mds2. agrees with the empirical data shown in Fig. 15-d. Fig. 15-b shows that there is howling in the Y direction. This results from the effect of the 5-cylinder compressor. Transient analysis allows for stress, velocity, etc. in this particular direction as well, though they are not exhibited in this paper. Under the unsteady vibration in the Y direction, stress, velocity, etc., were obtained for all locations. In addition, engine vibration was taken into consideration to enhance this transient vibration analysis. axis. Fig. 16ais the Fast Fourier spectrum result ofXand Y directional components of the shell side. Fig. 16b is the Fast Fouier spectrum result ofXandYdirectiona1 components of cylinder head. Fig. 16c shows Fast Fourier spectrum result of the z direction component of the shell side. As you can see in Fig. 16a and Fig. 16b, the 1st order of compressor rotation is dominant and the shell side amplitude is larger than the head side. This is caused by the balance of inertia which is the result from the location of the center of gravity since the center of gravity of the dynamic assembly (i. e. rotating assembly and wobble plate) are located toward shell side from the shaft center. In other words, the 1st order of compressor rotation is the major component of compressor vibration in X and Y direction. In Fig. 16c, the 5th order of compressor rotation (83.3 Hz) is dominant. This implies the compressor is vibrating in the axial direction due to the gas compression. Fig. 17 shows frequency response analysis results based on the data in Fig. 16. The horizontal axis in the figure represnets frequency (Hz), and the vertical axis is acceleration ( m d s ). The a and b in Fig. 17 correspond to location A and B in Fig. 15. The a, b, and c in Fig. 17 represent vibration amplitude of X, Y, and Z directions respectively. From Fig. 17, y and z directional components are larger than the X directional component. The outstanding amplitude peaks are multiple orders of rotation in frequency. Amplitude peaks are increased near the compressor resonance point. In an attempt to take an analytical approach, the Z directional load on the shell is nullified. The amplitude of each direction shown in Fig. 17 are reduced to 1/10 or less. Therefore, Z directional load on shell side is considered as the compressor vibration source in the X and Y directions. The Z directional load of the shell, which is explained by the reaction force of the gas compression, has high possibility for noise and vibration reduction. Future plans include continous analysis of stress and vibration with the expanded frequency range up to 2 KHz. frequency (Hz) (y-direction) 0 I I . , , . . I I I I 1 I 125 250 375 500 frequency (Hz) ,I a Pas t Fourier spectrua of Bearlng load on s i d e of Shell 16.7Hr tx-direc tion) I I I I I I I I 125 250 375 5.00 f requencY (Hz) 1 1 , ; 1 , 0 I I I l I I 125 250 375 500 freqaency (Ha) b) Past Fourter spectrum of Bearing Ioad on side of Head 910213 FREQUENCY RESPONSE 0 1 2 0 2 4 0 3 6 0 4 8 0 6 0 0" + ] + }, + { + "image_filename": "designv6_24_0002686_12.328516-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002686_12.328516-Figure2-1.png", + "caption": "Figure 2. Cutaway view of the SIRTF Observatory", + "texts": [ + " Past cryogenic space telescope designs, such as the Infrared Astronomy Satellite (IRAS) launched in 1981, and the Europeansatellitelnfrared Space Observatory (ISO) launched in 1995, had enclosed both the telescope and the science instruments inside the ciyostat. The vacuum shell resulting from such an arrangement is relatively large and massive. SIRTF uses Warm Launch Architecture[3} where only the science instruments are enclosed in the cryostat; the telescope is mounted externally to the vacuum shell, as shown in Figure 2. The size and mass of the cryostat is thus significantly reduced. This innovation combined with the cold orbit selection (eliminating earth as a heat source) allows lowering of the telescope temperature by passive cooling alone to a value close to the temperature of the outer shell (40 K). The Multiple Instmment Cavity (MIC) is located inside the ciyostat. It is a photon-tight chamber and contains the cold portions of the science instruments and the pointing calibration referencesensor (PCRS, a component of the pointing control system)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003001_taes.2010.5545206-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003001_taes.2010.5545206-Figure3-1.png", + "caption": "Fig. 3. 2D steering angles.", + "texts": [ + " This requirement is satisfied if INS error states are periodically estimated by the Kalman filter that uses GPS carrier phase measurements as filter observables [13]. For multi-directional steering in two dimensions, 1D FFTs for the synthetic and physical phased arrays are combined into a 2D FFT: Sarray(\u03bck,\u00c1l) = M\u00a11X m=0 N\u00a11X n=0 Sm,n exp(j\u00bcm(cos\u03bc0\u00a1 2fk)) \u00a3 exp \u00b3 j \u00bc 2 n(cos\u00c10\u00a14fl) \u00b4 fk = k M , k = 0, : : : , M 2 fl = l N , l = 0, : : : , N 2 \u03bck = arccos \u03bc cos\u03bc0\u00a1 2k M \u00b6 \u00c1l = arccos \u03bc cos\u00c10\u00a1 4l N \u00b6 : (5) In (5), k and l are index numbers of the antenna in the physical and synthetic arrays, respectively; steering angles \u00c1 and \u03bc are shown in Fig. 3. Outputs of the 2D FFT-based steering procedure correspond to postcorrelation complex amplitudes of incoming signals that are received for different antenna steering angles. Particularly, the real part of the FFT spectrum represents the in-phase (I) postcorrelation values and the imaginary part represents the quadrature (Q) postcorrelation values for the antenna steering angles that are defined by FFT frequencies as specified by (5). Hence, the 2D FFT mechanism formulated by (5) simultaneously provides Is and Qs for multiple beam steering angles" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000522_robot.2008.4543240-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000522_robot.2008.4543240-Figure2-1.png", + "caption": "Fig. 2. Hyper plate system with two active DOFs of X and \u0398", + "texts": [ + " Kaneko are with Department of Mechanical Engineering, Osaka University, 2-1 Yamadaoka, Suita, 565-0871, Japan {higashi,mk}@mech.eng.osaka-u.ac.jp K. Utsumi is with Hiroshima University, Higashi-Hiroshima, Japan. eventually lead to a simple manipulation scheme. We show that one of them is a similar arrangement of active DOFs of the plate to that of the pizza manipulation. By applying this arrangement of active DOFs to the plate, we design and develop a dexterous hyper plate system with a high-speed vision, as shown in Fig.2. We show a couple of experiments for manipulation which is far beyond human capability. Organization of Paper: This paper is organized as follows: In Section II, we review related works. In Section III, we show the analytical model and the problem formulation. In Section IV, we discuss the necessary condition for arrangement of active DOFs of the plate to manipulate the object. In Section V, we discuss a sufficient condition and show that one of the best arrangements of active DOFs is similar to that of the pizza manipulation" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001261_ip-h-2.1990.0008-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001261_ip-h-2.1990.0008-Figure1-1.png", + "caption": "Fig. 1 Symmetrical 4-arm antenna with arbitrary arm shape", + "texts": [], + "surrounding_texts": [ + "1. Introduction\nIn recent years the crosspolarisation properties of antenna systems have become more important as polarisation diversity techniques have been introduced to the fields of spaceborne communication, advanced radar and remote sensing.\nIn the case of circular polarisation one possible solution for an optimised polarisation performance is the application of symmetrical 4-arm antennas (e.g. a cross dipole or a 4-arm spiral). It can be shown that these antennas radiate perfect circular polarisation in the on-axis direction if the feeding network allows perfect control of the excitation voltages. In practice however, the quality of the feeding network limits the achievable degree of crosspolarisation discrimination (XPD) .\nA mathematical description of the polarisation purity for arbitrary excitations was provided earlier by the author [l]. Therein the XPD of an antenna system was regarded as the sum of two terms which described the influence of the feeding network and the antenna shape. Using this method for the design procedure for an antenna system, it is possible to choose between the optimisation of the feeding network, the optimisation of the antenna shape or a combination of both with less severe requirements for each system component.\nThis paper gives a brief description of the method and shows the relationship between amplitude and phase errors of the feeding network and the corresponding feeding network factor. To describe the influence of the\nPaper 70468 (Ell), first received 26th April and in revised form 24th July 1989 The author is with the Technische Universitat Berlin, Institut fur Hochfrequenztechnik, Einsteinufer 25, D-1000 Berlin 10, West Germany\nIEE PROCEEDINGS, Vol. 137, Pt. H , No. I , FEBRUARY 1990\nantenna shape, a figure of merit, called the polarisation stability factor, is defined which can easily be measured for real antenna structures.\nThe aim of this paper is the representation of a study into the optimisation of the antenna shape factor for planar spiral antennas. The fundamental radiation properties of these structures are reviewed in order to show basic limits for optimisation. Both calculated and measured values for the shape factor are presented yielding an optimised design for the antenna structure. A comparison of antenna systems with and without optimal antenna shapes is given. To meet the demands of advanced systems, which require improved polarisation performance for various directions, the polarisation is therein considered for an entire half-sphere.\n2 Polarisation properties of symmetrical 4-arm antennas\nis described by a feeding vector V which is composed of the four feeding voltages u l , . . . , u4 and which can be expressed in terms of 4 basic excitation modes VM\n(1)\n45", + "- 5 0 E -\n0 0.25 0.50 0.75 1.00\nCalculated current distribution for a small archimedean spiral Fig. 3 antenna a = 0.02 1; R,, = 0.275 1\nIt is shown in Reference 1 that, from the spherical wave expansion of an arbitrary polarised antenna field, the on-axis X P D of these structures can be written as the sum of two terms for any excitation and any shape of the\n46\nL - XPD z O d 0 - - -\nantenna arms :\n(3) E R H C P X P D = 20 lg -= X P D , + XPD,, I E L H C P I\nThe first term of this equation depends only on the actual feeding voltages and is given by\nIt represents the influence of the excitation configuration and can be called the 'feeding network factor'. Due to the independence of this term on the arm shape, it can be represented in a chart of general validity (Fig. 2).\nAs shown by the diagram, the radiation of pure RHCP for the z-axis can be obtained for any shape of the arms by c1 = - j , corresponding with a perfect excitation V = V3 . In practice however, the unavoidable errors of a real feeding network yield a deviation of the actual excitation from this ideal case. If the amount of the occurring amplitude and phase errors is known or estimable, the possible range of X P D , can be read from the diagram. Alternatively, the chart can be used to find the maximum acceptable feeding network error for a given X P D requirement.\nThe second term in eqn. 3 is given by\nand depends on the reaction of the antenna structure on perfect excitations with the basic feeding modes V 3 = vo (1, - j , - 1, j) and V = uo(l, j , - 1, - j) . This term is independent of the actual excitation and represents the influence of the antenna shape.\nTo obtain the complete X P D of the antenna system both terms X P D , and XPD,, have to be added. If XPD,, vanishes, the X P D requirement for the system can only be met by an optimisation of the feeding network. Alternatively, high values of XPD,, relax the requirements for the feeding network quality. Therefore XPD,, can be called the 'polarisation stability factor', since it determines the sensitivity or stability of the polarisation purity against variations of the feeding network performance. For practical applications a high degree of polarisation stability (XPD, , % 0 dB) is desirable, especially in case of very severe system requirements or commercial applications where the development costs for the feeding network can be important.\nFinally it should be noted that eqn. 3 also offers the possibility for the measurement of both factors X P D , and XPD, , . If one of them vanishes, the other one can simply be evaluated by measuring the X P D of the system.\nIn case of X P D , , the polarisation stability factor has to vanish. This can be achieved using a cross dipole, which reacts on excitations for L H C P and RHCP is the same manner and which shows therefore no polarisation stability. In case of XPD, , , a vanishing feeding network factor can be obtained by feeding two opposing arms without connection of the remaining two. Due to the symmetry of the antenna, the feeding vector is then given by V = uo(l, 0, - 1, 0) corresponding with X P D , = 0 dB. This technique was applied for the measurements of XPD,, in section 4.\n3\nPolarisation stability requires a different reaction of the antenna on excitations for R H C P and LHCP. The most promising antenna shapes are therefore chiral structures\nBasic characteristics of spiral antennas\nIEE PROCEEDINGS, Vol. 137, Pt. H , No. I , FEBRUARY I990" + ] + }, + { + "image_filename": "designv6_24_0001594_9783527646982.ch6-Figure6.12-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001594_9783527646982.ch6-Figure6.12-1.png", + "caption": "Figure 6.12 Finite element meshes of a sample with PI - enhanced stretchable interconnection. Forty percent elongation is applied on one end surface and another end surface is assumed to be a symmetrical plane.", + "texts": [ + " Moreover, these low and high levels formulate a so - called design space, in which the numerical results from the later FEM analysis can be interpolated for response surface calculation. To simplify the notation for the DOE matrix, the low actual level and high actual level are normalized to \u2212 1 and + 1, respectively. It should be noted that 0.1 mm PI width indicates that PI material covers the top and bottom sides of the metal conductor. No side wall coverage of PI material is applied on the metal conductor. Figure 6.12 shows an example of the three - dimensional FEM model with PI - enhanced stretchable interconnect. Taking the advantage of a symmetrical structure, only one unit of stretchable interconnect is modeled. A uniaxial elongation of 40% is applied to the substrate at the one end and the other end is assumed as a symmetrical plane. To obtain information on the sensitivity of the different factors defi ned in Figure 6.11 and Table 6.1 , two levels and four factors full factorial (2 4 ) analysis is performed by means of FEM simulation" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001764_tia.2010.2057398-Figure12-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001764_tia.2010.2057398-Figure12-1.png", + "caption": "Fig. 12. Results from FEA model\u2014stress.", + "texts": [ + " Based on the typical copper and aluminum ultimate stress values, the approximated S\u2013N plots are presented in Fig. 11. Although it is a design practice to achieve 108 cycle expectancy, the engineer can establish higher stress values. For example, the maximum stress allowable in the copper cage is 87 MPa to meet infinite life (108 cycles), but it would be possible to design, considering a maximum value of 145 MPa, if the target is 104 cycles. The better the stress is evaluated, the better the life expectancy can be estimated. A FEA is recommended whenever it is possible (see, for example, Fig. 12). Fig. 13 shows the results of the stress evaluation on the bars. Using the simplified equation (9), the peak stress during the stall can be calculated to determine its life. The steady-state and transient tests were made on laboratory machines for comparison to the calculations. The methodology described in Section II was tested on eight different motors to verify the stator temperature rise on a steady-state condition. The resistance temperature detectors (RTDs) were placed inside the stator winding per National Electrical Manufacturers Association recommendation" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000948_09544070jauto916-Figure14-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000948_09544070jauto916-Figure14-1.png", + "caption": "Fig. 14 Microstructure of the cell", + "texts": [ + " The optimization objects are the spot welds in regions A and B (see Fig. 13). In the present study, topology optimization by the homogenization method was carried out. The microstructure of the cell with the cavity is introduced into the topology structure in the homogenization method. The initial material density of the cell is set as 1, and then the density of the cell with the cavity, e, can be written as e~1{ 1{a\u00f0 \u00de| 1{b\u00f0 \u00de| 1{c\u00f0 \u00de \u00f07\u00de where a, b, and c are the dimensions of the cavity, as shown in Fig. 14. Obviously, e 5 0 when a 5 b 5 c 5 0, i.e. the cell is empty; e 5 1 when a 5 b 5 c 5 1, i.e. the cell is solid. The relationships between cavity sizes, material density, and structure strength are established first, so that the strength of the topology structure (material density) can be changed by adjusting the cavity sizes. Considering the isotropy of the material, the relationship can be written as Eijkl a, b, c\u00f0 \u00de~ekE0 \u00f08\u00de where E0 is the real elastic modulus of the material, k is the punishment factor (k 5 3 is selected), and Eijkl(a, b, c) are the hypothetical characteristics of material, determined by e, k, and E0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001389_6.2007-6221-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001389_6.2007-6221-Figure2-1.png", + "caption": "Figure 2. Left: The apodization function, A(\u03c1), for the NWO starshade, showing several different hypergaussian functions. Center: A 12-petaled, binary starshade with n=6. Right: Illustrations of starshade shapes with different petal numbers and values of n.", + "texts": [ + " ( ) a b a A n \u2265 \u2212 \u2212\u2212= \u03c1\u03c1\u03c1 exp1 (1) where a is the radius of a central region of zero transmission, b is the fall-off radius of the exponential function, n is the index that determines how quickly the hypergaussian function falls with radius, and \u03c1 is the starshade radius. The starshade is a binary mask with flower-like petals that are completely opaque. Each petal provides a fraction, T(\u03c1), of the total transmission of the mask: [ ]2 ( ) 1 ( )T A P \u03c0\u03c1 \u03c1= \u2212 (2) where P is the number of petals. The 1-dimensional apodization function and the 2-dimensional, binary starshade are shown in Fig. 2. The NWO architecture is very flexible2. Depending on the performance that is required and the size of mission that can be afforded, we would choose the starshade, telescope size, and starshade-telescope distance. In this paper we will focus on one example, though all the conclusions are valid for other mission scales. We will look at a 50 m diameter starshade and a 4 m telescope, operating 80,000 km apart. The mass of the starshade for this case would be about 3000 kg. This setup would be able to suppress the starlight by a factor of 10-10 with an IWA of ~50 mas" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002824_ias.2007.4347776-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002824_ias.2007.4347776-Figure4-1.png", + "caption": "Fig. 4. IPM rotor shape design.", + "texts": [ + " SHAPE DESIGN OF IPM ROTOR For the alteration of magnetization distribution, the selected design parameter is not a PM shape or a magnetized shape, but a rotor core which is due to IPM motor. The analysis model has 6 poles and 9 slots. By the analytical method of cogging torque, the position of magnetization dead zone, , and the angle of magnetization dead zone, , for reducing harmonic component of cogging torque are shown in Table . In this paper, the dead zone positions, 1\u03b1 and 2\u03b1 , are selected as design parameters for elimination of cogging torque. Fig. 4 shows the initial model and a shape designed model for elimination of the fundamental and third harmonics frequency of cogging torque. The IPM motor, however, has an intensive saturation in flux barrier of rotor core. Therefore, a numerical analysis such as FEM should be required for a precise analysis of cogging torque. Fig. 5 shows Equi-potential distribution of two models by using 2-D FEM and flux density distribution on airgap is shown in Fig. 6. In shape design model, flux density distribution is large distorted but the cogging torque is much reduced by slot combination" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001189_freq-2012-0100-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001189_freq-2012-0100-Figure3-1.png", + "caption": "Fig. 3: SLC and switch schematic for the GaN-HEMT.", + "texts": [ + " Brought to you by | New York University Bobst Library Technical Services Authenticated Download Date | 6/5/15 2:46 AM This level of power can be used e.g. for base stations, microwave radio links, etc. The HSA is designed to work with RF-power amplifiers for wide bandwidth systems like WCDMA and LTE [2], therefore a 5 MHz wide signal was chosen for the analysis of the response of the HSA. More details about the linear stage can be found in [7]. The schematic for the GaN-HEMT switch is shown in Fig. 3. For this transistor, VT is negative, i.e., the gate potential has to be lower than the potential at the source (VS) to turn the transistor off. VS is connected to the load via the diode shunted inductor and depends on the signal. In order to turn the GaN off, Vg has to be lower than the lowest VS, which is one diode voltage drop below zero. The simulated efficiency was calculated over the full N sample envelope from the total DC power in (1) from each source (Psup-DC), and the power delivered to the PA (Pload-DC) represented by a fixed resistive load of 15 \u2126 and 50 \u2126, in (2) and the efficiency in (3)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001939_2015-01-0088-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001939_2015-01-0088-Figure6-1.png", + "caption": "Figure 6. Gear Shift Fork - Reaction Force Results", + "texts": [ + " 4 shows the loads and boundary conditions - Fork rail constrained except translational Z direction because load applied in this point as a prescribed displacement of 10 mm in Z- direction - this displacement is shared by both the legs. Bottom of the legs constrained in translational Z - direction with using spring element and inside face of the bottom legs constrained in translational X-direction. To maintain the same stiffness for both the legs, the displacements and Reactions should be equal for both the legs. Fig. 5 shows, the Displacement results for Optimized design under operating loading condition - Displacement's are same or acceptable for the both legs. Fig. 6 shows, the Reaction forces results for Optimized design under operating loading condition - Reaction's are same or acceptable for the both legs. Now, this final design is ready for Strength evaluation using Nastran FE Solver. Operating load is applied on each leg constraining all DOF at Fork rail. Stresses in each leg should be below the endurance strength of the material. As the constraint of bulky design already mentioned with two pads, a third pad has been introduced to share the load. But this is done in a phase wise manner in which the third pad will only come to contact after there is specified deflection in the main two legs" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003630_6.1973-885-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003630_6.1973-885-Figure2-1.png", + "caption": "Fig. 2 de Havilland DHC-6 Twin Otter - 300 Series", + "texts": [ + " The Twin Otter is powered by two Pratt and Whitney PT6A-27 turbbprop engines. Maximum cruising speed is 180 KTAS and maximum operating altitude i s 10,000 feet because of an unpressurized cabin. Maximum range is 755 nautical miles; however, the average flight time is less than 45 minutes for normal commuter service in the United States. Approximately 90 of the 366 Twin Otters in service operate in the United States under FAA Part 135 rules. A three-view illustration of the Twin Otter is shown in Figure 2. I I. Design Criteria A survey was conducted of existing data on operational profiles of the DHC-6 Twin Otter. Based on this survey, typical climb, cruise and landing approach conditions were selected for design flight conditions, as tabulated in Table 1 Table 1 Study flight conditions Condition Airspeed (KIAS) Altitude (ft) Gross weight (Ib) Flap position (deg) CG location (% MAC) Glide slope (deg) Climb Cruise Landing Approaci- This study was conducted for the NASA Langley Research Center, Contract NASl-11683, under the direction of D" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002051_ieejjia.1.78-Figure16-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002051_ieejjia.1.78-Figure16-1.png", + "caption": "Fig. 16. Flat linear machine structures", + "texts": [ + " 15(b)), can constitute an important drawback, specially by its action on moving armature, if not considered at the design stage. The attractive force and its ripple can generate vibrations and damage the bearings. In order to reduce cogging force and normal force ripple, due to slotting, the skewing of stator or moving armature can be used. It should be noticed that the developed analytical model can be used to study the skewing of stator or moving armature by the mean of multislice approach. For the reduction of normal force mean value acting on moving armature, double sided sturctures (Fig. 16(b)) can be used (21). Even if this parametric study, which neglects magnetic non-linearities, has shown that slot opening height ho variations have very little impact on machines performances, the effect of magnetic saturation in tooth-tips should be more pronouced for small values of ho. Depending on the localisation of saturated parts of the machine, the effect of magnetic saturation can be globaly seen in different ways. The effect of magnetic saturation in tooth-tips can be globaly seen as an increase of slot opening width wo" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001322_mop.26917-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001322_mop.26917-Figure1-1.png", + "caption": "Figure 1 A Schematic diagram of probe-fed CPA with and without air gap", + "texts": [ + " However, here, we have proposed a very simple expression for computing the ereff as ereff \u00bc ere \u00fe 1 2 \u00fe ere 1 2 1\u00fe 10h pmn \u00f030:2a\u00de=\u00f01:59p\u00de 1 2 (4) where pmn \u00bc 0:1792 for m \u00bc 1 \u00bc 0:2203 for m \u00bc 2 \u00bc 0:2861 for m \u00bc 3 (5) and ere is the relative permittivity of the medium below the patch may be written as ere \u00bc er1er2h er1h2 \u00fe er2h1 (6) Accurate calculation of input impedance of the patch antenna is required for achieving the optimum performance. So, we have proposed a simple, efficient, and improved CAD model based on cavity model analysis of CPA with and without air gaps without using any rigorous and large mathematical steps. The input impedance of a circular patch of radius a with and without air gap seen by the coaxial feed, located at a distance q from the center of the patch shown in Figure 1, may be written as [22] Zin \u00bc R\u00fe j X \u00bc R\u00f0q\u00de 1\u00fe Q1 r f fr;mn fr;mn f 2 \u00fe j XF R\u00f0q\u00deQT f fr;mn fr;mn f 1\u00fe Q2 T f fr;mn fr;mn f 2 2 64 3 75 (7) where, XF, R(q), QT, and fr,mn are the reactance due to coaxial probe, input resistance at resonance when the feed is located at a distance q from the center of the CPA, total quality factor and mode dependent resonant frequency, respectively. XF may be calculated from Ref. 12 as XF \u00bc 3:77 f h c log c p f d0 ffiffiffiffiffi ere p (8) QT consisting of quality factor due to radiation loss (Qr), quality factor due to dielectric loss (Qd), and quality factor due to conductor loss (Qc) given by QT \u00bc 1 Qr \u00fe 1 Qd \u00fe 1 Qc 1 (9) The computation of Qr is more important because it determines the radiation efficiency" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000529_ecace.2019.8679276-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000529_ecace.2019.8679276-Figure3-1.png", + "caption": "Fig. 3. Schematic diagram of SEDCM", + "texts": [ + " GA search for the optimum (minimum) value of a function called cost function or object function from a given set of values known as population. The value of each population is called chromosome. There are three basic processes of GA namely selection, crossover and mutation. The GA uses these operations to find the fittest value based on the fitness value of each chromosome. The fig. 2 [31] shows flow chart of the genetic algorithm. III. MATHEMATICAL MODEL In recent years, DC motors have wide spread use [3] in industries because of its simple characteristics and stability. Fig. 3 represents the diagram of a voltage controlled SEDCM. In SEDCM the field is excited from a separate source voltage. The equation that describes the dynamic behavior of a SEDCM is given by [6], [26], [27]: b a aaa e dt diLRiv (4) Here, Va = applied armature voltage ia = armature current La = armature inductance eb = back emf Tm = motor torque Kt = torque constant Kb = back emf constant J = inertia of the rotor \u03c9 = velocity of the rotor (rad/sec) Again, the back emf is directly proportional to the rotor speed" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000855_0954406220936732-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000855_0954406220936732-Figure3-1.png", + "caption": "Figure 3. The biaxial circular flexure hinge: (a) deformation of biaxial flexure hinge; (b) circular contour curve.", + "texts": [ + " The deformation and force transformation relation of flexible link in reference coordinate system and local coordinate system can be derived from the following formulae Xp \u00bc JXg \u00f03\u00de Fg \u00bc JTFp \u00f04\u00de where J is the pose transformation matrix as follows J \u00bc p gR O3 3 O3 3 p gR \" # \u00bc Rot\u00f0zp, \u2019\u00de O3 3 O3 3 Rot\u00f0zp, \u2019\u00de \u00f05\u00de Based on equations (1) and (5), the flexibility matrix of the flexible link can be obtained at the reference coordinate system {Op}. Xp \u00bc CpFp \u00bc JXg \u00bc JCgFg \u00bc JCgJ TFp \u00f06\u00de Cp \u00bc JCgJ T \u00bc K 1p \u00f07\u00de Stiffness model of biaxial right circular flexure hinge The biaxial right circular flexure hinges are made from a cuboid by vertically cutting its four sides with two symmetric straight cylindrical surfaces, as shown in Figure 3. Thanks to its good positioning accuracy and large three-dimensional rotational deformation, the flexure hinge can be used as flexible spherical joint. The biaxial right circular flexure hinges have the advantage of easy machining and can save the machining cost to some extent. At the same time, compared with the straight circular flexure hinge, less likely to occur damage. When calculating the stiffness model of flexure hinges, it only concerns about the arc part and the rest, which is flexible thin beam, follows the same calculation of flexible link in the previous section. As shown in Figure 3(b), t is the minimum length of squared section of flexure hinges, and r is the radius of arc. Based on the theory of linear elasticity and small deformation hypothesis, the deformation equation of the biaxial circular flexure hinge in the reference coordinate system can be obtained.22 D \u00bc x y z a b c 2 666666664 3 777777775 \u00bc C11 0 0 0 0 0 0 C22 0 0 0 C26 0 0 C33 0 C35 0 0 0 0 C44 0 0 0 0 C53 0 C55 0 0 C62 0 0 0 C66 2 666666664 3 777777775 fx fy fz mx my mz 2 666666664 3 777777775 \u00bc CF \u00f08\u00de In which the elements of flexibility matrix are calculated as follows C11 \u00bc I1 E C22\u00bc C33 \u00bc 12I2 E C44 \u00bc I4 G C55\u00bc C66 \u00bc 12I5 E C26 \u00bc C62 \u00bc 12I3 E C35 \u00bc C53 \u00bc 12I3 E \u00f09\u00de where E and G are the Young\u2019s modulus and shear modulus, I1, I2, I3, I4, and I5 are intermediate variables, which are calculated as I1 \u00bc 1 2r 1 n\u00f0n\u00fe 2\u00de \u00fe 2 \u00f0n\u00f0n\u00fe 2\u00de\u00de\u00f03=2\u00de arctan ffiffiffiffiffiffiffiffiffiffiffi n\u00fe 2 n s " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003635_s12206-009-0321-8-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003635_s12206-009-0321-8-Figure2-1.png", + "caption": "Fig. 2. FOA equivalent model of sub-frame.", + "texts": [ + " \u2020This paper was presented at the 4th Asian Conference on Multibody Dynamics(ACMD2008), Jeju, Korea, August 20-23, 2008. *Corresponding author. Tel.: +82 2 2220 0446, Fax.: +82 2 2293 5070 E-mail address: hhyoo@hanyang.ac.kr \u00a9 KSME & Springer 2009 Fig. 1 shows the finite element model of #-type front sub-frame. The #-type front sub-frame consists of total eight subparts, that is, upper and lower cross member, left and right side members, A and G-point brackets. An equivalent model of the vehicle sub-frame that only consists of a beam element is constructed based on the FOA technique, as shown in Fig. 2. The #-type front sub-frame is a symmetric structure except for front and rear engine mounts. So, by dividing the finite element model into seven pairs of symmetric sides, a total of fourteen areas, the section properties, such as cross-sectional area, area moment of inertia and shear coefficient are extracted for constructing an FOA equivalent model. To verify the accuracy of the FOA model that only consists of beam elements, the modal properties obtained with the model are compared to those obtained with a full scale finite element model" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003168_6.1978-1006-Figure14-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003168_6.1978-1006-Figure14-1.png", + "caption": "Figure 14 Short Scarf Nozzle Weld Joint Configuration", + "texts": [ + " Fitup of the seal cylinder to rim flange can only be accomplished after the nozzle is welded to the rim. The actual fitup to this joint proceeds in a manner which is very similar to the manner discussed previously and needs no further elaboration. Fitup of short scarf nozzles is somewhat easier due to the lesser scarf angle involved with these configurations. In this instance, i t is possible to keep the weld joint in the nominal plane of the exit rim and obtain a weld joint which is weldable using conventional techniques. A typical short scarf nozzle weld joint configuration is shown in Figure 14. W E L O D I Q E C T I O N F O Q P U L S E A R C T I G \\ R I N V O Z Z L E Matching of the diameters for welding of both long scarf and short scarf nozzles is accomplished by either shrinking or expanding the parts until they meet within a tolerance of .002 mismatch on the critical mold line. In the case of the nozzle, the critical mold line is obviously the inner wall. In the case of the seal cylinder, i t is equally obvious that the critical surface is the outer wall. The actual shrinking and expanding is accomplished using hand and foot operated shrinking dnvicss " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003990_jjap.56.06gg01-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003990_jjap.56.06gg01-Figure2-1.png", + "caption": "Fig. 2. (Color online) Schematic process flow for the formation of a SOD film on a half-inch minimal silicon wafer. (a) Wafer cleaning, (b) spincoating of liquid dopant source, and (c) bake at 200 \u00b0C for 5min.", + "texts": [ + " In this work, we systematically investigate the solid source diffusion by SOD in minimal bulk Si and SOI wafers, and fabricate the minimal SOI-CMOS integrated circuits using the developed SOD diffusion technique. We also discuss the effectivity of a low temperature oxidation (LTO) process in boron (B)-diffusion and the top-Si layer thickness reduction of SOI wafers after the SOD diffusions. Finally, we demonstrate the successful operations of the fabricated SOI-CMOS inverters, static random access memory (SRAM) cells and ring oscillators. Figure 2 shows the schematic process flow of an SOD film formation on a minimal Si wafer. As the starting materials, Japanese Journal of Applied Physics 56, 06GG01 (2017) https://doi.org/10.7567/JJAP.56.06GG01 REGULAR PAPER 06GG01-1 \u00a9 2017 The Japan Society of Applied Physics we used (100)-oriented p-type minimal bulk Si wafers with a resistivity of 1\u201310\u03a9 cm. In the SOD film formation, we used liquid-state dopant sources. The compositions of phosphorus (P) and B liquid-state sources are as follows. 5 g of P2O5 and 5" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002573_iceaa.2015.7297178-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002573_iceaa.2015.7297178-Figure1-1.png", + "caption": "Fig. 1: Perspective view of a conventional electromagnetic band gap resonator antenna (ERA) and additional hypothetical planes considered for PCS design process. The value of phase is recorded in each cell by placing a virtual e-field probe at its center. The phase of a cell in PCS-IP (input phase) is denoted by \u03b8 m, n", + "texts": [ + " Although the PCS is proposed here for an ERA, this design principle can be extended to other radiating apertures with phase nonuniformity. The next section of the paper describes the PCS design for an example ERA while Section 3 will discuss the results predicted by the time-domain solver of CST Microwave Studio. 978-1-4799-7806-9/15/$31.00 \u00a92015 IEEE The working of the PCS has been demonstrated by designing a PCS for a conventional ERA. The design process is described below. A conventional ERA resonating at a design frequency of 11 GHz is used as a base antenna for the PCS. A perspective view of the ERA is shown in Fig. 1, where a linear polarized (E ) patch antenna printed on Rogers Ultralam substrate ( 2.4) is used as the feed antenna. An un-printed dielectric slab made out of 4.5 mm thick Rogers TMM4 ( 4.5), with a physical area of 6\u03bb 6\u03bb , is used as the PRS of the ERA where \u03bb is the free-space wavelength at the design frequency. The PRS is placed at a distance of \u03bb /2 from the ground plane. Together, the ground plane and the PRS create a cavity that resonates at the design frequency. Two hypothetical planes are defined along the direction of propagation: (a) a PCSinput plane (PCS-IP) is at around quarter wavelength spacing (~\u03bb /4) above the PRS and it is used to record the input (non-uniform) phase to the PCS and (b) a reference plane (RP) is located at a distance of S = 2\u03bb 54 mm from the PCS-IP, to observe the (hopefully \u201ccorrected\u201d and hence nearly uniform) output phase of the PCS. For phase correction, the inner 4.6\u03bb 4.6\u03bb 126 mm 126 mm regions of the PCS-IP and RP (referred to as correction aperture hereafter) are divided into rows and columns of square cells as shown in Fig. 1. The column number (n) increases along the positive x-axis while the row number (m) increases along the negative y-axis. Each cell (m,n) is identical having a length and a width of \u0394s, where \u0394s 9 mm [12] and hence the correction aperture has 14 rows and 14 columns. while that of a cell in RP (output phase) is denoted by \u0424 m, n . It is desired to have constant value of phase in RP, i.e. m, n: \u0424 m, n \u0424 . The phase delay (\u2206\u0424) required for each cell can be written as: \u2206\u0424 m, n \u03b8 m, n \u0424 [1] This \u2206\u0424 m, n is achieved by introducing dielectric material of appropriate height h m, n between the PCS-IP and RP in each cell" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000718_bmei.2010.5639976-Figure10-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000718_bmei.2010.5639976-Figure10-1.png", + "caption": "Figure 10. Force anlysis after bent in the second articulation", + "texts": [ + " In this way, the head turned quickly while inserting the needle. ItF is parallel to the direction of the head, which thrusts the tissue of human body to move the needle forward. When the head turns, there is a force rhF acting on the head as shown in Fig. 9. rhF is the reacting force generated by the tissue of human body. It resists the turning of the head. Assume that the lengths of the head and the first section are equal. After the head is bent an angle I\u03b2 and the first section is bent an angle II\u03b2 , as shown in Fig. 10, the thrusting force IIF generates a force IIbF on the second articulation: sin cos IIb II II IIt II II F F F F \u03b2 \u03b2 =\u23a7 \u23a8 =\u23a9 . (3) IIbF is in the opposite direction to the reacting force of IbF . They turn the first section together. Also, there is a reacting force 1rF on the first section. It resists the turning of the first section. IItF is parallel to the direction of the first section, which moves the needle forward. It should be noticed that IItF equals to IF . IV. BASIC STEERING ANALYSIS Through the above analysis, it is clear that the trajectory of the proposed needle is mainly influenced by the number of articulations, the tip angle, head and sections\u2019 lengths, and the maximum angles of articulations" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000478_lawp.2015.2496366-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000478_lawp.2015.2496366-Figure5-1.png", + "caption": "Fig. 5. Simulated and measured (a) E-plane and (b) H-plane normalized radiation pattern of the proposed antenna at 2.44 GHz.", + "texts": [ + " The dimensions of this truncated metal plate are 32 \u00d7 25 mm2. It is noted that the area above the clearance region of the ILA has no conductor. The corresponding return loss is also exhibited in Fig. 4. As can be seen, even though the proposed antenna is placed in the close proximity to a human body, it still receives more than 90 percent power, and the impedance matching is barely affected by the proximity loading of components. Next, the radiation characteristics were measured. The normalized radiation pattern of the proposed antenna is shown in Fig. 5. The majority of the electromagnetic (EM) energy radiates towards the broadside direction, and the half-power beamwidth (HPWB) is 85\u00b0 and 94\u00b0 in the E-plane and H-plane, respectively. Such a highly directive performance can be further confirmed by the peak directivity of the antenna (see Fig. 6). The reference ILA alone has a measured peak directivity of around 2.2 dBi whereas the proposed antenna has a measured peak directivity of around 6.3 dBi. These results indicate that little energy is radiated toward wrist tissue once the antenna is placed on a human body" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003325_icma.2006.257698-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003325_icma.2006.257698-Figure1-1.png", + "caption": "Fig. 1 Schematic diagram of five link mechanism.", + "texts": [ + " In this paper, the standard GA is modified and new genetic operators are introduced to improve its performance. Optimal design is an important subject in designing a hybrid mechanism. The main purpose of the paper is to present kinematic analysis and to investigate the optimal dynamic design of hybrid mechanism by deriving its mathematical model. By means of these equations, optimal design for hybrid five bar mechanism is taken by using GA. 1-4244-0466-5/06/$20.00 \u00a92006 IEEE The calculation results of the example are obtained in this system herein. II. HYBRID MECHANISM DESCRIPTION Fig. 1 represents five link mechanism configuration having all revolute joints except one slider on output link. Notations shown in Fig.1 are applied throughout the study. The hybrid mechanism has an adjustable link designed to include a power screw mechanism for converting rotary motion to linear motion by means of a small slider. The slider is assumed to move on a frictionless plane. The crank is driven by a DC motor (the main motor) through a reduction gearbox; the slider is driven by a lead screw coupled the second servomotor (the assist motor). Here the main motor is applied as a constant speed motor, and the constant speed motor profile is applied", + " a, b, d, e link lengths of the mechanism (m) , , , angular displacement of the links (rad) , , , angular velocity of the links (rad/s) , , , angular acceleration of the links (rad/s2) LLL ,, displacement, velocity, and acceleration of the slider on output link (m, m/s, m/s2) P the assist driving force (N) M0 the main driving torque (Nm) M drag torque on output link (Nm) ii YX , positions to the mass centre of the links in fixed coordinates (m) x iS , y iS (i = a, b, e, l) positions to the centre of gravity in local coordinates (m) 1 the engaging angle of main drive for C point in Fig.1 ( ) 2 the engaging angle of assist drive for C point in Fig.1 ( ) III. KINEMATIC ANALYSIS OF HYBRID DRIVING MECHANISM Kinematic analysis of five bar linkage is needed before carrying out derivations for the mathematical model. The mechanism is shown with its position vectors in Fig. 1. The output of system is dependent on two separate motor inputs and the geometry of five bar mechanism. Referring to Fig. 1, the output is given by , and the configuration represents inline mechanism. The output motion profiles , , can be designed for the system as , , . In general, the model of a mechanical system can simply be considered as inertial rigid system. Simplifying assumptions are required while developing the mathematical model. Friction and clearance in all joints are neglected. The mechanism operates in vertical plane and gravity effects are included. Since the hybrid five-bar mechanism has two degrees of freedom ( , L) in Fig. 1, equation (1) and (2) is found as below. dL L dd (1) or 21 ddd (2) where 1d , 2d are tiny displacement of output link caused solely by the main motion d and the assist motion dL respectively. If friction losses in all joints and the change in the kinetic energy are neglected, we can obtain 00 dMPdLdM (3) Since both d and dL are independent variables, M0 and P can be obtained from equations (1) and (3), MM 0 (4) L MP (5) By referring to Fig. 1, the loop closure equations is written as: EDAECDBCAB (6) The above equations may be also written in complex polar notation. By solving vector loop equation (6), angular positions of the link b and e are obtained. Having found the angular displacements of each linkage as , and in the five bar linkage, time derivatives can be taken to find angular velocity and accelerations. Some partial derivatives also can be found. They are definitely needed during the analysis of dynamic model. We can obtain 02 2222 bFFEELCCL (7) where )cos()sin( 00 aedEE , )sin()cos( 00 aeFF , )cos()cos( 000 adCC , 01 cos)( badA , 01 sin)( baB , )arctan(2 1 22 1 2 11 0 eB eBAA Therefore, 2222 FFEEbCCCCL (8) LL, , L and can be found from equation (8) )/()( CCeSSLCCL L )/())sin(( 0 CCeSSLCPFeSPFLda where SPFadSS SPF CPF )sin( ))()sin(( ))()cos(( 0 00 00 angular positions of link b and e are )cos( )sin(arctan 0 0 LEE LFF (10) 03 2 (11) Thus, we can get , , L from equation (10), and , from equation (11)", + " They are functions of design variables. 1) Inequality constraint related to the movable condition of hybrid mechanisms To ensure existence of the hybrid five-bar mechanism, the follow inequality constraints are to be satisfied. 0),,min( 0 0 0 22 22 22 2 dLeba dLeba dbLea Lebda 2 (16) 2) Inequality constraint due to the engaging angle Taking into account simultaneous existence of main and assist motion, the engaging angle of hybrid mechanism correlates with actual velocity direction and trend of C point in Fig. 1. The engaging angles can be obtained by taking velocity synthesis for the main and assist motion at C point, or substituting vector synthesis of tiny displacement (at C point) caused solely by the main and assist motion into (8). To ensure existence of the engaging angle, two constraints are obtained from the engaging angles of the main and assist motion: 0]tan[tan 0]tan[tan max2 max1 (17) where ddeddL ddL LCCddeddLeSSddL eSSddeddLLCCddL // /tan ))(//())(/( ))(//())(/(tan 2 1 where is allowable transmission angle of mechanism" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003581_012042-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003581_012042-Figure3-1.png", + "caption": "Figure 3: Reconfigurable matching network antenna [16]", + "texts": [ + " However, the drawback of this method is complicated antenna design and biasing circuit. 3 1234567890 \u2018\u2019\u201c\u201d Antenna with the reconfigurable matching network has been presented by [16], where three different frequency bands can be switched using two PIN diodes. The antenna is matched at 5.2 GHz by forward biasing the diode on the left and by forward biasing on the right, the antenna is matched to 6.4 GHz. Meanwhile, when neither of two diodes is biased, the antenna is matched at 5.8 GHz. The described matching network is illustrated in Figure 3. This design offers a simple antenna geometry, but gives limited frequency reconfiguration. A reconfigurable cedar-shaped microstrip antenna for wireless applications are presented by [17], invented the switching frequency method by changing the current flow of the patch. The author used is switched to reconfigure four different frequency bands. Slits are introduced at the edge of the cedar shaped patch. The proper configuration of the switches alters the current flow and changes the operating frequencies" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002482_9783527631506-Figure6.21-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002482_9783527631506-Figure6.21-1.png", + "caption": "Figure 6.21 Signal flow graphs for a first-order sampled-data transfer function with biphase clocking: (a) most-connected graph, (b) and (c) possible subgraphs of (a).", + "texts": [ + " Thus z\u22121/2 implies a half-period delay and z\u22121 implies a full-period delay. 3) The basic function to be carried out in SC filtering is to relate the z-transform of a voltage signal to the z-transforms of other voltage signals through a relationship of the form, Vj = aVi + bz\u22121/2 Vk + cz\u22121 Vm, that is, a weighted sum of delayed and direct versions of the voltage signals. In case of a feedback, it is possible to have Vk = Vj, and so on. The requirement in three of the above can be very easily appreciated by considering the SFGs shown in Figures 6.21a to 6.21c. In Figure 6.21a, the input voltage source Vi, with phases (1) and (2) corresponding to the clock signals \u03c61 and \u03c62, is shown. V (1) 1 and V (2) 1 are the voltages at the output of the first OA at the clock phases \u03c61 and \u03c62. A direct gain path between the input voltage and the output voltage is shown by the branch ti(i = 1, 2, . . .) without a hat. This represents a delay-free term like a capacitance ratio Ci/Cj in the SC network. A delay-free edge occurs between voltage nodes operated by the same clock phase (i", + " Since in an SC integrator, the possibility of sample-and-hold operation exists, the delayed gain edges t\u03026 and t\u03027 are inserted between V (1) 1 and V (2) 1 to accommodate this scenario. Since these two edges join the voltage variables at the same node (i.e., output of the first OA), they represent simply terms like z\u22121/2, that is, a delay by half a period. Since no switched path can precede the input source, only one delay edge (i.e., t\u03023) is inserted between V (1) i and V (2) i to take care of the possibility of a sample-and-hold operation at the input voltage source. The direction of this edge is arbitrary. Figure 6.21a is the most complete (i.e., most connected) SFG for a first-order transfer function in the z-domain. This can be verified by the application of Mason\u2019s gains formula (Kuo, 1967) to derive V (1) 1 V (1) i = [ t1 + (t2t6 + t3t4t6 + t3t5) z\u22121 ] [ 1 \u2212 t6t7z\u22121 ] (6.43) We can easily see that the elimination of t6 and/or t7 will reduce the above transfer function to a first-order function. This could be used as a building block to generate, by cascading, a higher-order polynomial function of z\u22121. This provides an example of realizing a finite impulse response (FIR) system" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000011_095765005x31108-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000011_095765005x31108-Figure3-1.png", + "caption": "Fig. 3 Detailed design of a centrifugal blood pump: (a) streamwise static pressure distribution in", + "texts": [ + " As for the boundary conditions, the total pressure was prescribed at the inflow boundary, whereas the mass flow rate was specified at the outlet section of the domain. Inlet duct and volute casing were set as absolute stationary. The surfaces of unshrouded-impeller blade and hub were the relative stationary except for the impeller shroud casing wall, which was modelled as stationary in the absolute frame of reference, i.e. there is a relative motion between the shroud region and the impeller passage. Figure 3 shows the representative flow dynamic phenomena for the design flow rate in a centrifugal blood pump developed by the present design and analysis method. The calculated results, in Fig. 3, have been obtained with water as a fluid because the hydrodynamic characterization of a centrifugal blood pump was performed in vitro using a steadyflowmock circulatory loop with water as the working fluid. Figure 3(a) shows that the computed streamwise static pressure of an impeller gradually increases without any abrupt change in the performance curve. Figures 3(b) to (e) represent the static pressure distribution on the impeller blade surface and the velocity distribution in the impeller passage at midspan. The surface pressure (blade loading) and the relative velocity vectors are ideally distributed around the blade surface for the design flow rate. The impeller blading and the incidence have been well-optimized lest the flow separation should occur in the blade passage. The deterioration of flow structure due to the flow reversal in the impeller blade passage could result in thrombus formations, which become crucial problems in long-term cardiopulmonary operation, so that proper design of a blood pump with a blood stagnation-free structure has to be accomplished. It is observed in Fig. 3(f) that the volute tongue has been properly aligned into the general direction of the flow (absolute velocity) leaving the impeller. Table 1 Specifications of a centrifugal blood pump impeller Inlet tip diameter (mm) 17.61 Inlet hub diameter (mm) 11.90 Exit diameter (mm) 30.00 Exit width (mm) 2.70 Number of blades 5 Tip clearance (mm) 0.30 Length in axial direction (mm) 5.70 Blade angle at inlet tip (8) 75.80 Blade angle at inlet hub (8) 70.55 Blade angle at discharge (8) 67.92 Unshrouded-impeller of which blade angles are measured from meridional plane", + " Power and Energy the impeller; (b) static pressure distribution (blade loading) around the impeller blade surface; (c) relative velocity vector distributions in the impeller; (d) relative velocity vectors at the impeller leading edge; (e) relative velocity vectors at the impeller trailing edge; (f) absolute velocity vector distributions in the volute casing; and (g) overall static pressure distribution in a centrifugal blood pump Proc. IMechE Vol. 219 Part A: J. Power and Energy JPE126 # IMechE 2005 Figure 3(g) illustrates the overall static pressure distribution in a centrifugal blood pump. The available kinetic energy at impeller exit has been uniformly recovered in the volute casing. 3 IN VITRO HYDRAULIC PERFORMANCE ANALYSIS A schematic diagram and photograph of the experimental set-up is presented in Fig. 4. Basic in vitro experimental measurements including pump static pressure differential, discharge flow rate, impeller rotational speed (r/min), and shaft torque have been taken using water in the present study. The wall static pressure rise across the Fig. 3 Continued JPE126 # IMechE 2005 Proc. IMechE Vol. 219 Part A: J. Power and Energy inlet duct and the volute discharge was measured with a differential-pressure transducer. The pump flow rate was measured with a positive displacement flowmeter of the mock circulatory system. Shaft speed was read from the r/min light sensor. In the case of torque measurement, unfortunately, the torque transducer has been broken down by the overloaded power consumption due to mechanical seals and bearings, so that no experimental data for shaft power could be obtained in this study" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000008_icelmach.2018.8506886-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000008_icelmach.2018.8506886-Figure4-1.png", + "caption": "Fig. 4. The stator calculation model of the motor: 1 - stator bearing element with a water jacket, 2 - stator core, 3 - winding, 4 - thermo-conductive resin, R\u017c - substitute thermal resistance of slot insulation, Rs - thermal resistance between the jacket and the stator core", + "texts": [ + " The prototype motor is based on the bearing and braking system from the new Fiat Panda III, while the external dimensions have been limited by the dimensions of the 17-inch rim (Fig.2). The motor consists of two main components: the rotor and the stator. The source of heat in this type of machine are losses in the rotor and in the stator (winding, core). A significant part of the losses are losses in the stator. In order to ensure adequate heat recovery from this element in the stator's supporting structure, a water jacket was made, and the empty space between it and the winding was filled with thermoconductive resin (Fig.4). The efficiency analysis of various structural solutions of the wheel motor cooling systems B. Bedkowski, P. Dukalski., T. Jarek, T. Wolnik W 978-1-5386-2477-7/18/$31.00 \u00a92018 IEEE 995 In order to assess the efficiency of the cooling system, an analysis based on finite element method (FEM) and computer analysis of fluid dynamics (CFD) was performed which is slower than other methods, such as thermal diagrams, and requires high computing power, but its big advantage is that allows analyzing devices of any geometry using any cooling systems. The only limitation of the method are computational capabilities of computer hardware [1],[3],[5],[6]-[9]. In order to conduct a thermal analysis, based on the finite element method, a simplified three-dimensional model of the stator of the engine was developed (Fig.4). The model has been prepared in such a way as to simplify the geometry that does not affect the efficiency of the cooling system and the thermal state of the machine. The applied model includes: an aluminium support element with a water jacket (1), a simplified stator core (2), a simplified winding model (3), a thermally conductive resin filling the space between the winding and the supporting structure (4). In the CFD analysis program, the model (shown in Fig.5) was additionally supplemented with a cooling medium in the water jacket channels" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003536_978-94-009-5063-4_1-Figure10-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003536_978-94-009-5063-4_1-Figure10-1.png", + "caption": "Figure 10 shows a structure consisting of two flexible bars or columns subjected to a load P. Denoting the length of the deformed column by L' we have", + "texts": [], + "surrounding_texts": [ + "A = EI (7T/L)2 , p = I/A (7T/L)2 ( 4) we introduce p* = PIA and N* = N/A (5) The relation between deformation 8/L and 10adP* is readily obtained in the parametric representation 8/ L = sin B - sin {3 2 p* = - sin {3 (1 - cos B/ cos (3). p (6) Prebuckling solution or fundamental equilibrium path: Numerical solutions of Eq. (6) are shown in Fig. 11 for two different structures, both with h/L = 0.1, one with p = 2 X 10-4 and one with p = 4 X 10-4 . With increasing load the stiffness ap* /ao decreases, and at a deformation corresponding to o/L = 0.04 (o/h = 0.4) a maximum occurs in each of the two load displacement curves. Consider the case in which P is a dead weight load. Then as the maximum is reached and the structure cannot carry additional load, it snaps into an inverted position such that the two columns are subjected to tension. The curves in Fig. 11 labeled p = 2 X 10-4 and p = 4 X 10-4 are analogous to the curve OA C in Fig. 7(a). Bifurcation buckling: We notice, however, that if N* > 1, i.e., N> E1(71/L)2 then the columns will buckle. From the post-buckled load-deflection curve for the column (Fig. 9) we see that for reasonably small buckling deflection the column deforms under constant load N* = 1. For all practical purposes, then, we can assume that the axial load for the buckled column is independent of the shortening and given by N* = 1. A secondary or post-buckled equilibrium form with slightly bent bars is represented by * tan B - o/b p = [1 -(tanB-o/b)2] li2 (7) This equilibrium form corresponds to N* = 1 and exists only for values of 0 larger than that for which buckling occurs. At the point of intersection between the fundamental equilibrium path Eq. (6) and the secondary solution represented by Eq. (7), the columns will begin to buckle. This occurs for the more slender columns, p = 2 X 10-4 , at p* = 0.155. The load cannot be increased beyond this value: The structure fails by bifurcation buckling with the columns temporarily bending during the process. For the structure with p = 4 X 10-4 , the point of intersection (bifurcation) occurs beyond the maximum in the primary load displacement curve, indicating that the columns are straight at the inception of snap-through. The behavior represented by the curve OA'B'D in Fig. 11 is analogous to that represented by the curve OABD in Fig. 6; the behavior represented by the curve OBD in Fig. 11 is analogous to that represented by the curve OBD in Fig. 7(a). In bifurcation buckling analysis it is often assumed that nonlinearities and geometrical changes in the pre buckling range can be omitted. As the columns buckle at N* = 1, the critical load of the structure in such a model is Post-bifurcation stability: Consider now a structure that has been slightly modified as shown in Fig. 13 by addition of a linear spring which carries a part of the load. Figure 14 shows load displacement curves for two structures with h/ L = 0.1 and p = 2 X 10-4 . One is without a spring (c = 0) and the other includes a spring with spring constant c = 2.5. The load displacement diagrams with spring, primary and secondary, are obtained by addition of p* SPRING = c(o / L) to the value of p* corre sponding to c = O. With a spring, buckling occurs, of course, at the same value of oil. However, if the spring constant is sufficiently large the slope of the line for the secondary solution becomes positive. The increase of the load in the spring is more than sufficient to com pensate for the decrease in the load carried by the columns. The two-column structure discussed here illustrates the behavior of structures of a more general nature. For example, the curve in Fig. 11 labeled OA'B'D is typical of failure of axially compressed cylindrical shells which buckle plastically and develop nonsymmetric folds after the load has reached its maximum value, as shown in Fig. 6. The curve in Fig. 11 labeled OBD is typical of shallow spherical caps under uniform external pressure in which nonlinear prebuckling effects are important but failure is by nonsymmetric bifurcation buckling. A rather thick cylindrical shell under axial compression deforms axisymmetrically throughout the collapse process. This would be indicated in Fig. 11 by a primary equilibrium path similar in shape to the curve OA 'B'C but lying under it and not intersecting the column bifurcation line at all. A very thin complete spherical shell under uniform external pressure would have a primary equilibrium path that is linear in the prebifurcation range OB. Similarly, very thin cylindrical shells supported in such a way as to prevent early buckling at the ends would display essentially linear prebifurcation behavior. Heavily stiffened shells display behavior similar to that represented by the curve OBD in Fig. 14. After the skin buckles at B, much of the load that was originally carried by it is gradually transferred to the stiffeners as the depths of the buckles grow in the post-bifurcation regimeBD. Loss of stability and imperfections: It is important to notice that in the passing of a maximum in the primary path the structure loses stability. Under a load exceeding this maximum there exists no equilibrium configuration in the immediate neighborhood. The structure is set in motion and the process of buckling is violent. On the other hand, the existence of a bifurcation point indicates only that the equilibrium on the primary path loses its stability. The consequences of this loss of stability on the primary path are not immediately clear. As the equilibrium on the primary path loses its stability at the bifurcation point, the structural behavior is governed by the conditions on the secondary path. Thus a bifurcation point signifies only a load level at which a new deformation pattern begins to develop. It does not necessarily indicate loss of structural stability in a physical sense. The equilibrium on the secondary path may be unstable. This is the case in our example of a two-column structure without the spring. In this case the loss of stability on the primary path results in the loss of stability of the structure. Buckling is violent and in addition the critical load is more or less sensitive to imperfections. With some initial crookedness the columns begin to bend in the pre buckling regime. There is no real bifurcation but the primary path approaches gradually the secondary path for a perfect structure. The behavior of imperfect structures is indicated by the broken lines in Fig. 14. On the other hand, if equilibrium on the secondary path is stable, as in the case with c = 2.5 in Fig. 14, the structure can take additional load beyond the bifurcation point. However, a new deformation pattern, in some sense orthogonal to the pre buckling configuration, begins to develop and the stiffness. of the structure may be considerably reduced." + ] + }, + { + "image_filename": "designv6_24_0002245_aim.2005.1511046-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002245_aim.2005.1511046-Figure5-1.png", + "caption": "Fig. 5 The first resonance mode of the mechanical structure at 378.8Hz.", + "texts": [ + " It appears that the granite makes machine structural element, not only lighter, but oscillation amplitude smaller than iron. This improvement on material dynamic performance makes the granite good for the high speed machine structural material. Based on the design process described before, the mechanical structure of an experimental feed drive system is designed as shown in Fig. 4. The masses of the base, the saddle, and the table are 570 Kg, 140 Kg and 60 Kg, respectively. And the result of the first resonance frequency for mechanical structure is shown in Fig. 5. The objective at the stage of the design is to get a simple but accurate enough model to predict the machine performance by computer simulation. As shown in Fig. 2, the mechanism components of the feed drive system include ball screw, nut, support bearing, linear bearing carriage, coupling and a motor. In order to analyze dynamic performance of the feed drive system, a multi-body model of the feed drive system is established as illustrated in Fig. 6. In this figure, parameters are defined as below: T : motor torque (N-m) mJ : rotor inertia (Kg-m2) sJ : ball screw inertia (Kg-m2) J : total inertia sm JJ tM : table mass (Kg) m : motor shaft angle (rad) tX : table position (m) sK : longitudinal stiffness of ball screw (N/m) bK : longitudinal stiffness of support bearing (N/m) nK : longitudinal stiffness of nut (N/m) allK : overall stiffness of the feed drive system (N/m), 1)/1/1/1( nbs KKK tC : damping coefficient of the guideway (N-s/m) p : lead of ballscrew (m) R : transformation ratio, 2/p (m/rad) In the mechanism components design process, the nut and the support bearings are dependent on the screw diameter" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002797_dscc2011-6167-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002797_dscc2011-6167-Figure5-1.png", + "caption": "Figure 5. FINITE ELEMENT MODEL OF FLEXIBLE TUBE", + "texts": [ + " In addition, due to their highly extended capabilities, commercial FE software tools are widely used by researchers to assist them in analyzing various phenomena. In order to examine the vibration of a catheter under environment disturbance inputs secondary to external loads applied by oscillatory blood flow inside left ventricle, a three dimensional finite element model of flexible catheter is developed in this paper. Geometry of catheter which is simply a thin wall hollow cylinder is discretized into 600 4-node doubly curved shell elements (S4R element in ABAQUS properly models the behavior of thin walled shell membranes). Figure 5 illustrates the meshed model. By assuming that blood flow does not have any influence on rigid section of the catheter, boundary condition is introduced in the model by fixing all degrees of freedom for nodes placed at the root of the flexible part. Our flexible catheter is made of an elastomer, whose stressstrain graph is obtained through an equi-bi-axial tension test [10] depicted in Figure 6. There is a rich literature about mechanical properties of rubber-like materials and their corresponding continuum mechanics constitutive models that describe their characteristics [11]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003388_euma.2000.338648-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003388_euma.2000.338648-Figure4-1.png", + "caption": "Fig. 4. Band pass filter with iris", + "texts": [ + " Also, Fig. 3 shows the modes positions and their weights (ie: Cin defined in (4)). It can be seen from this figure that the presence of the two modes positioned at 15.0 GHz and 15.04 GHz are the cause of the presence of the two pics seen in the S12 curve of Fig. 3. IV. B. Band pass filter with iris This test is performed for studying the influence of the iris dimension and the length of the cavity on the modes positions using the Taylor polynomial mapping. The dimensions of the circuit are shown in Fig. 4. The geometry of the circuit is parameterized with respect to the iris length \"w\" and the cavity length \"L1\". The parameters variation ranges are given in Table I. Twenty modes are needed to calculate the scattering parameters Si, and S12. A six order geometric derivation was used to build the multi-parametric Taylor polynomial model. Once the Taylor mapping is built for each mode and pole, the poles positions are determined for any value of the geometric parameters (LI and w). Fig. 5 and Fig. 6 show the first ten poles positions when one of the geometric parameter is varied inside its variation range" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002809_mop.30723-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002809_mop.30723-Figure1-1.png", + "caption": "FIGURE 1 Geometry of the frequency and pattern reconfigurable saline-water monopole antenna array. [Color figure can be viewed at wileyonlinelibrary.com]", + "texts": [ + " Although most of the previous papers were concerning about the frequency reconfigurable water antennas, the papers on pattern reconfigurable water antennas are relatively few. In this article, a water monopole antenna based on the theory of antenna array is proposed, and frequency agility and pattern reconfigurability can be achieved by the combinations of different antenna elements. The antenna array with simple and compact structure has a high efficiency, and the simulation and measurement results are also in good agreement. The geometry of the proposed frequency and pattern reconfigurable saline-water monopole antenna array is shown in Figure 1. Antenna array is composed of five monopoles numbered from E0 to E4, which are symmetrically placed along the axis. The driven water antenna is designed with a layer of teflon with the relative permittivity 2.1 and the relative permeability 1 to prevent the water and ground forming a shorting circuit and expand the bandwidth. The diameter and height of the layer are Dd and Hd. A polyvinyl chloride tube with the relative permittivity 4 and the relative permeability 1 is used to hold the water, while the inside radius and height of the tube are Ra and Ht, respectively, the tube thickness is t, and the height of the water cylinder of the driven element is Hw" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003936_isapm.1999.757285-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003936_isapm.1999.757285-Figure1-1.png", + "caption": "Figure 1. Schematic of (a) four 292\" x 292\" square alumina, and (b) sixteen 127\" round silicon tiles adhesivebonded to 600\" x 600\" square glass pallets.", + "texts": [ + " Furthermore, it is required that the out-of-plane warpage of the tiles-on-pallet assembly delivered at room temperature conditions to thin film processing be less than 25 pm over the entire area. Other considerations of significance include thermal stability of the adhesive in the 350 \"C to 400 \"C range, corresponding to thin film processing conditions, and the ability to thermally degrade the adhesive in order to detach the tiles. The Pallethation Approach Schematics of the palletized structures of the alumina and silicon substrates for subsequent thin film processing are shown in Figure 1. In this study, commercially available glasses are chosen as pallet materials. This choice was, in part, due to their close CTE match with both silicon and alumina. Furthermore, the glasses are readily available in large sheets, and excellent adhesion properties can be achieved with glasses, which also allow for laser ablation of the adhesive, if required. Relevant thermo-mechanical properties of the substrates and glasses are presented in Table 1. The thicknesses of the tiles and pallets were 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001126_s1474-6670(17)45315-8-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001126_s1474-6670(17)45315-8-Figure3-1.png", + "caption": "Fig. 3. Schematic diagram of impedance control", + "texts": [ + " As long as the motion speed is increased gradually, the system does not have to learn all the parameters at the same time, but simply needs to refine the previous results within a limited range in the whole parameter space, which is often large. The learning parameter space is therefore excited gradually and learning proceeds progressively. This simplifies the learning problem significantly. 3.1 . Task Description In this section, the concept of progressive learning is reduced to a concrete algorithm for high-speed robotic assembly. As shown in Figure 3-(a), the task is simply to insert a ball into a chamfered hole in an x - y plane. The controller is given a nominal trajectory Xd(t) = (Xd(t),Yd(t\u00bbT. How ever, due to the uncertainty inherent in the assem bly process, the hole is not precisely aligned with the trajectory and the ball often collides with a chamfer surface. Compliance control is necessary to cope with the geometric uncertainty of the as sembly process, but is not sufficient for high speed insertion. For example, when the ball approaches a chamfer at high speed and collides with the sur face, the quasi-static controller may not be able to prevent the ball from bouncing on the cham fer surface, which may lead to a failure of inser tion", + " Such dynamic control laws contain a number of parameters to be tuned to a specific task process. It is a difficult job to find the optimal values in a large param eter space, particularly when all the parameters must be learned on-line in real time. It should be noted that a failure in high speed assembly may incur serious damage to the robot as well as to the parts and the environment. Even for the purpose of learning, fatal mistakes must be avoided at all times. Therefore, we intend to apply progressive learning to cope with these difficulties. As shown in Figure 3-(b) , a ball is held with an appropriate impedance. We begin by formulating the impedance control law in accordance with [4] . The motion of the ball of mass mo is governed by the equation of motion given by f+p= mox (1) where x = (x, y)T is the position of the ball with an inertial reference, p = (Px , Py) T is the contact force acting on the ball, and f = (fx , fy)T is the actuator's force to be controlled. The objective of impedance control is to emulate a desired me chanical impedance by controlling actuator force f . The desired dynamics of the system shown in Figure 3-(b) is given by where Xd = (Xd , Yd)T is the nominal trajectory, and M, D and K are the desired inertia, damping and stiffness matrices respectively. The external force p is measured by a force sensor attached to the end effector. From eqs.(I) and (2), we can derive the impedance control law given by To formulate a learning algorithm, we need a means for evaluating control performance. In ac cordance with [11], we will define a performance index function, referred to as a reinforcement function" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002267_iros.1996.569005-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002267_iros.1996.569005-Figure4-1.png", + "caption": "Figure 4: (a ) T h e s u f i c i e n t 3-0 model /representat ion of t h e al ternator cover. (b ) T h e 2-D appearance analysis of a seed feature and its relative posi t ion t o a support ing feature.", + "texts": [ + "1 Object Modeling Submodule Assisted by human operator, the Object Modeling Submodule generates the model of target objects from a set of seed and supporting features in terms of 3-D geometry. It also analyzes the 2-D region appearance characteristics extracted by the Feature Detector/Predictor module and selects regions that are potential to be the prescribed seed/supporting features. Analysis is based on 2-D shape complexity and some other parametric feature descriptions (e.g. mean/standard deviation gray-scale value, etc.) as far as the task require. In our experiment, each alternator cover is sufficiently modeled/represented by the five circular regions as shown in Fig. 4(a). The large circular bearing hole at the center is defined as the seed feature and the remaining four small screw holes on the perimeter of the bearing hole are defined as the supporting features. The relative location of a screw hole to the bearing hole and the relative position of a screw hole to another screw hole are defined in 3-D geometry. We define the center of the seed feature p as the object origin and the normal vector n perpendicular to the plane of the seed feature (or the plane created by the supporting features) as the object orientation", + " For the type of seed features we used, several necessary parameters about ellipses were also obtained, and they are based on the facts about ellipses. Let p be the center of a seed feature C having the normal vector n with radius T , D be the projection of seed feature C onto the camera image plane, q o be the centroid of region D, and q be all the points on the boundary of region D. Let the moment of inertia M of the region D be computed with respect to the centroid of D [14]. The distance of q o to each point q normalized by its moment of inertia is computed as (1) T d = J(q - no) M-l(q - 9 0 ) . As shown in Fig. 4(b), if the region D is an ellipse, then d should be constant over all perimeter points q. In our experiment, to accept a region as an ellipse, the standard deviation Ud of the distances from q o to all q is computed, and then tested as in Eq. (2). Uthreshold is obtained through simulations and trials of various centroid locations p and normal vector n angle rotations in the workspace. Empirically, it is found that the test g l h r e s h o l d 5 0.1 unit performs satisfactorily. Furthermore, we also obtained the analysis of the locations of supporting features relative to the centroid of their associated seed feature", + " The definition of the objective function is as follows: For each set A of four pairs of supporting features, say, compute the 3-D positions of the centroids of these supporting features. Let pi be the 3-D positions of the centroid of the supporting feature estimated from q i l e f t and qir ight . The objective function to be minimized for a set A is where Cjj is the 3-D distance between the centroids of two supporting features i and j which is obtained from the 3-D model of the object (e.g. in Fig. 4(a), Cij is either fir,,pport or 2rsupport) , and FA is bounded by some value Fmax. This bound is derived empirically and is highly specific to the physical system. To find an optimal correspondence, the algorithm performs an exhaustive search over all sets of possible correspondences. Finally, after the optimal correspondence is established, the algorithm estimates the pose of the object with respect to seed features and supporting features. If the supporting features are optimized with the objective function case of FA < F,, over both image frames, then the 3-D position of the supporting features are estimated", + " This substitute is then used to compute the pose of the object (i.e. vectors p and n). The result of this module is shown in Fig. 2, where the accuracy of the pose estimation is shown by reprojecting the seed and supporting features onto the original images. The result of the 3-D pose estimation is then passed to the Manipulator Interface Module ( M I ) as in Fig. 3 where motion-path-plans are generated for grasping. 7 Experimental Results In our experiments, several types of alternator covers that share the common model shown in Fig. 4 were used as target objects. The objects were randomly cluttered, with possibility of being upside down. The radius of the bearing hole is 15.0\" and the distance between the origin of the bearing hole to each neighboring screw hole is 22.5\" (i.e. r and rsvpport as in Fig. 4(a)). As a testbed we used a gripper-mounted camera on a PUMA 700 manipulator. Two gripper/camera positions were chosen as the left and right viewpoints with a distance of 243.5\" and vergence angle of 20\". The approximate distance from the camera locations to the objects is 350.0\". For each trial, stereo-pair images of 512 x 480 pixels are digitized, and processed individually up to the SCPE module as shown in Fig. 3. On our SUN Sparc 1000 server, the average processing time (i.e. until completion of pose-estimations) is approximately 1" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002643_60.790876-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002643_60.790876-Figure3-1.png", + "caption": "Fig. 3 Radial cross-section for c=O and 5=1 and appropriate curve of mutual inductances per unit of length, between stator coil 1-6' and rotor loop 1 (between bars 1 and 2)", + "texts": [ + " When the rotor bars are skewed uniformly (linear skewing) the mutual inductance between the stator coil i and rotor loop j can be defined, per unit of length, in the following manner, (15) where 6 represents the length along the axial direction of the rotor, {e(O,l), and y is the mechanical angle of skewing in radians. When 5 increases and the skewing of the rotor bars is in a positive direction of 8, then the sign in brackets in equation (15) is positive. Equation (15) can be interpreted as follows: in any radial cross-section of the machine, mutual inductance, per unit of length, between stator coil i and rotor loopj, has the same shape, but is displaced in space, as shown in Fig. 3. Total mutual inductance between stator coil i and rotor loopj, for skewed rotor, is determined when equation (15) is integrated along the axial axis of the rotor: -112 It is obvious from (16) that the total mutual inductance between the stator coil and the rotor loop, obtained by the described method, becomes a function of one variable. IV. EXAMPLE OF CALCULATION OF INDUCTANCE The proposed method is applied to a four pole, three phase induction machine whose cross-section is shown in Fig.4. The stator phase windings are symmetrical and consist of eight shorted pitch coils, connected in series" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000934_j.still.2017.12.011-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000934_j.still.2017.12.011-Figure3-1.png", + "caption": "Fig. 3. Yield function of the extended Drucker\u2013Prager model.", + "texts": [ + " In this work, the soil was assumed as a continuum non-linear elasticplastic material which exhibited material hardening. The extended Drucker-Prager model has been used by several researchers to simulate the mechanical behavior of soil during the soil-tool interaction process (Bentaher et al., 2013; Li et al., 2013; Naderi-Boldaji et al., 2013; Ibrahmi et al., 2015; Tagar et al., 2015; He et al., 2016; Li et al., 2017). This model is suitable to model frictional materials, which are typically granular-like soils and rock (Simulia, 2013). The yield function of the extended Drucker-Prager model, as shown in Fig. 3, is defined as: = \u2212 \u2212 =F t p \u03b2 dtan 0d i (7) = + \u2212 \u2212t q K K r q2 [1 1 (1 1 )( ) ]d 3 (8) = \u2212 + +p \u03c3 \u03c3 \u03c31 3 ( )1 2 3 (9) where F is the yield function, td is the deviatoric stress, p is the equivalent pressure stress, \u03b2 is the slope of the linear yield surface and di is the td-axis intercept in the p\u2212 td plane, q is the Mises equivalent stress, r is the third invariant of deviatoric stress, K is the ratio of the yield stress in triaxial tension to the yield stress in triaxial compression and thus controls the dependence of the yield surface on the value of the intermediate principal stress" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003715_aero.2004.1368030-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003715_aero.2004.1368030-Figure2-1.png", + "caption": "Figure 2 -Node A And B, Begin Negotiating A Docking Maneuver. Both carry their own respective hardware and software. Each node runs their systems independently, meeting hard real time deadlines locally. Communication is limited to soft real time ranging and azimuth data.", + "texts": [], + "surrounding_texts": [ + "CONTROL SYSTEMS The Virtual Machine Architecture for Control Systems (VMACS) is intended to provide an operating platform for high-level autonomous agent control applications and intervehicle cooperation. VMACS creates a unified logical system image from distributed independent vehicle node processors. Low-level system tasks such as interrupts, memory unification, timing, synchronization, messaging, resource management, and configuration management are handled by VMACS, presenting to the controlling autonomous agent applications a vastly simplified and more idealized operating environment. Like other virtual machines, VMACS exports to applications a 'virtual' processing platform complete with processor, memory, interrupts, and other resources. Where VMACS distinguishes itself, is that it is explicitly built to support distributed, real time, controls applications. VMACS offers virtual CPU instructions for matrix math operations, compound memory access, extremely fast process context switches, and intelligent inter-process communications. Participating nodes in a VMACS system are mapped into a common global resource space. Nodes can interface to each other as if they are operating on a single hardwired multiprocessor computer, . removing the need for explicit wireless network coordination. Distributed shared memory, distributed process locks, and process migration are all functions of VMACS, which reduces the complexity of the high-level autonomous agent applications. VMACS is also responsible for implementing the most important requirement identified, a long-term plug-and-play upgrade path. This is accomplished by isolating the applications that run on the system from the details of the flying hardware. The abstraction is called Platform Neutrality. Platform Neutrality The primary purpose of VMACS is to provide the autonomous agent applications, which control the system, with independence from the processing hardware. This capability is justified in re-configurable satellites due to the dynamic nature of the system. Over the long term, as nodes of the system are upgraded, new nodes are added and older nodes retired, the coordinating agents much incorporate the changes into the operational system. New nodes will be built using new processors, different memory architectures, and better components. Because the original system developers do not have visibility into all possible future upgrades, building the system on an abstract representation of the hardware leaves a path for integration of the upgrades without full prior knowledge of how the upgrade will work. In the short-term, high-level autonomous agent software, which acts as a spacecraft and mission control system, will not always be able to run on a fixed platform. An example is the attitude control subsystem of two docking vehicles. While in free flight, each vehicle would typically have its own attitude subroutines executing separately, each on its own respective processor. Once docking is complete, a unified attitude control system must be created. Even though the momentum hardware does not change, the controlling software must now take into account the combined vehicle dynamics, capabilities, and it's distributed momentum effectors. With VMACS, the actual flight software for the momentum subsystem of one vehicle could migrate to the other vehicle and execute from within the new processing environment. The migration of the software takes with it the control software for the unique hardware aboard the first vehicle, including dynamics and capabilities. With this capability, the two attitude subsystems could condense, for purposes of real-time performance, without having to build multiple binary versions of the control algorithm for every dock-able node combination. 3 6. HARD REAL TIME SUPPORT Communication Channel. VMACS exposes Node B resources to Node A through the distributed computing model. The high speed link enables a dedicated link which can support hard real time communication. Applications are free to move from node to node. VMACS treats the processing elements of each node as hot plug-able members of the same multiprocessor computer. VMACS Traditional real time software frameworks offer a path to solve fixed hard real time problems. Hard real time in this case implies that a deadline may not he missed, and that a missed deadline is the same as an error. Common real time operating systems today provide an application programming interface (API) which allows the creation of tasks that must meet a given deadline plan. Changing the deadline plan often means stopping the system, uploading new compiled software, and restarting the system. Autonomous re-confgurable spacecraft missions will not always he able to tolerate an interruption of service to update the nmning real time processes. The system will require continuous operation of some real time tasks, even during transition from one controller to another. The least invasive way to accomplish this task is to isolate the running controller from the starting controller. With complete isolation in place, the system can initialize the new controller with the data from the running system. Once initialization of the new controller is complete, the output of the first controller is ignored, and the new controllers output is instead issued to the effector. When the new controller is in operation, the original controller is descoped, stopped, and its resources can reclaimed. attempts to resemble a classic Non-Uniform Memory Architecture machine. VMACS offers a graceful method of accomplishing this handoff scenario. Each process in VMACS can be fully time and space partitioned from other processes; where partitioning implies that the behavior of one process cannot APPLICATIONS affect the behavior of another. In addition, VMACS 5. OPTIMIZATION FOR CONTROL SYSTEMS VMACS offers an opportunity to tune the instruction set for controls applications. Through advanced hyte-code instructions, specialized functions that are common to all control systems can he created. These specialized functions are similar to the media and vector extensions of the common microprocessor market. Specialized mathematical functions, including matrix operations and even simple PID controllers can be exposed to the application level as native instructions or instructions for a dedicated coprocessor. These instructions may take advantage of the custom hardware VMACS executes on, or he processed by VMACS itself with built-in algorithms. supports a dynamic process resource database (PRD). A process resource database is used in advanced real time operating systems to store information about a processes allotted memory and time resources. All mainstream real time operating systems currently implement a static PRD, where changes are not allowed unless the whole system is reloaded with a new binary image. With proper protection of a dynamic PRD, VMACS can allow running processes to manipulate the PRD at run time, without stopping other tasks. Processes may register their inputs and outputs with the PRD for inclusion in a real time control hand off scheme. Due to the built-in partitioning of tasks in VMACS and the complete introspection a virtual machine offers, planned and unplanned task hand offs is possible. If a capability is needed that doesn't exist on the executing local platform or is available in hardware elsewhere in the system, a remote procedure call to another VMACS virtual machine may he made. This capability to distribute compute tasking at an instruction level simplifies the application Programmer job by offloading the task of distributing the problem from the application to VMACS. Unplanned hand offs themselves offer a new capability not normally thought in outside of business Sewer systems. Should a node fail in a VMACS system, which is responsible for controlling a given function, vMACS can fail over to a secondary controller gracefully. As all VMACS nodes are sharing a common resource pool, when a controller falls offline another controller may catch the exception and takeover the inputs and outputs of the failed node. With U 0 globally shared and backup copies of the controller spread system-wide, critical control systems can be made to be fail-over or fail-safe capable. 4 7. TIGHT COUPLING OF DEVELOPMENT TOOLS, VIRTUAL MACHINE, AND APPLICATIONS Our investigations into the cost of software development has lead to the consideration of several rapid application development tools. These tools are based on commercially available, fifth generation graphical programming languages. Each of these tools allows the controls and software engineering staff to develop high level representations of system controllers in a precise, domain specific way. From these tools, C or Ada code is created and grafted into the fmal product operating system for flight. In attempting to capture these tools into the existing process framework, it has become clear that these tools lack a capacity for introspection into the final composite flight code and its subsequent operation. While this issue is resolvable, a closer look at the cause reveals that the problem is rooted in the way the abstraction from the highest block to the lowest register was implemented. Vendors of fifth generation graphical programming languages are all laden with the same constraint; the underlying hardware is always different. The output of most of these tools is limited to a text based, fourth generation language. This decision allows the output of the tool to be compiled to any platform with a compliant compiler. Introspection issues arise when trying to gain visibility into the operational system with tools that are built to work with forth generation tools. While the debugger can let the user see data and functional flow, the debugger cannot translate that into the original block model of the fifth generation tool. The high-level controls engineer is unable to see operation in the context of the block diagram, and the software engineer is required to wade through auto-coded source. VMACS can help ease the introspection issue in several ways. The primary method is a simple virtue of the virtual machine itself. Unlike current tools, which must remain hardware agnostic, a tool targeting VMACS is guaranteed a stable virtual 'hardware' platform. The fifth generation tools could skip the intermediate steps of producing C, Ada, and assembler and target the bytecode of VMACS directly. This removes several tools from the build tool chain, and allows the controls engineer to gain in-circuit class visibility into the operation of the code from the native block representation tool. The stability of the instruction set also has implications in the quality of the build tools themselves. As the target instruction set architecture (ISA) is not a moving target, the tools vendors can focus on improving the output of the tool. This reduces the resources needed to keep up with constantly changing architectures. As more systems are implemented with VMACS, despite differences in mission or underlying hardware, the cost of the tool chain can be spread out and justify further refinement. An example of this concept can be seen in the Java tools from Sun Microsystems, Inc. Java is the name for three different technologies, the Java virtual machine, the Java programming language, and the Java object libraries. The Java system is able to run on many different hardware and software platforms, including Microsoft Windows on Intel processors, Macintosh OS X on ISM PowerPC processors, and Sun's own Solaris OS on SPARC processors. This abstraction of the host environment keeps the development tools blissfully unaware of the lower level complexity. As a result, industry has published several major implementations of the Java byte-code compiler, some run on Java itself to allow developers to use any platform to build Java code for any other platform. Java's rapid and widespread adoption is a validation of the 'write once, run anywhere' virtue of virtual machines, including VMACS. Sun\u2019s Java and Microsoft\u2019s .NET technologies are two virtual machine based technologies. Each is built with the goal to provide binary portability for World Wide Web applications. In the area of business computing, Java is heavily used to implement server side services while .NET is still growing in the enterprise. A Java working group recently released a \u2018real-time\u2019 specification to extend the Java language with a soft real time applicationprogramming interface. To date, there are no implementations of a fully compliant hard real time Java interpreter. Neither Java nor .NET offer the critical features beyond virtualization that a re-configurahle satellite mission will require. The real time specification for Java is able to provide processes with pseudo-deterministic execution, hut neither offer the time and space portioning VMACS proposes. In addition, the instruction sets for both virtual machines are oriented to business and Web applications, control applications with their specialized requirements, are not optimally supported. 8. CONCLUS~ON The Virtual Machine Architecture for Controls Systems presented in this paper specializes in the needs of distributed real time controls systems for re-configurable spacecraft. By adopting a platform based on VMACS, reconfigurahle spacecraft are enabled with a viable, long term development and operational foundation. Separation of the computing and communication hardware from the highest level control applications is critical to sustain a system with heterogeneous node elements over several generations of upgrades. While several technologies exist for this abstraction, VMACS offers the hest solution by simplifying the development process, increasing reuse of code through target stability, and higher order abstraction. BIOGRAPHY Thom Kreider is a Senior Project Engineer in the Advanced Technology Department at Honeywell DSES-Glendale. He specializes in hard real time and operating systems software architecture for electronic, mechanical, power, and explosivehigh- energv applications. The last three years he has focused on aerospace applications with emphasis on advanced technologv for sofhvare systems and total systems engineering; several related patents are pending. He holds an B.S. in Computer Science from Park University. Jamie Ross is a Senior Principal Engineer in the Advanced Technology Department at Honeywell DSES-Glendale. He holds a BSc. in Astronautical Engineering from MIT, BSc in Computer Science for Univ. of Mryland and a MSc in b4echanical Aerospace Engineering from ASU. His previous work involved development of Honeywelrs ERADS autonomous ultraviolet attitude determination and navigation sensor as well as systems engineering work on the IRIDIUWM satellite constellation system. His current research interests include engineering information systems. distributed computing, meta-control systems dmign, and dark beer. 6" + ] + }, + { + "image_filename": "designv6_24_0000255_0094-5765(80)90086-7-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000255_0094-5765(80)90086-7-Figure3-1.png", + "caption": "Fig. 3. ERDSAT in launch configuration.", + "texts": [], + "surrounding_texts": [ + "transmitted in S-band (SSA-Link) and the sensor data in K,-band (KSA-Link). For satellite operations a data link to the European Control Center will be established via an Intelsat link. The sensor data will be recorded on HDDTs at the TDRS-receiving station and then conventionally mailed to European data centers.\nModel payload The first model payload instrument is a Modular Optoelectronic Multispectral Scanner (MOMS) working in a pushbroom mode with linear CCD (charge coupled device) arrays in the focal plane. In total, 9 spectral channels are envisaged covering a swathwidth of approx. 200 kin. The spectral bandwidths of the different channels are only preliminary and they are assumed to be:\nChannel 1 0.53-0.57/zm Channel 2 0.584).62/~ m Channel 3 0.63-0.67/zm Channel 4 0.76-0.95 #m Channel 5 1.6-1.7/xm Channel 6 2.1-2.3/xm Channel 7 10.5-12.5/~m Channels 8 panchromatic in the visible region.\nand 9\nThe spatialresolution is approx. 30 m in the 4 visible channels, 50 m in the 2 medium IR channels and 100m in the thermal IR channel. In addition, 2 panchromatic channels in the visible spectral range provide stereo images with decreased swathwidth but with an altitude resolution of 15-20 m. In a future payload definition phase attention should be paid to the spatial resolution of the several spectral channels because the different spatial resolutions should be integer multiples from each other. This will ease the image data processing procedure. The HgCdTe detector arrays of the TIR channel are cooled to 85-100 K with a two-stage radiation cooler.\nThe other payload instrument is a digital Synthetic Aperture Radar (SAR). The antenna dimensions are approx. 12m \u00d7 1 .2mx0,12m and the minimum depression angle of the SAR antenna is adjustable from 71 \u00b0 to 45 \u00b0. It is a coherent pulse radar and it is assumed to be operated at a frequency of either 5.3 or 9.6 GHz. The SAR with 71 \u00b0 depression angle and 5.3 GHz seems to be a good compromise for ocean and land observations including soil moisture determination. For this SAR the swath width is about loo km, the spatial resolution is loom x 100 m; the DC power required is 1000 W nominal. This example has been used as the baseline in this study. A SAR with 45 \u00b0 depression angle and 9.6 GHz is most suitable for land applications (but requires significantly more power). Presently there is no decision which of these SAR modes should be selected. Polarisation will be selected later (HH or VV). The gray level resolution should be as good as possible and it is intended to attain 1 dB resolution. Therefore a special multilook SAR-processing is required. The average SAR operation time is presently limited to 6 min per orbit and to a maximum of 3 orbits per day.", + "Another payload element is a Data Collection System (DCS), (CNES, 1977), equivalent to the French ARGOS-system. This system applies the random access technique and receives the in situ measurement data from ground based Data Collection Platforms (DCPs) in the UHF-band. Certain reference data can be used for MOMS and SAR data processing. The onboard equipment of the DCS includes the UHF antenna and receiver as well as control and processing units.\nSatellite concept The major requirements determining the ERDSAT concept are: - -ARIANE launcher compatibility (e.g. 3 m fairing) --Mission requirements as given before --Three years lifetime (preoperational) spacecraft platform", + "---Growth potential to a multimission ARIANE platform - - \" L o w cost\" approach.\nThe selected configuration (Figs. 3 and 4) features: --modularity by separate modules for payload and bus compartments --thermostable structure for the payload platform --ease of integration ---extensive use of space proven hardware elements ---design experience from German ZKS programme.\nTwo versions were analysed with differences in SAR antenna mounting and, consequently, arrangement of optical sensors as well as solar generator geometry. Both allow slewing of the SAR antenna minimum depression angle form 45 \u00b0 to 71 \u00b0. Figure 5 shows Version II.\nThe bus allocated mass is 880 kg including payload telemetry, attitude and orbit control system and power conditioning, whereas the payload mass budgeted is 450kg. Therefore a margin of more than 1200kg remains, with respect to the ARIANE payload capability of 2600 kg (see Fig. 6).\nThe maximum power (see Fig. 7) needed is 1750 W (peak), which will be provided by a 1330W (BOL) solar generator and batteries. Half of the power required is necessary for battery loading as a 6 rain, 1000 W SAR operation was assumed even during eclipse." + ] + }, + { + "image_filename": "designv6_24_0003488_978-3-319-06590-8_14-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003488_978-3-319-06590-8_14-Figure5-1.png", + "caption": "Fig. 5 Hysteresis measurement test rig: a general, b detailed view, c samples and samples\u2019 supports expanded view", + "texts": [ + " The normal displacement-normal force curve has been obtained from interpolations of experimental data [18], later confirmed by Br\u00e4ndlein\u2019s theoretical investigations [19]. The hysteresis curve for a cylinder pressed against a plane has been measured using the already existing high-precision, high-temperature resistance flat-on-flat fretting test apparatus designed and set up by the AERMEC laboratory [20] originally dedicated to contact parameters measurement during wear process. The rig was modified to accomodate a cylindrical specimen (part C in Fig. 5c) and to avoid any relative rotation between flat and cylindrical specimens. The cylindrical specimen (C) is connected to a fixed support (A) by means of a \u201cseat\u201d (B). This part has a double function, it offers a flat surface on which to point the laser measuring system and it allows to test on the same support A different cylinder geometries (different radii) by redesigning the seat. The flat specimen (D) is fixed to a mobile support, excited by a shaker. The tangential force is measured by means of a load cell connected to support A, while the relative tangential displacement is measured by means of a laser doppler vibrometer" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001219_bf00862617-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001219_bf00862617-Figure2-1.png", + "caption": "Fig. 2.", + "texts": [ + ", the least error is introduced into the measurement results by a device, the moving part of which, while executing the movement relative to the measuring axis along with the body, is centered relative to this axis. Consequently, by using the scheme E, which facilitates centering of the body as well as the moving part of the device relative to the measuring axis, the best results can be obtained from the point of view of the reduction in the measuring error resulting from the inaccuracy in determining the coordinates of the center of mass. Here, the device should remain centered during the centering of the solid body. The device in [I] in which the scheme E has been used is shown schematically in Fig. 2. The frame 4, the load-carrying pallet 6, and the pallet counter-weight 3 are secured to the two-wire suspension 1 (any other vibrating system or a system with an acelerating rotating movement can also be used, for instance, Attwood's machine). The pallet counter-weight and the pallet are moved by means of the drive 2 in opposite directions along the frame such that the movement does not cause the displacement of the device's center of mass lying on the measuring axis X-X. The design of the mechanical drive may differ. In Fig. 2, it includes two lead screws connected by a gear drive. By the movement of the pallet, the center of mass of the body is led along a single coordinate to the plane containing the axis X-X and the orthogonal direction of pallet movement. The rotating support 8 and the support counterweight 5 which can be moved by the mechanical drive 7 (which is functionally and structurally analogous to the drive 2) are mounted on the pallet. The support 8 and the counterweight 5 are moved by means of the drive 7 in opposite directions along the axis orthogonal to the axis of pallet movement and the X-X axis" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000948_09544070jauto916-Figure16-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000948_09544070jauto916-Figure16-1.png", + "caption": "Fig. 16 Stress distribution of the transverse beam before and after optimization: (a) before optimization; (b) after optimization", + "texts": [ + " When the objective was met or the stress change between two neighbouring steps was zero, the calculation stopped. During the optimization, the number of spot welds after optimization is equivalent to or less than the original number, and the distance between two spot welds should be larger than 20 mm in order to avoid stress concentration. The analysis was carried out using ALTAIR/Optistruct software (Altair Engineering, Inc., Michigan, USA). The optimization of the spot weld layouts on the transverse beam is shown in Fig. 15, and the stress results are shown in Fig. 16. After optimization, the stress distribution of the component was improved and the maximum stress was reduced by almost 20 per cent. The fatigue life of the transverse beam increased from 3.956104 to 4.756104 km, as shown in Table 5. In the present study, fatigue life analysis and improvement of the autobody in a sports utility Proc. IMechE Vol. 222 Part D: J. Automobile Engineering JAUTO916 F IMechE 2008 at East Carolina University on July 1, 2015pid.sagepub.comDownloaded from vehicle (SUV) were carried out" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000370_ijmee.27.1.1-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000370_ijmee.27.1.1-Figure3-1.png", + "caption": "Fig. 3. Contact force versus gear teeth relative displacement characteristics. (a) Forward contact 3\u20134. (b) No contact 2\u20133. (c) Reverse contact 1\u20132.", + "texts": [ + " W1, W2, W3, W4 are the loads in bearings B1, B2, B3, B4, while \u00b5 \u03b8 \u00b5 \u03b8 \u00b5 \u03b8 \u00b5 \u03b8( \u02d9 ), ( \u02d9 ), ( \u02d9 ), ( \u02d9 )1 2 3 4 are the coefficients of friction of bearings B1, B2, B3, B4, and R is the bearing radius. T2 and T3, are the dynamic torques between the gears in mesh, while the control force F is the load developed between the gear teeth and i is the gear ratio. Equations (1)\u2013(4) are applicable under the following conditions: \u2206\u03b8 is the gear teeth angular clearance, \u03bb = =1 3 2i r r and T T r r T3 2 3 2 2= = \u2206( ) . The tangential loads between the gear teeth are T2 = Fr2 and T3 = Fr3 and r2, r3 are the gear radii. The contact force versus relative displacement for the gear teeth is illustrated in Fig. 3. The teeth are in forward contact between 3 and 4, they lose contact between 2 and 3, i.e. the clearance is open, and they re-contact on the backside between 1 and 2. When the gear teeth are in contact with each other during rotation, a variation in stiffness exists. Therefore care must be exercised when deciding on the type of function that will represent the stiffness. Usually a periodic function is used that can be approximated by a Fourier expansion. Under steady-state conditions the gear teeth, and the masses connected to them, move together with the same velocity and the backlash is taken up in one direction" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003599_tmag.2017.2698722-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003599_tmag.2017.2698722-Figure7-1.png", + "caption": "Fig. 7. Flux density distribution of PMs at different thickness. (a) 1mm. (b) 2mm. (c) 3mm. (d) Scale.", + "texts": [ + " And the magnet thickness hp is increased from 1 to 5 mm and the values of the lead in the range of 10 to 40 mm. It can be seen that the thrust force density up to some point, and then reduces. By using the largest thrust force density per volume of approximately 6.33 MN/m3. Therefore the magnetic screw lead \u03bb is designed as 20 mm. There is a point that the demagnetization of the PM must be taken into account when using HTS coils due to the heavy flux density. It is observed that thinner PMs are more prone to demagnetization, as shown in Fig. 7. Hence, based on the considerations of size, weight and cost, the PM thickness is selected to be 3 mm. The thrust force variation with the translator slot tooth length lt is shown in Fig. 8. Obviously, as the length of slot tooth increases, the thrust force of the EMSHTS gradually decreases. In order to produce the ideal helical magnetic field distribution, the length of slot tooth lt is designed to be 6mm. In order to demonstrate the proposed EMS-HTS, the electromagnetic screw and PM magnetic screw will be quantitatively analyzed and compared" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002224_3527608117-Figure1.60-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002224_3527608117-Figure1.60-1.png", + "caption": "Fig. 1.60 Cast brake disk particle of reinforced aluminum for the ICE 2 [64].", + "texts": [ + " Comparable construction unit characteristics are attainable only with the application of powder metallurgical aluminum alloys or when using heavy iron pistons. The reason for the application of composite materials is, as already described, the improved high temperature properties. Potential applications are in the area of undercarriages, e.g. transverse control arms and particle-strengthened brake disks, which can be also applied in the area of railmounted vehicles, e.g. for undergrounds and railway (ICE), see Fig. 1.60. In the 48 1 Basics of Metal Matrix Composites Tab. 1.9 Applications of metal composites. I. Drive shaft for people and light load motor vehicles (Fig. 1.61) [65]: Material: AlMg1SiCu + 20 vol. % Al2O3P Processing: extrusion form cast feed material Development aims: \u2013 high dynamic stability, high Young\u2019s modulus (95 GPa) \u2013 low density (2.95 g cm\u20133) \u2013 high fatigue strength (120 MPa for n = 5\u00d7107, R = \u20131, RT) \u2013 sufficient toughness (21.5 MPa m1/2) \u2013 substitution of steels II. Vented passenger car brake disk (Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003178_aps.2015.7304572-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003178_aps.2015.7304572-Figure1-1.png", + "caption": "Fig. 1. (a) Geometry of the proposed hexa-band antenna for mobile phone applications and (b) Antenna dimensions.", + "texts": [ + " For mobile handsets, the SAR level should not exceed the maximum permissible exposure levels set by the FCC [5], and designing a small multiband antenna with low SAR poses considerable challenges. In this paper, we present the design of a novel hexa-band mobile phone antenna, with a metal backing for SAR reduction. The proposed antenna comprises a coupling strip and two radiating strips, that are connected to a chip-inductorloaded strip. Details of the antenna design and the simulation results for the antenna performance are presented below. II. ANTENNA CONFIGURATION AND RESULTS The configuration of the proposed hexa-band antenna is shown in Fig. 1. The antenna is printed on the no-ground portion of the system circuit board of the mobile phone and covers a small area of 15\u00d729.5mm2. A 0.9-mm thick FR4 substrate (relative permittivity \u03b5r = 4.4 and size 50\u00d7115 mm2) is used to simulate the circuit board, and a ground plane of size 50\u00d7100 mm2 is printed on its back side. In order to increase the bandwidth of the antenna at high frequencies, a coupling strip is employed (see Fig. 1). In addition, a 9 nH chip inductor is used to achieve good matching of the antenna impedance at low frequencies. The proposed antenna has been simulated and optimized by using the software HFSS (High Frequency Structure Simulator). The dimensions of the final antenna model are shown in Fig. 1(b). 372978-1-4799-7815-1/15/$31.00 \u00a92015 IEEE AP-S 2015 The simulated S11 of the antenna is presented in Fig. 2. It is observed that the lower and upper bands cover the LTE band 13 and the DCS 1800 (1710-1880 MHz); the PCS 1900 (1850- 1990 MHz); WCDMA (1920-2170 MHz); LTE Band 40 (2300-2400 MHz); and Band 41 (2496-2690 MHz). The antenna gain varies from -0.5 dBi to 3.4 dBi, while the antenna radiation efficiency ranges from about 65.3% to 89.7% in the absence of the head phantom. Fig. 3 shows the SAR results in the presence of the SAM head model" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001283_iemdc.2013.6556163-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001283_iemdc.2013.6556163-Figure2-1.png", + "caption": "Figure 2. Definition of the stator and rotor synch frames. Definition of the synchronous s", + "texts": [ + " Finite ns of the transient with pull into step curve of the ated as a function of the ll tested LSSyR solutions e an IM competitor having laminations and windings. uipped with a cage made filling the saliencies, is the transient with motion under investigation: LS1 is a is a state of the art SyR rotor LS3 is the proposed LSSyR 978-1-4673-4974-1/13/$31.00 \u00a92013 IEEE II. MODELING OF THE LSSYR MA The dynamic model of the LSSyR machine concurrent presence of rotor saliency, as in and a short circuited rotor cage, as in an Indu The dq reference frame, synchronous to the ro Fig. 2. The rotor speed, in electrical radians, i vector, also in Fig. 2, is imposed by the AC the angular frequency \u03c9 and the synchron where p is the number of pole pairs. In the voltage slips at \u03c9 \u2212 \u03c9r and the slip s is defined \u03c9 \u03c9\u2212\u03c9 = rs (1) The phase angle \u03b4 of the voltage vector i Fig. 2. The rotor cage is not housed in usual ro same and regularly displaced along the airga would be in an asynchronous motor. On the c conductors here are aluminum bars that fill the SyR rotor completely, as in the examples With reference to the circuital model repo the d and q axes, the electrical equations of th sr s sss j dt d iRv \u03bb\u03c9+ \u03bb += dt d i R R r r rq rd \u03bb +\u22c5\u23a5 \u23a6 \u23a4 \u23a2 \u23a3 \u23a1 = 0 0 0 dt \u03bbdiR m fefe +=0 ( The subscript s stands for \u201cstator\u201d varia subscript r stands for \u201crotor\u201d. The non isotro cage bars is reflected into the rotor paramete and then the rotor resistances are different for in (3)", + " The two impedances Zpd and Zpq in (13) are called operational impedances: they are complex numbers intended as phasor operators, and their values are a function of the slip frequency s\u03c9 or, in other words, of the actual rotor speed. Yet, the stator current phasors have to be calculated. The voltage to current relationship is obtained by manipulation of (10) and (13). ( ) ( ) \u23a5 \u23a5 \u23a6 \u23a4 \u23a2 \u23a2 \u23a3 \u23a1 \u22c5 \u23a5 \u23a5 \u23a6 \u23a4 \u23a2 \u23a2 \u23a3 \u23a1 \u03c9+\u03c9\u2212 \u03c9\u2212\u2212\u03c9+= \u23a5 \u23a5 \u23a6 \u23a4 \u23a2 \u23a2 \u23a3 \u23a1 sq sd pqspd pqpds sq sd I I ZjsRZ)s( Z)s(ZjsR V V 1 1 (14) The voltage dq phasors are: \u23aa\u23a9 \u23aa \u23a8 \u23a7 \u03b4+\u03c9\u22c5= \u03b4+\u03c9\u22c5\u2212= )tscos(V\u0302V )tssin(V\u0302V sq sd 0 0 (15) Being the voltage phase angle \u03b4 referenced to the q axis of the rotor, as reported in Fig. 2. When the slip speed is not zero, the d and q components of the voltage vector are then phasors in time quadrature, regardless of the term \u03b40, that is the load angle at synchronous speed: \u23aa\u23a9 \u23aa \u23a8 \u23a7 = = V\u0302V V\u0302jV sq sd (16) The inverse of (14), finally, expresses the stator current phasors: ( ) ( ) \u23a5 \u23a5 \u23a6 \u23a4 \u23a2 \u23a2 \u23a3 \u23a1 \u22c5 \u23a5 \u23a5 \u23a6 \u23a4 \u23a2 \u23a2 \u23a3 \u23a1 \u03c9+\u03c9\u2212\u2212 \u03c9\u2212\u03c9+\u22c5=\u23a5 \u23a6 \u23a4 \u23a2 \u23a3 \u23a1 V\u0302 V\u0302j ZjsRZ)s( Z)s(ZjsR DI I pdspd pqpqs csq sd 1 11 (17) where ( ) ( ) pqpdpqpdssc ZZsZZRjsRD 22 21 \u03c9\u2212++\u03c9+= (18) By substitution of (17) into (13), also the stator flux components are determined" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003345_amm.813-814.964-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003345_amm.813-814.964-Figure4-1.png", + "caption": "Figure 4. Mesh model of Brake Pedal.", + "texts": [ + " This pedal is hinged about base plate by a hinged pin our foot apply force on this pedal. As shown figure 2, the 3D cad model of brake pedal is drawn in creo2.0 software as per dimension specification of old brake pedal.As existing pedal has more mass,best cad models are developed with consideration of different aspect as shown in figure 3. The material used for the brake pedal is is cold rolled steel sheet(FePo3 En10130 series) series which has following chemical and mechanical properties as shown in Table 1. As shown in Figure 4, with the help of ANSYS software, the geometric model was divided into tetrahedral finite elements of higher order (solid187). Each element is defined by 10 nodes \u2013 at corners and midside of edges of the tetrahedron. Each node of this element has three degrees of freedom allowing translational movement in x, y and z directions. The middle elements ensure high accuracy of calculations in the case of a complex geometry model. Finite Element mesh is refined at stamping portion. In the brake pedal, following loads are assumed - Generally human weight is about 700N - 1100N (as per the Knorr Bremse company standard) which forms main consideration for design of brake pedal" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003937_978-1-4471-5104-3_5-Figure5.3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003937_978-1-4471-5104-3_5-Figure5.3-1.png", + "caption": "Fig. 5.3 Positioning a rectangular spillway", + "texts": [], + "surrounding_texts": [ + "In the project and installation of hydroelectric power plants, it is essential to know the water flow rate and the fall height to access the potential energy. It is still important to know the maximum flow or maximum water discharge for construction of the engine room in a safe place outside of the flooding area. Knowing the water flow rate in various periods of the year can also set the pluviometric ranges (amounts of rain) in the area. The output power of a water turbine is based on a Bernoulli\u2019s derived equation as follows: Pt \u00bc gtqgQHm where gt is the turbine efficiency q is the water density in kg=m3 g is the acceleration of the gravity in m=s2 Q is the water flow in m3=s Hm is the water head in m:" + ] + }, + { + "image_filename": "designv6_24_0002224_3527608117-Figure6.13-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002224_3527608117-Figure6.13-1.png", + "caption": "Fig. 6.13 Turning of SiC-particle-reinforced brake drums: process operations, requirements, and process parameters.", + "texts": [ + " The surface and the peripheral zone of the Al alloy is altogether slightly plastic deformed. This is made clear by optical microscope images of the peripheral zone. Near the processed surface only broken SiC particles can be seen. Starting with the tests to select the cutting parameter and cutting materials, which are carried out with the machining of the brake surface of a SiC-particle-reinforced brake drum, a turning process took place in the area of the center, the fixing holes and the brake surface as well as the processing of the outside. Figure 6.13 shows the clamped, processed brake drum with the processing steps and the selected cutting conditions. The brake drum is clamped in a conventional three-jaw chuck, which is adjusted to the outside geometry of the drum. With this clamping a roundness of 14 \u00b5m can be reached in the area of the brake surface after finishing. As tools CVD diamond inserts with a corner radius of 1.2 mm are used. The complete turning process takes place at a cutting speed of 500 m min\u20131 in dry machining. In the area of the center of the fixing holes and when machining the out- side a feed rate of 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000230_t-ed.1985.22174-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000230_t-ed.1985.22174-Figure7-1.png", + "caption": "Fig. 7. Schematic diagram of the lambda lines and load lines in case A .", + "texts": [ + "A10) vcm = VCE- ('41 1) Substituting (A9)-(All) into (A2) and (A3) and then from (A7), we have F = - (4d&d [e 13 - (zsl0/pFI) [,(VDD - V W V T - 11 11 (VaEZ - vCE)ivT - APPENDIX I1 According to the magnitude of the slopes of both load line and X line, three different cases must be considered. 1) Case A: The magnitude of the slope of the load line - HC 1 - Zslo [e'vDD - vcE)fzvr - 11 -I- Zl . (A:L3) is smaller than that of the h line, i.e., R2 > - (Zs,o/@R,) [e('Bn - vCE)/VT - WU et al.: NEW APPROACH TO MODEL CMOS LATCHUP 1653 [mi. This case is indicated in Fig. 7 where according to the value of VcEAz, there are three subcases. i) Subcase AI: VcEAI < Vp. As may be seen from the curves of Fig. 7, there is only one true intersection point of Class III. Using (7), the condition V,,,, < Vp can be rewritten as 11 < (VDD - vP)/R2 - IpO (B1) where Zpo is equal to Zp - Z1 and is independent of ii) Subcase A2: VcEA, > Vv Only one intersection 11. point of Class I exists. The condition is I1 > (VDD - vV)/R2 - ZVI) (B2) where Z, is equal to Zv - Zl and is independent of iii) Subcase A3: Vp < VcEAl < Vv. The artificial intersection point is just the true intersection point of Class 11. There are two other true intersection points of Class 1 and Class III" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001074_iccv.1995.466882-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001074_iccv.1995.466882-Figure4-1.png", + "caption": "Figure 4: A collection of 2D rigid bodies under bounded translational motion relative to a 1D camera. Each body can translate by AX and AY, as shown for body E .", + "texts": [ + "2 Occlusion Graphs The occlusion relations for a multi-body system with no ambiguities can be represented by a directed occlusion graph. The graph is a pair ( V , E ) , where the vertex set V contains all of the bodies. To construct the edge set, E , consider all pairs t,y E V . Since there are no occlusion ambiguities, one of 2 y, x + y, or y b 3: must be true. In the first case no edge is added, while the other two cases add the directed edges (2, y) and (y, 2) respectively. Consider the collection of 2D rigid bodies viewed by a 1D camera which is illustrated in Fig. 4. Figure 5 (a) shows v (b) The first step in analyzing the existence of binary occlusion relations for an arbitrary pair of bodies is to model the bounded motion between them. We fix A and let M ( B ) denote the union of all possible positions of B. Its convex hull, C H [ M ( B ) ] , can be partitioned from A by a separating plane if the occlusion is unambaguous. This is illustrated in Fig. 3 (a) for two 2D bodies viewed by a 1D camera. The partition creates two half-spaces. If the image plane projections of A and CH[M(B)] don\u2019t overlap, A E B" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002295_b978-0-08-011216-9.50010-9-Figure7.1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002295_b978-0-08-011216-9.50010-9-Figure7.1-1.png", + "caption": "FIG. 7.1", + "texts": [ + " Introduction In previous chapters the engineering theories of bending and torsion have been developed and applied separately to appro priate problems. This chapter is concerned with components which are subjected to the combined effects of bending and torsion, such as shafts, coil springs or curved members. These provide useful examples of complex stress systems. The bending and torsion relationships derived in earlier chapters will be used frequently in this chapter and are quoted below for ease of reference \u03c3\u03c7 Mz E /1 1 \\ 7.2 Combined Bending and Torsion of a Circular Shaft The first example is illustrated in Fig. 7.1 in which a solid circular propeller shaft is supported in bearings at A and B and transmits a torque T. The weight of the propeller W causes bending in the shaft, but any bending due to the weight of the shaft will be considered as negligible. The torque is constant 229 along the length of the shaft and hence from equation (7.2) the maximum shear stress at the outer surface is 16\u0393 \u00abmax \" \u03c0 \u03af / 3 where d is the diameter. The transverse shear stress due to bending is zero where the maximum bending and torsional shear stresses occur and therefore need not be considered" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000800_941748-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000800_941748-Figure4-1.png", + "caption": "FIGURE 4. A 3D Model of a Flywheel and Corresponding Finite Element Analysis.", + "texts": [ + " The students are introduced to the CAD software by modeling a simple extruded object like that shown in Figure 2. Parts such as a connecting rod are used to further the complexity of the models. In later exercises, the students create 3D models such as the turbine blade and 2 flywheel shown in Figures 3 and 4. In addition, the students learn to create conventional engineering drawings automatically from the solid models. They are later asked to redefine a model to take advantage of symmetry and to export the meshed geometry to the FEA software for analysis as shown in Figure 4. After an introduction to basic finite element concepts using a planar example, the students are introduced to concepts of structures and elements. Simple problems such as analysis of beams that the students have already been exposed to in the core engineering courses can be used to illustrate finite element concepts. The closed form solutions to these problems can be used as a comparison to the finite element results. Students are exposed to concepts of plane-stress, plane-strain, axisymmetric problems and associated elements" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001362_j.cma.2012.05.009-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001362_j.cma.2012.05.009-Figure2-1.png", + "caption": "Fig. 2. Initial triangulation of five tetrahedra.", + "texts": [ + " The necessary and sufficient condition that no point of S is contained in the circumsphere of any tetrahedron in the triangulation is that any two adjacent tetrahedra in the triangulation are Delaunay with respect to each other\u2019s vertices. For the construction of Delaunay triangulation in three and higher dimensions, point insertion algorithm is the most popular, and many interesting methods have been proposed [19\u201323]. For a set of 3D points, the initial triangulation is a cuboid consisting of five or six Delaunay tetrahedra large enough to contain all the given points as shown in Fig. 2. The Delaunay triangulation is achieved by inserting points one by one into the initial triangulation. Each cycle of point insertion can be divided into three steps. (i) For a newly inserted point, identify all the tetrahedra whose circumsphere contains the point in its interior. The cavity left behind upon removal of these tetrahedra forms a starshaped insertion polyhedron. (ii) Owing to the finite precision arithmetic, the triangulation facets on the boundary of the cavity have to be verified with the visibility check and corrected before they are connected with the inserted point to form tetrahedra" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000193_0954406220906246-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000193_0954406220906246-Figure1-1.png", + "caption": "Figure 1. Mathematical models of leaf spring system: (a) main model indicating K\u2019, (b) first derived model indicating Kt, and (c) second derived model indicating Ks.", + "texts": [ + "37 In the present curved beam model, left end is directly hinged to fixed support and the right end is also hinged to another fixed support through a rigid link. Moreover, interaction between the curved beam and rigid link is not assumed to maintain tangency as well. The constrained rotational motion of the shackle is modeled using a rotational spring at the hinge point. The incremental mathematical leaf spring model simulating its large deformation behavior under current load step i is shown through schematic diagram in Figure 1(a). In the figure, dashed lines represent configuration under previous load step Wi 1 and solid lines represent deformed configuration under current incremental load Wi \u00bcWi 1 \u00fe Wi. In line with the virtual theoretical rotational restraint of the leaf spring model (Figure 1(a)), two more virtual but theoretically viable models are derived from it through post processing. In the first derived model, a virtual longitudinal spring connects the beam ends, whereas in the second model, a virtual vertical spring is used to define the system stiffness. The stiffnesses associated with these virtual models are termed as K\u2019, Kt, and Ks. The derived leaf spring models are shown in Figure 1(b) and (c), and described later on in connection with kinetic analysis of the original model. The incremental leaf spring model is presented in four sub-sections. The model ensures satisfaction of kinematic and kinetic restraints of the total system with respect to deformed configuration at current load step i. Such kinematic and kinetic descriptions of the leaf spring system in deformed configuration are presented in the \u2018\u2018Kinematics of leaf spring system\u2019\u2019 and \u2018\u2018Kinetics of leaf spring system\u2019\u2019 subsections, respectively", + " Force balance conditions of the curved beam in vertical and horizontal directions give the following two equations Ri AV \u00fe Ri BV \u00bcWi \u00f010\u00de Ri BH Ri AH \u00bc 0 \u00f011\u00de Similarly, moment balance condition of the curved beam about point A gives \u00f0Lc \u00fe R cos \u2019i\u00deRi BV \u00f0R sin \u2019i\u00deRi BH \u00bcWi Xi W Mi B \u00f012\u00de Resultant of reaction forces Ri BH and Ri BV at point Bi along ABi direction is given by Ti BR \u00bc Ri BH cos \u00f0 i\u00de \u00fe Ri BV sin \u00f0 i\u00de \u00f013\u00de As mentioned earlier to address the sheer impossibility of existence of rotational restraint of the present model, two more models are derived from the main model through post processing. True stiffness of the longitudinal spring Ki t, as used in modeling the longitudinally constrained motion of the shackle for the first derived model (Figure 1(b)), is now determined from the corresponding incremental force\u2013displacement as given below Ki t \u00bc Ti BR Li \u00f014\u00de In the above equation, incremental force Ti BR and span increment Li are defined as Ti BR \u00bc Ti BR Ti 1 BR and Li \u00bc Li Li 1, respectively, whereas total span increment L0 \u00bc Li L0 and total force Ti BR give total stiffness of the longitudinal spring as given below K0 t \u00bc Ti BR L0 \u00f015\u00de Similarly, true and total stiffnesses of the vertical spring of the second derived model (Figure 1(c)) are determined as given by Ki s \u00bc Wi Yi W \u00f016\u00de K0 s \u00bc Wi Y0 W \u00f017\u00de where Y0 W is total deflection of loading point defined as Y0 W \u00bc Y0 W Yi W. Deformation analysis of curved beam Deformation behavior of curved beam under the current incremental load Wi \u00bcWi 1 \u00fe Wi, as observed with respect to body fitted frame, is shown in Figure 4. Let the incremental displacement fields in curvilinear frame be ws and us for bending and stretching deformations, respectively. Kinematic boundary conditions of the two displacement fields are ws \u00bc 0 at si 1 \u00bc 0 and ws \u00bc 0 at si 1 \u00bc Si 1 \u00f018a\u00de us \u00bc 0 at si 1 \u00bc 0 and us \u00bc unknown at si 1 \u00bc Si 1 \u00f018b\u00de Strain\u2013displacement relation in curvilinear frame for curved beam under combined bending and stretching is derived as \"iss \u00bc 1 \u00f01 i 1 ni 1\u00de u0s \u00fe 1 2 \u00f0u 0 s\u00de 2 \u00fe 1 2 \u00f0w 0 s\u00de 2 ni 1 w00s , following literature" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001472_9781118354179.auto089-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001472_9781118354179.auto089-Figure2-1.png", + "caption": "Figure 2. Most common clutch installation\u2014attached to the engine flywheel. (Reproduced from Reimpell, Stoll, and Betzler, 2001. \u00a9 Elsevier.)", + "texts": [ + " Before entering into design considerations, it is necessary to summarize main tasks of the clutch: \u2022 Establishment and interruption of power flow from the engine to other transmission assemblies; \u2022 Execute the above functions gradually, enabling full control of the engagement (disengagement) process; \u2022 Enable vehicle start and gear change; \u2022 Enable full engine operation when \u201cgears\u201d are fully engaged; \u2022 Dampen torsional vibrations during the engagement (disengagement) process and at all times when the clutch is engaged; \u2022 Limit the torque input, by slipping, to protect transmission system from overload in critical conditions (e.g., wheel locking when braking). Main transmission clutch is typically attached to the engine flywheel, as indicated in Figure 2 (Reimpell, Stoll, and Betzler, 2001), which represents by far the most common installation. However, other installations are possible, when the engine and gearbox are split, usually to achieve better weight distribution between vehicle axles. In such cases, the clutch can be still attached to the engine flywheel or designed integral to the gearbox, the latter shown in Figure 3 (Nunney, 1998). Having in mind previous consideration, the most commonly used type of the main transmission clutch will be considered, having the following characteristics: \u2022 Friction type: Torque transfer is based on dry friction principle" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003257_s11771-015-2507-9-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003257_s11771-015-2507-9-Figure1-1.png", + "caption": "Fig. 1 Structure of quadrotor helicopter", + "texts": [ + " According to above reasons, this study designs a controller based on PIDNN control scheme in order to solve the control problem of a quadrotor under influence of wind disturbance. Due to friction, swirl and other reasons, wind often appears accompanied with turbulence [24]. There is usually turbulent wind field acting on quadrotor because the flying altitude is low in real conditions. To make the simulation results accord with the practical situation, the turbulent wind field is generated according to Dryden model and taken into consideration as the disturbance source of the quadrotor helicopter. The quadrotor is composed of body and four rotors, as presented in Fig. 1. The thrust force is generated by four rotors. The motion of quadrotor is controlled by varying the rotation speed of four rotors to change the thrust and torque produced by each one. Four rotors are divided into two pairs: pair (1, 3) and pair (2, 4). The rotate direction of the two pairs is contrary in order to counteract the aerodynamic torque generated by rotors\u2019 rotation. Increase or decreasing the rotation speed of the four rotors simultaneously will generate vertical motion. Independently varying the speed of the rotor pair (1, 3) can control the pitch angle about y-axis and the translational motion along x-axis" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002153_amm.813-814.915-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002153_amm.813-814.915-Figure2-1.png", + "caption": "Figure 2: Loads, boundary conditions & meshing", + "texts": [ + " All DOF\u2019s are fixed at the hub of the wheel and pressure of 6MPa is applied on the face of the wheel rim for static structural analysis. For modal analysis all DOF\u2019s are fixed at the hub. Cylindrical co-ordinate system is assigned to the wheel, because wheel is in cylinder shape. entities. The model is meshed with the 10 node tetrahedral 3-D solid element termed as solid 187 in Ansys software. The size of the element gradually decreased to get the meshing convergence. Static Structural Analysis. A pressure of 6 MPa is applied on the face of rim and all DOF are fixed for the hub inner face as shown in Fig 2b. The response of the wheels in terms of Total deformation, von mises stress, structural stiffness and specific structural stiffness are calculated and compared the results for all 6 design with three materials discussed in previous section. Tabular Values Spoke type/Result type Inclined spokes Curved spokes Y-shape Spokes Deformation [mm] 0.4257 0.2956 0.2327 Von mises stress [Mpa] 574.76 384.85 791.4 Stiffness [N/M] 813.26 1171.3 1487.9 Mass[Kg] 9.3295 9.8311 10.831 Specific structural stiffness[N/M-kg] 87", + "38 stiffness of 3 spoke curved design made up of steel is better than others. \u2022 All 5 spokes designs produced the higher von mises stress than the 3 spokes designs. It can be observed that the von mises stress does not vary much based on the type of the material chosen for manufacturing. All three spoke designs lower von mises stress than the other designs. Modal Analysis Results. It is done to find out the natural frequencies of the structure. All degrees of freedom are fixed at the hub as shown in fig 2a and performed the modal analysis for all designs. The typical contours are shown in below section. Contours Tabular Values Type/Mode Number Inclined spokes Curved spokes Y- shape 3 5 3 5 3 5 1 126.7 178.9 134.6 196.5 118.2 166.1 2 132.7 179.2 141.9 195.6 125.2 166.4 3 168.5 245.9 172.2 257.1 173.0 269.9 4 222.9 248.5 230.7 262.8 207.5 283.9 5 225.0 249.3 234.7 267.9 270.1 327.0 6 233.1 342.2 276.9 380.5 270.4 328.2 Mass (Kg) 1.914 2.139 1.985 2.258 2.123 2.483 Observations. All 5-spoke designs have higher natural frequency than the 3-spoke designs" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003634_powereng.2015.7266300-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003634_powereng.2015.7266300-Figure2-1.png", + "caption": "Fig. 2. Designed 6 kW BLDC outer-runner motor.", + "texts": [ + " kC shows the Carter factor and formulated with the following equation [5]; ( )2 1 1 5 s C open ss s ss open T k k b T b k \u03b4 = \u2212 + (9) where Ts ,which is the stator tooth opening and the common parameter in both outer and inner-runner BLDC motors can be indicated as; s s DT Q \u03c0= (10) By using the above equations, the electromechanical parameters of the 6 kW outer-runner, on-drum BLDC motor have been determined. The analytical results are presented in Table 4 and also the 3D representation of the designed motor is shown in Fig.2. III. MAGNETIC FIELD ANALYSIS BASED ON FINITE ELEMENT ANALYSIS (FEA) AND CO-SIMULATION Finite element method (FEM) is applied for optimizing the construction during the electrical machine design procedure. FEM is composed of many methods and sub-modules and which can help the designer to shorten the development period before the real model is built. Since transient analysis which is the basic analysis component of finite element method is dependent on time and computes the circuit as a function of time for a given or predefined time range, it has been used for the check up on analysis algorithms, control options with different initialization parameters while designing 6 kW, 60 V BLDC motor" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001865_icicv50876.2021.9388494-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001865_icicv50876.2021.9388494-Figure3-1.png", + "caption": "Fig 3. Micropatch element of H shaped design", + "texts": [], + "surrounding_texts": [ + "1. Introduction 2. Proposed Method:\nThe H-shaped antenna is shown in figure.1.\nFig.1 H-shaped antenna\nThe figure shows that H-shaped patch antenna is available for the transfer of energy waves.\nFig.2 H-shaped antenna \u2013 front\nview\nThe H shaped antenna was studied for Epsilon value of 2.33 and loss tangent of 0.0012 values. The loss microstrip antenna can be depicted as follows. A very advanced antenna element is developed in which the function across all predominating conditions and environment functions.\n978-1-6654-1960-4/21/$31.00 \u00a92021 IEEE 58\n20 21\nT hi\nrd I\nnt er\nna tio\nna l C\non fe\nre nc\ne on\nI nt\nel lig\nen t C\nom m\nun ic\nat io\nn T\nec hn\nol og\nie s\nan d\nV ir\ntu al\nM ob\nile N\net w\nor ks\n( IC\nIC V\n) | 9\n78 -1\n-6 65\n4- 19\n60 -4\n/2 0/\nAuthorized licensed use limited to: Carleton University. Downloaded on May 25,2021 at 13:04:54 UTC from IEEE Xplore. Restrictions apply.", + "The glass reinforced epoxy resin laminate patch antenna design is shown in Fig.5.\nIn our analysis of antenna design, this is very important information. This material is a reinforced glass\nportion that greatly improves the work of the antenna. This diagram and architecture of radiation pattern will further proceed in our synthesis of research.\nThe impedance or resistivity function is given in the following figure.\nThe yagi-uda antenna was associated with the following strips.\n978-1-6654-1960-4/21/$31.00 \u00a92021 IEEE 59\nAuthorized licensed use limited to: Carleton University. Downloaded on May 25,2021 at 13:04:54 UTC from IEEE Xplore. Restrictions apply.", + "For 15 strips of element the figure for YagiUda is given as:\nThe further radiation analysis is given as:\nYagi-uda was the basic antenna element that was later used to study the various other antenna derivatives of the antenna.\nThe radiation pattern due to our dipole description of H shaped element is given in the following figure.\nThe folded H shaped radiation pattern looks as follows.\n978-1-6654-1960-4/21/$31.00 \u00a92021 IEEE 60\nAuthorized licensed use limited to: Carleton University. Downloaded on May 25,2021 at 13:04:54 UTC from IEEE Xplore. Restrictions apply." + ] + }, + { + "image_filename": "designv6_24_0003536_978-94-009-5063-4_1-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003536_978-94-009-5063-4_1-Figure6-1.png", + "caption": "Figure 6. Load-end shortening curve with limit point A, bifurcation point B, and post-bifurcation equilibrium path BD (photographs courtesy Sobel and Newman [42]).", + "texts": [ + " D \" ~ Limit Load of Imperfect : Shell (Asymptotic 4 Analysis) I I I OL-----------------------~ BUCKLING MODAL DISPLACEMENT, wb (b) Figure 7. Load-deflection curves showing limit and bifurcation points. (a) general nonlinear analysis, (b) asymptotic analysis. perfect structures, the two phenomena loosely termed \"buckling\" are collapse at the maximum point in a load vs. deflection curve and bifurcation buckling. These two types of instability failure are illustrated in Figs. 6 and 7. The axially compressed cylinder shown in Fig. 6 deforms approximately axisymmetric ally along the equilibrium path OA until a maximum or limit load AL is reached at point A. If the axial load A is not sufficiently relieved by the reduction in axial stiffness, the perfect cylinder will fail at this limit load, following either the path ABC along which it con tinues to deform axisymmetrically, or some other path ABD along which it first deforms axisymmetrically from A to B and then nonaxisymmetrically from B to D. Limit point buckling, or \"snap-through\" occurs at point A and bifurcation buckling at point B. The equilibrium path OABC, corresponding to the axisymmetrical mode of deformation, is called the fundamental or primary or pre buckling path; the post bifurcation equilibrium path BD, corresponding to the nonaxisymmetrical mode of deformation is called the secondary or post-buckling path. Buckling of either collapse or bifurcation type may occur at loads for which some or all the structural material has been stressed beyond its proportional limit. The example in Fig. 6 is somewhat unusual in that the bifurcation point B is shown to occur after the collapse point has been reached. In this particular case, therefore, bifurcation buckling is of less engineering significance than axisymmetric collapse. A commonly occurring situation is illustrated in Fig. 7(a). The bifurcation point B is between 0 and A. If the fundamental path OAC corresponds to axisymmetrical deformation and BD to nonaxisymmetrical deformation, then initial failure of the structure would generally be characterized by rapidly growing nonaxisymmetrical deformations", + " This occurs for the more slender columns, p = 2 X 10-4 , at p* = 0.155. The load cannot be increased beyond this value: The structure fails by bifurcation buckling with the columns temporarily bending during the process. For the structure with p = 4 X 10-4 , the point of intersection (bifurcation) occurs beyond the maximum in the primary load displacement curve, indicating that the columns are straight at the inception of snap-through. The behavior represented by the curve OA'B'D in Fig. 11 is analogous to that represented by the curve OABD in Fig. 6; the behavior represented by the curve OBD in Fig. 11 is analogous to that represented by the curve OBD in Fig. 7(a). In bifurcation buckling analysis it is often assumed that nonlinearities and geometrical changes in the pre buckling range can be omitted. As the columns buckle at N* = 1, the critical load of the structure in such a model is Post-bifurcation stability: Consider now a structure that has been slightly modified as shown in Fig. 13 by addition of a linear spring which carries a part of the load", + " However, if the spring constant is sufficiently large the slope of the line for the secondary solution becomes positive. The increase of the load in the spring is more than sufficient to com pensate for the decrease in the load carried by the columns. The two-column structure discussed here illustrates the behavior of structures of a more general nature. For example, the curve in Fig. 11 labeled OA'B'D is typical of failure of axially compressed cylindrical shells which buckle plastically and develop nonsymmetric folds after the load has reached its maximum value, as shown in Fig. 6. The curve in Fig. 11 labeled OBD is typical of shallow spherical caps under uniform external pressure in which nonlinear prebuckling effects are important but failure is by nonsymmetric bifurcation buckling. A rather thick cylindrical shell under axial compression deforms axisymmetrically throughout the collapse process. This would be indicated in Fig. 11 by a primary equilibrium path similar in shape to the curve OA 'B'C but lying under it and not intersecting the column bifurcation line at all" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002590_0022-2569(71)90034-6-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002590_0022-2569(71)90034-6-Figure1-1.png", + "caption": "Figure 1. To rus t r a ced by point Aa.", + "texts": [ + " *Assistant Lecturer in Mechanical Engineering, Cauifield Institute of Technology, Victoria 3145, formerly Research Scholar, Department of Mechanical Engineering, Monash University, Clayton, Victoria 3168, Australia. SDean, Faculty of Engineering, Monash University, Clayton, Victoria 3168, Australia. 241 Arrangement of the Linkage Suppose a link 3 to be connected to a base 1 by two revolute pairs 12 and 23 in series, these two pairs being mutually perpendicular and at a distance q apart: q i_s the length of the crank 2. A point on link 3 can lie anywhere on a surface when both revolutes 12 and 23 are perfectly free. The point Aa, at a distance b from the axis of 23 (as shown in Fig. 1) must be somewhere on the surface of a circular torus whose pr imary and secondary radii are respect ively q and b. Nex t suppose a link 4 to be connected to the same base 1 through acyl indr ica l pair 14. A point A4 at a distance r from the axis of 14 must lie somewhere on the surface of a circular cylinder, of radius r, axis 14 (see Fig. 2). Axis 14 is shown intersecting 0z orthogonally in P and lying parallel with 0)': 0P = p. In general this cylinder intersects the torus in a spatial closed curve that has two separate branches" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003315_detc2005-84562-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003315_detc2005-84562-Figure1-1.png", + "caption": "Figure 1. Contact tooth load by PISE method", + "texts": [ + " Ksi = Fi \u03b4i (1) Contact force in meshing gears at contact region Gear mesh stiffness at a point of contact between a gear pair greatly depends upon the displacement at local contact point in each body and varies along the line of action. In this paper, the novel Pseudo-Interference Stiffness Estimation(PISE) method is introduced to estimate the contact force at the contact region of the gear teeth. The displacement at local contact area is assumed to be equal to the overlap region of gear teeth, as shown figure 1. With the obtained displacement, the gear mesh contact force, Fcont , can be calculated as follows: Keq cont = n \u2211 i=1 Keq i (2) where Keq i = K1 ciK 2 ci K1 ci +K2 ci (3) 2 wnloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 02/01/2016 Te Kci = Ksimax dh h = Ksimax n (4) Fcont = n \u2211 i=1 Keq i \u03b41.05 i (5) When Ksimax , h, \u03b4i and n are singular stiffness at the point of maximum constructive interference, the length of contact area, the constructive interference at point i in contact area, and number of discretizations" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001244_tasc.2016.2543267-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001244_tasc.2016.2543267-Figure5-1.png", + "caption": "Fig. 5. FE model for calculating forces at the end of the rotor", + "texts": [ + " When skew is applied, Torque ripple is dramatically reduced compared to the no-skew model. The Vskew model has reduces torque ripple as much as the conventional skew model. As mentioned above, generated force is calculated as (5)-(7). Axial force is mostly generated at the ends of the rotor due to the magnetic flux density in the z-direction, which is caused by magnetic saturation. FEA models to calculate forces at the ends of the rotor are designed with air regions covering the ends of the rotor to apply Maxwell stress tensor as in Fig. 5. In Fig. 6, unexpected axial forces at the ends of the rotor in comparative analysis models are shown. Axial force is produced in every end in every model. In the non-skew model, axial forces at the ends of the rotor are the same as in the opposite direction. Therefore net axial force is zero. In contrast, the conventional skew model has different amounts of axial force at the ends of the rotor due to the unbalanced magnetic saturation, which results in remnants of 1051-8223 (c) 2015 IEEE. Personal use is permitted, but republication/redistribution requires IEEE permission", + " Skew Angle Skew Angle Skew Angle 1051-8223 (c) 2015 IEEE. Personal use is permitted, but republication/redistribution requires IEEE permission. See http://www.ieee.org/publications_standards/publications/rights/index.html for more information. Characteristic differences exist between Cases 1 and 2 due to the magnetic flux density distribution. Product of \ud835\udc35\ud835\udc5f and \ud835\udc35\ud835\udf03 decides the magnitude of the torque with Maxwell Stress Tensor Method according to (2)-(4), since the size of element at air-gap is uniform as shown in Fig. 5. Therefore sum of \ud835\udc35\ud835\udc5f \u2219 \ud835\udc35\ud835\udf03 in unit stack length along with the z-coordinate are shown in Fig. 12. Magnetic saturation at the ends of the rotor gives rise to the bypass linkage flux. Thus, \ud835\udc35\ud835\udc5f \u2219 \ud835\udc35\ud835\udf03 occurs over the boundary of stack length. Since case 1 is more magnetically saturated in the ends as described in Fig. 11, bypass linkage flux is produced more in case 1 compared to that of case 2. Magnetic saturation of rotor end is almost uniform so that bypass linkage flux has almost constant value in simulation time period" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000934_j.still.2017.12.011-Figure11-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000934_j.still.2017.12.011-Figure11-1.png", + "caption": "Fig. 11. Plastic strain distribution with different cutting depths, (a) at 50mm, (b) at 100mm and (c) at 150mm.", + "texts": [ + " Soil failure can be defined as the permanent deformation of the soil (Stafford, 1984; Rajaram and Gee-Clough, 1988). For experiment, the failure angle was usually measured by connecting the points where the deformation of soil is most remarkable. Therefore, for simulation, the plastic deformation of soil was selected as a measure of soil failure and the failure angle was obtained from plastic strain distribution. As shown in Fig. 10 (a), (b) and (c), the plastic strain increased evidently with the increase of cutting angle. However, as shown in Fig. 11 (a), (b) and (c), the plastic strain kept almost unchanged with the increase of cutting depth. The failure angle \u03b3 is defined as the angle between the direction of the shear deformation zone and the direction of proceeding of the cutting blade. The failure angle decreased as the cutting angle increased, whereas it kept almost unchanged with different cutting depths. These results are in consistent with the experimental ones of Hatamura and Chijiiwa (1976a). The soil chip thickness l reflects the accumulation of soil caused by cutting. Fig. 10 and Fig. 11 show that both the increase of the cutting angle and cutting depth would lead to the increase of the chip thickness. The ratio of the chip thickness to the cutting depth ranges from 1.31 to 1.71 (Table 3). This ratio increased with the increase of cutting angle. However, it remained almost unchanged with the increase of cutting depth. Elijah and Weber (1971) noted that for flow failure with inclined flat blades, the thickness of the soil chip normal to the blade surface was greater than the depth of cut because of soil deformation" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000854_6.1995-1779-Figure12-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000854_6.1995-1779-Figure12-1.png", + "caption": "Figure 12a. Comparison of Test and CFD Predicted Pressures", + "texts": [ + " There are no significant differences between computational and experimental data which can be seen on this scale. 15 2 American Institute of Aeronautics and Astronautics Figure I l a . Comparison of Test and CFD Predicted Pressures F/A- 18E Sting and Distortion Model - Real Afterbody, Top View Mach= 0.85 a =3.5\" Figure 11 b. Comparison of Test and CFD Predicted Pressures F/A- 18E Sting and Distortion Model - Real Afterbody, Bottom View Mach=0.85 a =3.5\" F/A- 18E Sting and Distortion Model - Distorted Afterbody, Top View Mach=0.85 a =3.P Figure 12b. Comparison of Test and CFD Predicted Pressures F/A- 18E Sting and Distortion Model - Distorted Afterbody, bottom View Mach = 0.85 a = 3.5\" 15 3 American Institute of Aeronautics and Astronautics A more sensitive measure of the accuracy of the predicted surface pressures is a comparison of the integrated lift and drag from these pressures. Note that the pressure instrumentation was designed to obtain the difference in drag between the real and distorted geometries. Accurate measurement of the full configuration absolute force levels was not a test objective and would have required significant increases in tap density" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001784_92.250205-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001784_92.250205-Figure4-1.png", + "caption": "Fig 4 Plot\\ ( 1 1 the n u n u d l l d ) o u t \\ produccd l o r the X N O R gdc", + "texts": [], + "surrounding_texts": [ + "[ l ] F. Augustine and R. Varadarajan, \u201cEfficient mappings for multi-di- mensional systolic arrays using flexible buffer structures,\u201d UF-CIS Tech. Rep. TR-92-009, 1992. [2] P. R. Cappello and K. Steiglitz, \u201cUnifying VLSI array design with geometric transformations,\u201d in Proc. Inr. Conf: Parallel Processing, Bellaire, MI, Aug. 1983, pp. 448-457. [3] H. T. Kung and C. E. Leiserson, \u201cSystolic arrays (for VLSI),\u201d Sparse Matrix Proc., SIAM, Duff er a l . , eds., Philadelphia, PA, 1978, pp. 245-282. [4] P. Lee and Z. M. Kedem, \u201cSynthesizing linear array algorithms from nested For loop algorithms,\u201d IEEE Trans. Comput.. vol. 37. pp. 1578-1598, Dec. 1988. [5] -, \u201cMapping nested loop algorithms into multidimensional systolic arrays,\u201d IEEE Trans. Parallel Distrib. Syst., vol. 1, pp. 64-76, Jan. 1990. [6] G. J. Li and B. W. Wah, \u201cThe design of optimal systolic arrays,\u2019\u2019 IEEE Trans. Comput., vol. C-34, pp. 66-67, Jan. 1985. 171 W. L. Miranker and A . Winkler, \u201cSpace-time representations of computational structures,\u201d Computing, vol. 32, pp. 93-1 14, 1984. [8] D. I. Moldovan and J. A. B. Fortes, \u201cPartitioning and mapping algorithms into fixed size systolic arrays,\u201d IEEE Trans. Compur., vol. 35, pp. 1-12, Jan. 1986. [9] P. Quinton, \u201cAutomatic synthesis of systolic arrays from uniform recurrent equations,\u201d in Proc. IEEE 11th Int. Symp. Comput. Architect., Ann Arbor, MI, 1984, pp. 208-214. 1101 S. V. Rajopadhye and R. M. Fujimoto, \u201cSystolic array design by static analysis of program dependencies,\u201d Parallel Architectures and Languages Europe, J . de Bakker, A. J. Nyman, and P. C. Treleaven, Eds. New York: Springer Verlag, 1987, pp. 295-310. [ 111 S. K. Rao and T. Kailath, \u201cArchitecture design for regular iterative algorithms,\u201d in Systolic Signal Processing Systems, E. E. Swartzlander, Ed. New York: Marcel Dekker, 1987, pp. 209-297. [I21 W. Shang and J . A. B. Fortes, \u201cTime-Optimal and Conflict-Free Mappings of Uniform Dependence Algorithms into Lower Dimen- tal results based on a variety of logic cells which demonstrate the benefit of the LI in terms of cell area and routing flexibility. Simulation results indicate that this benefit is without any detrimental effect on electrical performance. I. INTRODUCTION The local interconnect (LI) layer in full custom CMOS technology is created after the polysilicon and the S/D diffusions are in place, but before the isolation oxide is deposited. This makes it possible to directly connect LI to polysilicon gates and/or the S/D diffusions without etching contact cuts in an isolation oxide. The isolation oxide is deposited after the formation of LI connections. Hence, normal metallization procedures can be used to make connections to source/drain diffusions, polysilicon as well as LI. LI was first introduced by Texas Instruments in their 0.8 pm biCMOS technology as a very thin layer of TiN that forms as a byproduct of the self-aligned TiSiz fabrication process [I] . It is possible to pattern the LI layer to make connections using a single additional mask. Among other existing approaches, Phillips Research Laboratories introduced a TiSi2 LI layer in their 0.5 pm CMOS process [ 2 ] . This TiSi2 LI layer is specifically crated and it is not a by-product of fabrication. Both TiN and TiSi2 are highly conductive layers and have a sheet resistance comparable to that of polysilicon (e.g., 14 Q/sq for TIN versus 20 Q/sq for typical polysilicon), so both types of LI are suitable for making signal connections. LI has been used for just this purpose in SRAM cells [ I ] , [2] to reduce overall cell size and increase speed. In this paper we propose to use LI for selected connections in other digital CMOS circuits to improve cell area and routing flexibility. We present a methodology for using LI in custom logic circuit designs. Because the methodology is simple, it can easily be encapsulated in an automatic layout synthesis package for cell generation. We provide experimental results that compare circuits routed with and without LI connections for both manual layouts Manuscript received March 15, 1993; revised July 20, 1993. This work was supported in part by the Natural Sciences and Engineering Research Council of Canada, and Micronet. The authors are with the Department of Electronics, Carleton University, Ottawa, Ont., Canada KlS 5B6. IEEE Log Number 9213101. 1063-8210/93$03.00 0 1993 IEEE IEEE T R M C L = C M \" ON VERY LARGE SCALE INTEGRATION (VLSI) SYSTEMS, VOL. I , NO. 4, DECEMBER 1993 567 and synthesized layouts. We also present some simulation results to show that the use of LI connections has a negligible impact on the performance of logic cells. 11. LAYOUT METHODOLOGY Our methodology is based on the line-of-diffusion style of layout [3], which is very common in both manual design of standard cells and automatic cell synthesis. Potential uses for the LI layer are guided by two objectives. The first and most important objective is a reduction in layout area. The secondary objective is a methodology aimed at providing maximum routing flexibility by a) providing an I / O port on polysilicon at the top and bottom of the cell for every extemal signal, and b) providing Z/O ports on metal 1 at the sides of the cell for as many of the external signals as possible, in addition to the providing an Z/O port on metal 2 for every extemal signal. Based on these objectives, we identified three possible uses for the LI layer. The first use is for connecting the n-device and p-device diffusions together (see Fig. I ) . Up to this point, such connections have been made in metal 1 with a minimum of two contacts required'. Use of LI here eliminates the need for contacts. Also, in many routing programs, the preferred direction for metal 1 connections is horizontal since the polysilicon gates usually run vertically between the p- and n-diffusion lines. Since LI and metal 1 make no direct connection (a contact would be required), the conflicts between horizontal metal connections and p- to n-diffusion connections area avoided. A second potential usage of the LI layer is for directly connecting sourceldrain diffusion regions to polysilicon wires as shown in Fig. 2. As with the n-device/p-device connections, such connections are presently made in metal 1, and require a minimum of two contacts (or three if both n-device and p-device diffusions are involved). LI eliminates the need for any contacts. The third use for LI is aimed at increasing global routing flexibility by facilitating the creation of polysilicon Z/O ports at the top and bottom of the cell (see Fig. 3). Leading edge routing programs are capable of exploiting a large number of routing layers for making connections to the cells. At the cell design level, this translates into a need for having as many of the Z/O ports as possible available on each of the routing layers. Currently, to make signals implemented in diffusion available on polysilicon, (usually at the top and bottom of the cell) it is necessary to run a small metal strap off the diffusion to which a polysilicon strip can be connected. This again requires a minimum of two contacts. It also usually implies that there is a break in the diffusion line for the polysilicon strip because the power and ground lines normally run horizontally at the top and the bottom of the cell, making it impossible to run the metal strap off the diffusion at the top or bottom. LI eliminates the need for the contacts, and for the break in the diffusion line. 111. IMPACT OF LI ON LAYOUT Q U A L I T Y This section summarizes the results of two layout experiments aimed at quantifying the benefit of using the LI layer in terms of layout quality. A . Manual Layout We performed manual routing following automatic transistor placement for five circuits, both with and without LI connections. The transistor placement algorithm uses a composite metric which applies connectivity and optimal chaining considerations simulta- a) with LI b) with Metall Fig. 3 . Polysilicon I / O port from diffusion. neously, [4]. We used standard CMOS design rules with 0.8 pm minimum feature size. We chose the n-device widths (4.4 pm) to be 60 percent of the p-device widths (7.2 pm). This is in line with common practice in standard cell design, and allows a total of five over-the-transistor metal 1 tracks for intracell routing. For LI, we used the design rules stated in [2] as a guideline. Whenever no specific rule for LI was indicated (e.g., for overlaps and enclosures), we assumed that the comparable rule for polysilicon applied to LI as well. We further assume that diffusion regions are silicided such that source and drain resistances are not critical during the layout process. This allows the use of single contacts from metal 1 to diffusion located anywhere on the source or drain region. This layout style promotes routing over the transistors and is most efficient in terms of area. The five circuits used for our comparisons were specifically chosen because they are highly interconnected and known to be difficult to route. They are as follows. 1) a 2-input XNOR gate [ 5 ] 568 IEEE TRANSACTIONS ON VERY LARGE SCALE INTEGRATION (VLSI) SYSTEMS, VOL. I . NO. 4. DECEMBER 1993 2) d non-dual tul l addc'r conrdining tour non-dual gale\\ and two inverters [6] 3) a master-slave D-type flip flop with asynchronous set and reset 4) a 3-to-8 decoder that contains eight 3-input NAND gates and three inverters, and 5) a full adder comprised of eleven logic gates. Plots of the layouts of these circuits which were routed with and without LI can be found in Figs. 4 to 8. For each circuit layout (ten in total), the following measurements were taken: I ) cell area (including cell length and cell height), 2) number of metal 1 contacts (contacts between metal 1 and either diffusion, poly, or LI), 3) number of LI wires and contact points to diffusion and polysilicon, and 4) number of metal 1 I / O ports available as an indicator of rout- a) with Local Interconnect ing flexibility (see Fig.-4). b) without Local Interconnect Table I contains a summary of comparisons between the two Fig. 5 . Plots of the manual layouts produced for the non-dual full adder cell. sions (with and without LI connections) for each of the five circuits. The results in the table are expressed in terms of a comparison of the improvement in the layout with LI versus the layout without LI. B. Automatic Layout As a second layout experiment, the simple rules for the use of local interconnect in leaf cell design were included in the Picasso leaf cell synthesis system [4]. Following transistor placement as two parallel rows of transistors, all the nets corresponding to the patterns identified in Section I1 are prerouted using the local interconnect layer. Then, a modified maze router is invoked to realize the remaining nets in metal 1 and polysilicon. Finally, layout compaction is accomplished with a two-dimensional compactor which supports automatic jog insertion and octagonal geometry. A total of 13 circuits were used for comparison in the automatic layout context, including the five circuits used for manual layout. Results are summarized in Table 2. Sample layouts produced for two of the benchmark circuits are shown in Figure 9. C. Evaluation of the Results A number of conclusions can be drawn from the results of the two layout experiments. Manual layouts with LI use on average 1 I percent less area (and as much as 20 percent less for one test circuit) than the manual layouts of the same circuits without LI. For automatic layout, the area reduction varies between 0 percent and 15 percent and is 6 percent on average. Generally, manual layouts that use LI have more metal 1 I / O ports available. In all but one test case, there were at least 2 more metal 1 I / O ports available. - a) with Local Interconnect 4 b) without Local Interconnect Finally, the use of LI in logic cell design may have a small impact on the yield of resulting circuits through a reduction of the number of contact windows required. For instance, manual layouts that use Fig. 6. Plots of the manual layouts produced for the D-type flip flop cell IEEE TRANSACTIONS ON VERY LARGE SCALE INTEGRATION (VLSI) SYSTEMS, VOL. I , NO. 4 , DECEMBER 1993 569 LI require an average of 22 percent fewer contacts (and as much as 38 percent fewer contacts for one test circuit). This savings is realized at a cost of a number of LI contacts to diffusion or poly, which is comparable to the number of metal 1 contacts saved. In terms of processing, the LI layer can be realized with a single additional mask layer. The net impact on circuit yield and break-even point for area savings are dependent upon specific processes. IV. IMPACT OF LI ON ELECTRICAL PERFORMANCE Because the LI layer has much higher sheet resistance (and possibly higher capacitance) than metal 1, it is important to consider the impact that the introduction of LI wires may have on the electrical performance of the cell. We therefore extracted and characterized all our cells implemented with and without local intercon- 570 IEEE TRANSACTIONS O N VERY LARGE SCALE INTEGRATION (VLSI) SYSTEMS, VOL. I . NO. 4 . DECEMBER 1993 nect. We used nominal values of 14 Q/sq for the resistance and 0.065 fF/pm2 for the capacitance of local interconnect features. We found that the performance degradation was negligible in all cases with these nominal values of capacitance and resistance. We further found that the performance degradation was still acceptable if the resistance and capacitance of the local interconnect layer were increased to ten times their nominal values. This is true for propagation delay, output transit times, and switching current under a wide range of output loading and input transit time conditions. As an example, the extracted circuit for the XNOR circuit realized with the local interconnect layer is shown in Fig. IO. Parasitic capacitors are not shown for readability. The worst case impact is for the fall time of the output signal (OUTB), since a local interconnect resistor is placed in series with the discharge path of the output load (CL). SPICE simulation results for the output fall time of the XNOR gate under three different loading conditions are summarized in Fig. 1 1. The horizontal axis indicates the values of parasitic resistance and capacitance used in the simulation, normalized to the nominal values of 14 Q/sq and 0.065 fF/pm2, respectively. For example, a normalized value of 0 is an optimistic approximation of the performance of the cell without local interconnect. A normalized value of 1 assumes the nominal values of both resist- ance and capacitance, whereas a normalized value of IO indicates that both LI resistance and capacitance values are 10 times their nominal values. V . SUMMARY In this paper we have shown that it is possible to use the local interconnect layer (LI) to make selected connections in digital CMOS circuits. We identified three ways to use LI: i) for con- IEEE TRANSACTIONS ON VERY LARGE SCALE INTEGRATION (VLSI) SYSTEMS, VOL. I . NO. 4. DECEMBER 1993 57 I necting n-device diffusions to p-device diffusions, ii) for connecting diffusion outputs to polysilicon inputs, and iii) for taking polysilicon 1/0 ports directly off diffusions. Layout experiments based on thirteen benchmark circuits have shown that the overall circuit area is reduced in circuits that use LI. Furthermore, global routing flexibility is improved in circuits that use LI because the number of I / O ports available on metal 1 is generally higher. ACKNOWLEDGMENT Micronet is a network of centers of excellence focusing on microelectronics and funded by the Government of Canada. The SUN SPARCstations used for the layout development and experimentation were graciously provided by the Canadian Microelectronics Corporation. This paper is an extendedlrevised version of the paper initially presented at the 1992 International Symposium on Circuits and Systems, held in San Diego, CA, May 10-13, 1992. The authors are grateful for constructive comments made by the reviewers. REFERENCES [ l ] J. R. Lineback. \u201cT1 finds a new way to shrink SRAM cells,\u201d Electronics, 1987, pp. 63-64. 121 R. D. 1. Verhaar, R. A. Augur, C. N. A. Aussems, L. de Bruin, F. A. M. Op den Buijsch, L. W. M. Dingen, T. C. T. Geuns, W. 1. M. Havermans, A. H. Montree, P. A. van der Plas, H. G. Pomp, M. Vertregt, R. de Werdt, N. A. H. Wils, and P. H . Woerlee, \u201cA 25 pm 2 bulk full CMOS sram cell technology with fully overlapping contacts,\u201d Dig. IEDM. 1990. [3] D. D. Gajski and D. E. Thomas, \u201clntroduction to silicon compilation.\u201d in Silicon Compilurion, D. Gajski, Ed. Reading, MA: Addison-Wesley, 1988, pp. 1-48. [4] M. C. Lefebvre and D. F. Skoll. \u201cPicassolI: A CMOS leaf cell synthesis system,\u201d in Proc. MCNC Int. Workshop Layout Synthesis, 1992, pp. 207-219. [5] N. H. E. Weste and K . Eshraghian, Principles of CMOS VLSI Design: A Systems Perspective. Reading, MA: Addison-Wesley, 1985, p. 317. [6] -, Principles of CMOS VLSI Design: A Systems Perspective. Reading, MA: Addison-Wesley, 1985, p. 314. Automated Pin Grid Array Package Routing on Multilayer Ceramic Substrates Changsheng Ying and Jun Gu, Senior Member, IEEE Abstract-As the number of input and output ( I / O ) pins increases, CAD tools are necessary for high-performance packaging design. In this paper we present a routing tool for automated pin grid array (PGA) package routing on multilayer ceramic substrates. Routing is divided Manuscript received November 3, 1992; revised April 30, 1993 and July 13, 1993. This work was supported in part by NSERC Strategic Grant MEF0045793, NSERC Research Grant OGP0046423, and Federal Micronet Grant. The authors are with the Department of Electrical and Computer Engineering, University of Calgary, Calgary, Alta.. Canada T2N 1N4. IEEE Log Number 92 I3 102. 1063-8210/93$03.00 0 1993 IEEE" + ] + }, + { + "image_filename": "designv6_24_0001397_icitacee.2016.7892413-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001397_icitacee.2016.7892413-Figure2-1.png", + "caption": "Fig. 2 Linear ball screw design", + "texts": [ + " Efficiency approaching 978-1-5090-0890-2/16/$31.00 c\u00a92016 IEEE 68 98% so that the gear is widely used to make the drive transmission to the shaft. There are several factors that determine measures on gears can be show in Table 1 as follows: Linear ball screw actuator is a mechanical equipment which uses the rotational motion of the moving objects with minimal friction. Screw on the shaft serves as groove ball bearing so that the removal can be done with precision / high-precision position that can be shown in Figure 2. The Figure 2 shows that the linear ball screw was installed directly with Smartmotor with specific gear box. The design linear lead screw can be seen in Figure 3. The calculation for the soft contact lens inspection can be determined by (1)-(5). SmartMotor is a servo motor in which it is equipped by a servo control system. In a Smartmotor already include servo motor controller, amplifier and encoder. Things that are required to operate a Smartmotor is the power supply, internal programming and serial communication command from the outside (or both)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001587_stherm.2011.5767218-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001587_stherm.2011.5767218-Figure6-1.png", + "caption": "Figure 6: The mounted sensor chip", + "texts": [ + " For measuring the electrical resistance the 4-wires arrangement is used in the chip. The width of the chip is 12.5 mm while that of the active area is 10 mm. The heat flux is measured only in the active area. Advantage of this arrangement is that the obviously uneven parts of the heat flow field on the periphery are excluded from the measurement. The sensitivity of the chip is about 40 Vcm2/W in heatflow sensor mode. In the temperature sensor mode we exploit the above mentioned temperature dependence of the sensor electrical resistance. A photograph of the chip is shown in Figure 6. The chip is bound to a very small printed wire plate, which in turn is connected to the preamplifier. The gluing of the chip has to be made with great care since uneven thermal resistance toward the copper pyramid causes measurement error. 2 HFLHFUHF Vass-Varnai, New Level of Accuracy in TIM Measurements 27th IEEE SEMI-THERM Symposium The theoretical limit of the resolution of the heat flux measurement is the LSB value determined by the amplifiers and the A/D converter. This LSB value is 4 mW. This very good value will obviously be reduced by the electronic noise" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000025_lapc.2011.6114144-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000025_lapc.2011.6114144-Figure3-1.png", + "caption": "Fig. 3 a) AMC unit cell structure: on the top the square patch, in the middle the proposed inner layer and on the bottom the ground plane. b) Top view of the extended inner layer and in dot-line the AMC unit cell.", + "texts": [ + " Keeping the length of the patch constant, (2) and (8) are related by tan(\u03b2'L)>< >>: \u00f03:5\u00de Here, Genetic Algorithm Toolbox in Matlab R2013b [31] is used to get the optimal solution. In this section, the mesh stiffness calculation model is verified by comparing with the FEM results. As shown in Fig. 10, a 2D FE model is established in Abaqus CAE 6.12 using plane stress elements (CPS3 and CPS4R) [30]. The inner ring nodes of pinion and wheel are coupled with the master nodes located at the center of pinion and wheel, respectively. The master node located at the center of the wheel is constrained in all directions, while the master node located at the center of the pinion is only constrained in x axis and y axis with an input torque T being applied. The entire FE model contains 73,486 elements and 44,880 nodes (rint \u00bc 25mm), and such an accuracy is generally considered to be adequate to obtain a result close to the real physical model" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001826_870305-Figure9-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001826_870305-Figure9-1.png", + "caption": "Fig. 9", + "texts": [], + "surrounding_texts": [ + "870305\n5\nTHE SBEAR HODULUS AS \"ONE\" SELECTION CRITERIOK - Not all components are subject to the same requirements. I t is necessary t(\\ be cost conscicJus in the use of the wide range of plas tics (see also Fig. 3). Figure 8 shows an example of how price is dependent upon mRterial characteristics .\nHead Impact Simulator\nFig. 7", + "6 870305\nINTERIOR TRIN OF THE VH GOLF - Figures 9 to 11 give an overview\u00b7 C'f the interior trim of the VH Golf.", + "870305\n\\ VW Golf GL - Luggage Compartment\nfig. II\n7\no Appliciltion for p12sticR in the passenger compartment of the \\'H CoJf - In the folJowing are given a few examples of important applicatiofls of plastics for interior trim and data on the ffi8terials used.\nThe trends listed refer to - Material selection, - Process and prorluction engineering, - Design. o Instrument panel, consoles, trays\nVW Golf C - Injection-Molded Instrument Panel, Individual Components\nFig. i ~" + ] + }, + { + "image_filename": "designv6_24_0001801_j.triboint.2016.11.027-Figure13-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001801_j.triboint.2016.11.027-Figure13-1.png", + "caption": "Fig. 13. Schematic of bearing static load performance measurement test rig.", + "texts": [ + " This is because the bearing needs time to reach the heat balance. The foil temperature reaches 21.5 \u00b0C and stabilized at 21.5 \u00b0C after the moment 300 s. The reason is that the gas film is built between the journal and bearing surface, thus few heat is generated and the bearing achieves heat balance. The total temperature rise is approximately 1.5 \u00b0C, which indicates that the test NSFB has a superior start-stop characteristic. Moreover, the foil temperature stabilizing at 21.5 \u00b0C confirms the feasibility of the test NSFB. Fig. 13 shows a schematic diagram of the bearing static load performance measurement test rig. The test rig is same as the test rig in Fig. 10 except the load device. The midplane of the tested NSFB is connected to a load device, which consists of a micrometer head, two Rotor Load cell Eddy current sensorMicrometer head rods, and a spring, through a strain gauge type load cell. The connection prevents the bearing from rotating caused by the drag torque between the bearing and the rotor. The spring between the NSFB and the micrometer head is used to transform the load and confirm that the applied load is centred at the bearing" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003625_tpel.2015.2477456-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003625_tpel.2015.2477456-Figure3-1.png", + "caption": "Fig. 3. Structure of the MEs: (a) original single-phase ME; (b) proposed two-phase ME.", + "texts": [ + " As a good result, mature control methods for the traditional three-stage wound-rotor generator system can be used in the generation mode of this new proposed integrated starter/generator system. Details about the control methods in the generation mode can be found in [30, 31], and will not be repeated in this paper. ME Design of a two-phase ME was carried out based on an original single-phase ME of a traditional three-stage woundrotor generator. The structure of the original single-phase ME was illustrated in Fig. 3a: The stator field winding was formed from individual coils wound around each of the 12 stator teeth, and the rotor used 12-pole three-phase distributed winding which was connected to the rotating diode rectifier. In order to reduce the redesigned section and get better comparison between the original single-phase ME and proposed two-phase ME, reconfiguration was only undertaken for the stator while the rotor structure and armature winding remained unchanged. Besides, the main dimensions of the stator, such as the active diameter and active length, were constrained to the existing mechanical envelope", + " So the key point of the design lay in the reconfiguration of stator slots and the two-phase field winding. Because the rotor winding structure, containing 12 poles, was kept unchanged, the stator should also have 12 poles after reconfiguration. And in order to reduce harmonic components of the rotating magnetic field generated by the two-phase AC excitation, distributed winding structure was selected for the two-phase field winding. Taking into account the structural strength of the stator and the flux density in stator teeth, 72-slot was selected for the stator as illustrated in Fig. 3b. The two-phase field winding had a spatial difference of 90\u00b0 electrical angle, and the connection diagram was shown in Fig. 4 (only for one pair of poles). The number of coils of the two-phase field winding was chosen to make the two-phase ME have the same output capacity for the MG in the generation mode when excited by the same DC field current as the original single-phase ME. So in the generation mode, the original GCU can be still used and the characteristic of the MG will be the same as the original one" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002203_0010-4655(91)90210-c-Figure25-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002203_0010-4655(91)90210-c-Figure25-1.png", + "caption": "Fig. 25. Variation of the active impedance versus frequency for shunt-loaded arrays with and without the inhomogeneity (a = 0.60 m, d = 0.07 m, Br= 2.8, h = 0.01 m).(a) b=0.47 m, W=0.26 m, XL=O.54m, f(MHz): [240,250,260,270,280,290,300]. (b) b=0.47 m, W=0.26 m, B~=12.8, XL=O.54m, X,.=0.05 m, X", + "texts": [ + " Curve b shows the scan variation of the curve a due to the effects of dielectric loading. It array at the new resonant frequency of 234 MHz, should be noted that in curve c the position of the At this frequency, the unit cell is small enough to shunt relative to the strip\u2019s edge (XF) is also push the blindness nearly to endfire. adjusted in an attempt to keep XF/W nearly Next, the effects of loss, size and position of constant. In figs. 26 and 27, the scan variation for the substrate mhomogeneity are studied. The dithe cases in fig. 25 are shown. The calculations electric material is positioned parallel to the shunt represented by curves a, b and c are performed at strip, spanning the major portion of the substrate\u2019s 386 f-P.R. Bayard, D. H. Schaubert / Infinite arrays of2-D microstrip structures 0.05?., ~r = 2.5, h = 0.01?.). 0.39?., d = 0.05?., ~r = 2.5, h = 0.01?.). thickness (don 0.07 m, Z~= 0.01 m, Zf on 0.06 m). The dimensions of the cell are identical to those of versus angle for the homogeneous case and for fig. 25a. Figure 28 shows the element gain pattern two widths of the substrate\u2019s lossy inhomogeneity, which is centered between adjacent strips. Note that in this case, I R~I would not be a useful XF = 0.05/u quantity because of the substrate loss. As com-- - - XF = 0.07/u -\u2014 = 0.09/u pared to the homogeneous case represented by \u2014-\u2014Xp = 0.12.\\ curve a, the introduction of a small lossy material 1.4 does not affect significantly the scan variation increases, it starts to perturb the fringing fields which are the primary source of radiations" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001651_s001670050083-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001651_s001670050083-Figure1-1.png", + "caption": "Fig. 1 Schematic of the experimental setup with an ankle specimen in the testing apparatus. The instrumented spatial linkage (ISL) was mounted on the tibia proximally and could be attached to mounts which were fixed to the talus or calcaneus distally. Here the distal end of the ISL is attached to the mount in the calcaneus", + "texts": [ + " The objectives of this study were: (1) to measure the forces in the ATFL and CFL during simulated weightbearing in different ankle joint positions; and (2) to simultaneously record and compare the movements in the tibiotalar and subtalar joints. The experimental protocol in this study involved the measurement of ligament force and joint motion during the application of external loads to cadaver ankle specimens. An MTS test machine (Model 858 Bionix Test System, MTS Corporation, Minneapolis, Minn., USA) was used to apply compressive loads to ankle speci- mens held in a specially designed testing apparatus in which the position of the ankle joint could be varied in a controlled manner (Fig. 1). The experimental setup has been described in detail previously [2]. The forces in the ATFL and CFL were measured with buckle transducers [17]. Tibiotalar and total ankle joint motion were measured by an instrumented spatial linkage (ISL) [13, 14]. Specimens Eight fresh human ankle specimens (mean donor age 42 years, range 28\u201357 years) were harvested soon after death, frozen at \u201320\u00b0C, and thawed at room temperature before testing. Specimens with evidence of degenerative disease or ligamentous injury were excluded from the study", + "0 software (Asyst Inc., Rochester, N.Y., USA). Information about calibration procedures and reproducibility of the buckle transducer measurements have been published in a separate paper [2]. In a pilot study performed to develop the technique, we were unable to instrument the PTFL adequately due to limitations in space [6]. Since the PTFL and CFL buckle transducers would likely impinge on each other, the PTFL was kept intact but eliminated from further investigation. Joint motion measurement The ISL (Fig. 1) is a six degree-of-freedom electrogoniometer that measures three-dimensional ankle motion. This device consists of seven metal links interconnected by six revolute joints containing electrical potentiometers. The ISL was secured to the medial side of the tibia by means of a mounting block and two 4 mm diameter screws. The distal end of the ISL was either fixed to a mounting block on two 2.5-mm threaded Steinmann pins which were firmly drilled into the calcaneus, or to an equivalent block and two pins in the talus", + " Throughout the test sequence, the loading apparatus was set to allow unconstrained internal-external rotation of the tibia, as well as unconstrained anterior-posterior and medial-lateral translation of the foot plate. The entire test sequence outlined above was performed with the distal end of the ISL mounted either on the calcaneus or talus. This allowed the measurement of motion between the tibia and calcaneus (total ankle joint motion) and between the tibia and talus (tibiotalar joint motion). In describing the motion about the foot and ankle, we have adopted the definitions of Brostr\u00f6m [3, 25] (Fig. 1). The neutral position was defined as the position where the tibia was mounted vertically in the MTS machine and the foot plate was in a horizontal plane. Internal and external rotation of the foot occurred about a vertical axis (A-A\u2032) through the shaft of the tibia and perpendicular to the MTS table. Supination and pronation occurred about an axis (C-C\u2032) which is described by a line through the foot parallel to the MTS table. Dorsiflexion and plantarflexion occurred about an axis (B-B\u2032) perpendicular to the other two axes in the frontal plane" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000984_access.2018.2882914-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000984_access.2018.2882914-Figure3-1.png", + "caption": "FIGURE 3. Simulated E-field distributions inside the structure at 26 GHz. (a) Without MDs. (b) With MDs.", + "texts": [ + " After an extensive simulation on the parameters of corrugated edges, ds, ls, ld and dt, their optimal values are selected to be 0.7 mm, 0.4 mm, 0.2 mm and 0.7 mm, respectively. Furthermore, two triangular metal patches called metal directors (MDs) are added on the two surfaces of the radiation aperture of the CAVA with GCEs in order to improve the gain at higher frequencies. The corresponding model is shown in Fig. 1(c). To explain effectiveness of the MDs on gain improvement, the E-field distributions inside the structure with and without MDs at 26 GHz are given in Fig. 3. Employing MDs with the proposed structure provides a different propagation environment on the antenna aperture. This pair of MDs strengthens the field coupling between two arms of the antenna so that the EM wave that radiates to the space is closer to the plane wave after the superposition of the coupled fields. A parametric study on the various length (d) and height (h) of triangular MDs for gain variation is performed, resulting in that the highest gain is obtained at d = 4 mm and h = 1 mm (figure not shown for brevity)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003952_978-3-642-33457-3_7-Figure7.8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003952_978-3-642-33457-3_7-Figure7.8-1.png", + "caption": "Fig. 7.8 Rotor design\u2014concept", + "texts": [ + " The stiffness limit was critical for the case of AMB (estimated in order of 105 N/m) while the force limit was critical for the aerostatic bearings. 7. Structural robustness of the rotor was an equally important requirement for the design. Proper retaining of the magnet was crucial, particularly for a high-speed rotor with a high diameter to length ratio. All these requirements affected the design of the motor whose conceptual design is depicted in drawings in Fig. 7.7 The motor concept is explained in the rest of this section (Fig. 7.8). A laminated, slotless stator core has protrusions corresponding to the axial direction for good thermal contact with the housing. Advantages of slotless machines for very-high-speed operation were discussed in Chap. 2. Exclusion of stator teeth removes slotting-effect harmonics from the PM field while, at the same time, reduces impact of armature-field harmonics in the PM rotor. As a whole, a slotless motor is prone to be more efficient and less susceptible to rotor overheating than its slotted counterpart" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001529_iwat.2016.7434821-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001529_iwat.2016.7434821-Figure1-1.png", + "caption": "Fig. 1. Detailed geometry of the proposed antenna: top view (left) and side view (right).", + "texts": [ + " With a full ground plane configuration, the antenna possesses the insensitivity to the human body proximity. To determine the human body\u2019s effect on the antenna, its performance is studied numerically in both free space and on frequency dependent homogeneous human muscle-equivalent phantoms. It is revealed that the antenna bandwidth and gain are very stable with respect to the human body loading as well as structural deformation. Moreover, the antenna has less impact to the human tissue, i.e. low maximum specific absorption rate (SAR). Fig. 1 shows the geometry of the proposed antenna which is designed based on a microstrip technology. It consists of three arc-shaped radiator patches on the top surface and a ground plane printed on the opposite side of a dielectric sheet with relative permittivity of 3 and thickness of 3 mm. Unlike most of the ultra-wideband printed antennas which have partial ground planes for ultra-wide bandwidth enhancement, this 978-1-5090-0267-2/16/$31.00 \u00a92016 IEEE 127 antenna has a full ground plane. It is known that a conventional thin patch antenna with a full ground plane is inherently narrowband" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003745_tap.2007.908537-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003745_tap.2007.908537-Figure5-1.png", + "caption": "Fig. 5. Twyman-Green interferometer.", + "texts": [ + " A final lift-off process was employed to remove the unwanted metal and undeveloped resist. The ZEP layer was then lifted-off in a methylene chloride ultrasound bath and the remaining PMGI layer was removed using EBR PG remover in ultrasound. The phase shift on reflection of the reflectarray stripes, compared to that of the adjacent reference regions between the stripes was measured at 28.3 THz (10.6- m free-space wavelength) using a commercially available Twyman-Green interferometer [18], the Wyko IR3. For a typical Twyman\u2013Green interferometer (Fig. 5), a beam from a coherent light source (a 10.6 m CO Synrad laser for the IR3) is initially expanded and collimated with a two lens telescope, to achieve near plane wave excitation. The source was also linearly polarized along the largest dimension of the reflectarray stripes. The collimated beam impinges on a beam splitter where half of the beam will reflect off of the test device back into the interferometer and the other half of the beam will reflect off of a flat reference surface inside the interferometer, typically a gold mirror for tilt adjustment and correction" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000885_j.apm.2014.04.032-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000885_j.apm.2014.04.032-Figure2-1.png", + "caption": "Fig. 2. The positioning parameters of the ball joints: bolts positioning (a), casings positioning (b).", + "texts": [ + " (1)\u2013(5)) that are required to be known for establishing the kinematic functions in the guiding system. Beside these, for obtaining the angular capability of the ball joints, there are required the positioning parameters of the joints casings and of the axes of the bolts with spherical head; these parameters are defined in the local reference frames. The ball joints M, N and C have the spherical bolts fixed in the wheel carrier, each bolt being positioned in the local reference frame x2y2z2 of the wheel carrier by two angles, e and k, relative to the local axis z2, as follows (Fig. 2a): eM; kM; eN; kN ; eC ; kC \u00f06\u00de The angles e and k have positive \u2018\u2018+\u2019\u2019 or negative \u2018\u2018 \u2019\u2019 sign in the local reference frame x2y2z2, e corresponding to the positive rotation around x2, while k corresponds to the positive rotation around y2; the rotations e and k define the passing z2 ? zp(zM,zN,zC), zp being the bolt axis (Fig. 2a2). In addition, for modeling the orientation as a continuous sequence of two successive rotations, there are represented the angles k0 (k0M;N;C), the second rotation (with k0) being performed around the axis yp obtained after the first rotation (with e), where k0 = arctg (tg k cose). The joints casings M, N and C belong to the lower and upper suspension control arms and to the steering tie-rod. The planes of the casings are positioned in the local reference frames, x1y1z1, x3y3z3 and x4y4z4, by two angles, l and g, relative to the elements planes (Fig. 2b). For the ball joints M and N, the planes of the suspension control arms are (Mx1y1) and (Nx3y3), the positioning angles of the spherical casings in these planes being (Fig. 2b1): l1; g1; l3; g3; respectively l4; g4 for the spherical casing C: \u00f07\u00de In Fig. 2b2, a detail of the scheme shown in Fig. 2b1 is presented, the first rotation (with the angle l) being relative to a line that passes through M/N, parallel with x1/x3 (obtaining the dot-line plane). The second rotation (angle g) is around a line that passes through M/N, parallel with y1/y3", + " The angular capability is defined by two angles, r and s: r represents the angle between the unit vector up of the nut and the normal axis ns to the casing plane (S), while s is used to define the orientation of the unit vector up, being defined between the characteristic line dc of the casing plane and the projection up p of the unit vector up on this plane (Fig. 8a). In these terms, Please degree r \u00bc \\\u00f0 ns; up\u00de; s \u00bc \\\u00f0dc; up p\u00de: \u00f015\u00de The angular parameters r and s will be represented in a circular diagram, with s in trigonometric direction and r having a radial measure, marking a point J for a current pair rj\u2013sj (Fig. 8b). With the previous notations (see Fig. 2a), the unit vectors up of the nuts rods are uM , uN and uC . They are positioned by the angles eM\u2013kM, eN\u2013kN, eC\u2013kC relative to the local axis z2 in the wheel carrier reference frame x2y2z2 (the nuts are fixed in the wheel carrier). In a similar way, the unit vectors ns of the normal axis to the casing plane will be ns1 , ns3 and ns4 (see Figs. 2b and 3). They are positioned by the angles l1\u2013g1, l3\u2013g3, l4\u2013g4 relative to the local axis z1,3,4 in the control arms reference frames x1y1z1, x3y3z3, respectively in the tie-rod reference frame x4y4z4", + " (16) and (17), which are used to calculate the angles rM,N, contain the expressions of the unit vectors in the suspension arms reference frames, so that, in accordance with the specifications in (31) and (32) for the angles l1,3/g01;3 and the corresponding matrices, the normal vector ns1;3 to the casing plane (direction zs1;3 ) will have the following form (in the local reference frame x1,3 y1,3 z1,3): \u00f0ns1 \u00de1 \u00bc MS1 1 \u00f0ns1 \u00des1 \u00bc Ml1 S1 1 h i Mg10 S1 1 h i 0 0 1 2 64 3 75 \u00bc nx1 s1 ;ny1 s1 ;nz1 s1 ; \u00f034\u00de \u00f0ns3 \u00de3 \u00bc MS3 3 \u00f0ns3 \u00des3 \u00bc Ml3 S3 3 h i M g03 S3 3 h i 0 0 1 2 64 3 75 \u00bc nx3 s3 ;ny3 s3 ;nz3 s3 : \u00f035\u00de At the same time, the unit vector of the nut rod uM;N (the direction zM,N) will be expressed in the suspension arm reference frame (x1,3 y1,3 z1,3) by using the sequence of angles k0M;N , eM,N, M2\u20130, w01,3, t001;3 and u1,3 (see the rotation axes in (31)\u2013(33) and Fig. 2), \u00f0uM\u00de1 \u00bc \u00bdMp 1 \u00f0uM\u00dep; \u00f0uN\u00de3 \u00bc \u00bdMp 3 \u00f0uN\u00dep; resulting the following expressions: \u00f0uM\u00de1 \u00bc M/1 0 1 h i M t001 0 1 h i Mw01 0 1 h i \u00bdM2 0 MeM p 2 h i M k0M p 2 h i \u00f0uM\u00dep \u00bc ux1 M ;u y1 M ;u z1 M ; \u00f036\u00de \u00f0uN\u00de3 \u00bc M/3 0 3 h i M t003 0 3 h i Mw03 0 3 h i \u00bdM2 0 MeN p 2 h i M k0N p 2 h i \u00f0uN\u00dep \u00bc ux3 N ;u y3 N ;u z3 N ; \u00f037\u00de in which reference frame \u2018\u2018p\u2019\u2019 reference frame xM,N yM,N zM,N and \u00f0uM;N\u00dep \u00bc \u00f00;0;1\u00de. Eqs. (19) and (20), which are used to calculate the angles sM,N, are expressed in the casings frames xs1;3ys1;3zs1;3, so that the rotations l1,3/g01;3 are added to (36) and (37) (see Fig. 2): \u00f0uM\u00des1 \u00bc M g01 1 s1 h i Ml1 1 s1 h i \u00bdMp 1 \u00f0uM\u00dep \u00bc uxs1 M ; uys1 M ;uzs1 M ; \u00f038\u00de cite this article in press as: P. Alexandru et al., Modeling the angular capability of the ball joints in a complex mechanism with two s of mobility, Appl. Math. Modell. (2014), http://dx.doi.org/10.1016/j.apm.2014.04.032 Please degree \u00f0uN\u00des3 \u00bc M g03 3 s3 h i Ml3 3 s3 h i \u00bdMp 3 \u00f0uN\u00dep \u00bc uxs3 N ; uys3 N ;uzs3 N : \u00f039\u00de Eqs. (18) and (21), which are used to calculate the angles rC/sC, contain the expressions of the unit vectors ns4 and uC in the local reference frame x4y4z4 of the steering tie-rod, as well in the reference frame xs4ys4zs4 of the casing from C" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003317_twc.2013.100313.120861-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003317_twc.2013.100313.120861-Figure7-1.png", + "caption": "Fig. 7. A series of plane waves originated from a distant cluster, impinge on a circular array of radius R. Two horizantally oriented dipoles are present in the array, on position r1 and r2 respectively. The plane waves are characterized by the wavevector k, whose associated angle is \u03d1,\u03d5. Observe that in practical implementations the dipole elements are vertically oriented. The array is mounted on a cylindrical platform.", + "texts": [ + " The analysis of the antenna response directly accounts for the normal Fourier tranform in 2-D. This basis function represents a plane wave that is expandable in a series of cylindrical functions, namely the circular harmonics. As expected, a series of plane waves impinge on the array with a power r and from an angular direction \u03d5k. All plane wave contributions on the array\u2019s aperture are then weighted and integrated to yield the pattern-weighted response. A graphical representation of the above scenario is illustrated in Fig. 7. Note that the array is mounted on a cylindrical platform. The decomposition of a plane wave into circular harmonics on a circular aperture of radius R = r can then be expressed by the following series [23]: eik Tr = eikr cos(\u03d5r\u2212\u03d5k) = \u221e\u2211 n=\u2212\u221e inJn (kr) e in(\u03d5r\u2212\u03d5k), (14) where Jn (\u00b7) denotes the Bessel function of order n. In (14), k is the wavevector of the incoming planewave with k = \u2016k\u2016 = 2\u03c0f/c, where f denotes the frequency of the planewave and r is the position of the observation point. The above equation is also known as the Jacobi-Anger expansion" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001291_bf03378615-FigureI-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001291_bf03378615-FigureI-1.png", + "caption": "Fig. I-Basic configuration of C-B-S single stand rolling mill.", + "texts": [], + "surrounding_texts": [ + "CONTACT-BEND-STRETCH rolling is a new rolling method developed by the General Electric Co. that is particularly applicable to sheet and strip and has special advantages in the fabrication of hard-to roll materials and foil. Since its inception in 1958, a considerable amount of experience has been ac cumulated in terms of rolling of a wide variety of thicknesses-ranging from 0.150 to 0.0002 in. During this time, several experimental mills have been built and tested at the General Electric Research and Development Center, including a fully instru mented 3-in. mill designed for single stand and tan dem operation. Many specific programs have been undertaken, including most recently an investigation of low carbon steel strip and foil rolling. The present report is a description and status of this process as of Dec. 1966. Description of C-B-S rolling Contact-Bend-Stretch (C-B-S) rolling is a process that incorporates plastic bending in conjunction with longitudinal tension and rolling pressure in strip or foil reduction. In addition, it utilizes \"speed ratio\" as a means for determining and controlling reduc tion, in place of the conventional rigid roll gap. These two features lead to a mill concept that is a substan tial departure from methods presently employed in rolling. The configuration of a C-B-S mill is shown in Fig. 1. The strip enters the mill from the right, and is threaded around a large roll, A, called the entry contact roll, then around a small floating roll, called the bend roll, which is cradled in the gap between roll A and another large roll, B, called the exit con tact roll. The strip finally passes around roll Band emerges from the mill. The strip is maintained under tension by means of the entry and exit tension reels, not shown. Sufficient tension is applied at entry and exit to prevent slipping between the strip and the two contact rolls. Rolls A and B are driven at a fixed ratio of surface speed with respect to one another, by means of an appropriate mill drive system, invok ing speed ratio as the means for producing strip re duction at a controlled gage. For example, if roll B is driven so as to have twice the surface speed of roll A, then the strip is reduced in thickness by 50%. Reduction occurs at the two \"bite\" points between the single bend roller and the two contact rolls. Reduction is the consequence of the drawing or stretching of the strip around the small bend roll and forcing it up into the gap between the two con tact rolls where it is squeezed, bent, and rolled suf- L. F. COFFIN, JR. (member AIMEl is a mechanical engineer in the Metallurgy and Ceramics Laboratory, Research and Development Center, General Electric Co., Schenectady, N. Y. 14-JOURNAL OF METALS, AUGUST 1967 by L. F. Coffin, Jr. ficiently at both reduction points to match the speed ratio pre-established by the mill drive system. Specific features of the process are the following: 1) a simple, straight-forward and inexpensive means for supporting a small diameter roll across wide strip widths; 2) direct control of gage during re duction by the use of speed ratio rather than by roll gap; 3) self-adjusting 'and self-determining tensions and roll forces arising during rolling at each point of bite, compatible with a specifically selected speed ratio; 4) the advantages offered by a small diameter roll through lowered roll force needed for a partic ular reduction; and 5) means for producing plas tic bending and the accompanying lowered tensions and roll forces in strip reduction. The basic reason for incorporating plastic bend ing into the process comes as an outgrowth of some early work by the author on the resistance of metals to cyclic plastic strain.I ,2 It was observed during the course of these investigations that the surfaces of test specimens subjected to reversed plastic strain were easily indented by very light pressures, such as those produced by dial gages used to measure the change in dimensions of the specimen. Based on these observations, an exploration of some new and use ful means by which the principles underlying this effect could be exploited was undertaken, Subse quent studies have led to the above described method of rolling in which plastic bending is incorporated. It should be evident that the process described above will function equally well when the degree of plas tic bending becomes negligibly small, so that the absence of plastic bending imposes no limitations to the strip thickness or the bend roll diameter in the operation of the mill. Aside from utilizing small rolls and plastic bending to reduce roll forces, there is an additional feature inherent in the method which becomes important in the rolling of wide, thin strip, and foil. It will be noted that the roll gap, or distance between the two contact rolls, acts only to determine the rela tive position of the bend roll, as given by the angle (x, (called the toggle angle), in Fig. 1. Small toggle angles increase the roll force and decrease the strip tension, while large toggle angles lower the roll force and raise the tension necessary to meet the reduction demanded by the speed ratio. The rolling process is consequently insensitive to roll gap and, in fact, a change in the roll gap has no effect on the percent reduction. Thus, in the support of a wide strip, deflections of the rolls A and B are of little consequence, and the same reduction tends to occur at the center of the strip as at its edges. This is in contrast to conventional rolling where stiffness of the mill frame and roll rigidity are of prime im portance in preventing deflections and roll gap varia tions that directly influence the relative reductions occurring at the center and edges of the strip and result in distortions in the product shape. In the C-B-S configuration, deflections in the contact rolls due to rolling forces are readily accommodated by the deflection of the bend roll to maintain uniform rolling conditions across the strip width. Precise gage control is possible in a C-B-S mill. If the incoming material has uniform thickness along its length, then \"speed ratio\" insures that this uni formity in thickness is preserved. If longitudinal thickness variations are present in the incoming strip, by sensing the strip thickness and adjusting the speed ratio accordingly, longitudinal thickness uniformity can be developed. Another feature of the C-B-S process is the sim plicity in adapting it to tandem rolling. This is shown in Fig. 2 for a 2-stand tandem mill. The basic features of C-B-S rolling are incorporated, and an additional large contact roll C, and a small bend roll 2, have been added. Contact roll C is, of course, driven at a speed ratio with respect to roll B, just as B is driven in speed ratio with respect to A. Additional stands can be added in the same manner." + ] + }, + { + "image_filename": "designv6_24_0003952_978-3-642-33457-3_7-Figure7.2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003952_978-3-642-33457-3_7-Figure7.2-1.png", + "caption": "Fig. 7.2 5DOF magnetic bearings, as proposed by Kimman [6]", + "texts": [ + " It was apparent that, for reaching higher rotational speeds, the spindle length would need to be significantly decreased. However, that was hardly achievable with the same motor-bearings configuration. Hence, Kimman et al. [4] proposed a whole new approach for high-speed spindles: to use a short (disc-shaped) rotor suspended in AMB. The inspiration was found in an idea of 3DOF combined axial and radial magnetic bearings envisaged in [5]. Kimman et al. [4] proposed using such bearings for supporting 5DOF of a disc thus benefiting from reducing rotor tilting and higher resonance frequencies (Fig. 7.2). In essence, it would mean that all the bearings from the original setup\u2014Fig. 7.1\u2014would be grouped around the axial-bearing disc and that would, in turn, drastically reduce the spindle volume. Advantages of using short rotor follow also from analyses of Chap. 5. A rigid short rotor has one critical speed less than its long/slender counterpart and it may also benefit from increase of critical speeds as a result of gyroscopic stiffening (see Fig. 5.12 in Sect. 5.5). Still, the greatest advantage of such a rotor clearly comes from the increase of flexural resonance frequencies as a result of the great reduction of the rotor slenderness. In that way, stability threshold at the first flexural critical speed is too high to be reached and it ceases to be the limiting factor for the rotational speed. The next step in the development of a short-rotor spindle was to integrate an electrical motor into the 5DOF bearings from Fig. 7.2. It was apparent that close spatial integration of AMB and motor was needed. Several concepts for spindle motor were considered including bearingless motors [7, 8] and axial-flux machines. It was a standard, radial-flux PM motor, however, that offered possibility of motor-bearings integration without merging their function and without changing the original AMB concept. The conceptual design of the new spindle is depicted in Figs. 7.3 and 7.4. In proposed magnetic bearings (Fig. 7.3) control actions in axial and radial directions are decoupled: permanent magnets are utilized both to create bias flux and to separate control fluxes in different directions" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003412_mnl.2018.5220-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003412_mnl.2018.5220-Figure1-1.png", + "caption": "Fig. 1 Device architectures a 3D schematic of Si-step-FinFET [18] b 2D schematic of Si-step-FinFET [18] c 2D schematic of C-FinFET [18]", + "texts": [ + " In this work, we analysed the impact of WFV on various electrical parameters for varying gate oxide thickness. Various electrical parameters investigated are: threshold voltage (\u03c3VT), subthreshold swing (\u03c3SS), on current (\u03c3Ion), and off current (\u03c3Ioff) in the presence of WFV of Ti metal gate for Si-step-FinFET. Furthermore, comparative study of electrical parameters of step-FinFET and C-FinFET in the presence of metal gate WFV is also presented. computer-aided design (TCAD) model: Figs. 1a and b show the three-dimensional (3D) and 2D schematic of Si-step-FinFET and C- FinFET, respectively. Fig. 1c gives the 2D schematic of C-FinFET. As observed, step-FinFET has two different fin widths and gate oxide thicknesses, but same fin height for both the devices. Therefore, the fin height is equal for both step-FinFET and C-FinFET, whereas fin width and gate oxide thickness are different. The advantages of the Si-step-FinFET over the C-FinFET is pointed out in our previous work [18]. Simulations were carried out through 3D TCAD numerical device simulator [20], and in the simulator the calibrated TCAD physics model with the fabricated results [21] was considered" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003624_1.23323-Figure26-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003624_1.23323-Figure26-1.png", + "caption": "Fig. 26 Crossflow and crossflow vortices. From Vargas and Reshotko [3,4] (1998).", + "texts": [ + " Near the leading edge the negative pressure gradient for the NACA 0012 swept wing damps out the streamwise turbulence but creates a crossflow velocity that favors the crossflow instability. The crossflow instability appears in regions of strong pressure gradient on swept wings. The wing sweep and the pressure gradient curve the streamlines in the leading edge region, and in the boundary layer the presence of the wall lowers the momentum of the fluid and a velocity profile develops in a direction normal to the inviscidflowvelocity. Thisflow is called a crossflow [30] (Fig. 26). The crossflow velocity profile has a zero velocity at the wall and at the boundary layer edge, with an inflection point in between. The inflection point makes this velocity profile dynamically unstable and causes crossflow vortices. The direction of rotation of the vortices is in a plane normal to the streamline direction. For the NACA 0012 swept wing tip at the conditions of the 1996 experiment, the crossflow instability was the only instability that may be present near the leading edge where the ice accretion is formed" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000645_tsmcb.2006.886947-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000645_tsmcb.2006.886947-Figure4-1.png", + "caption": "Fig. 4. Two examples of imitation across similar embodiments (humanoid). Both demonstrator (left in both examples) and imitator (right in both examples) share the same humanoid embodiment. In the example on the left, the identity mapping is used as the correspondence mapping. In the example on the right, the left arm and leg of the demonstrator are mapped on the right arm and leg of the imitator (and vice versa) with a weight of minus one, resulting in a mirror symmetry. The gray traces visualize the body-part trajectories.", + "texts": [ + " The demonstrator performs a series of actions, and the imitator tries to minimize the correspondence induced relativestate metric. Continuously using the components of S for which \u03b5j = 0 as the current subgoals for each DOF j, the imitator performs actions that attempt to reduce the contribution of error in each such component. Here, the rate of change was restricted to half the componentwise error. Of course, many other selection mechanisms are possible for both immediate or deferred imitation. 1) Identity and Mirror Symmetry Mappings: Two examples of imitation across similar embodiments are shown in Fig. 4. Both demonstrator and imitator are humanoid. In the first example, the identity correspondence mapping is used. In the second example, using the same demonstration, symmetry is achieved by mapping the left body parts of the demonstrator to the right body parts of the imitator and vice versa (see also examples in Fig. 2). 2) Multiple Mappings Between Dissimilar Bodies: The model of an AIBO robot is used as an imitator in the examples shown in Fig. 5. In the first example, the right arm of the demonstrator is mapped on the right front leg of the robot, the left arm to the left front leg, the right leg to the back right leg, and the left leg to the back left leg" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003027_imece2014-38788-Figure12-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003027_imece2014-38788-Figure12-1.png", + "caption": "Figure 12: Solid model of the compliant mechanism designed in ex. 3", + "texts": [ + " Using equation (11) a constrained optimization formulation is generated giving the following spring constants for the energy values of The torsional springs designed above may be translated to either four SLFPs or two fixed-guided segments. Let us consider a compliant mechanism with two fixed-guided segments. We know that for a compliant mechanism with fixed-guided segments and (14) Using equation (14) and assuming , we have Considering rectangular cross-section, width , and thicknesses The resulting compliant mechanism is shown in Figure 12. Figure 13 shows the coupler curve obtained with the PRBM. The precision-position locations obtained from both PRBM and ANSYS\u00ae are shown plotted Figure 13. The strain energy stored in the mechanism at precisionposition is summarized in Table . Table 2: Strain energy stored in the compliant mechanism at various precision-positions Precision Position Strain energy stored (in.-lb.) PRBM ANSYS\u00ae 0 0 15 13.677 45 41.215 A generalized approach for the design of compliant mechanisms has been presented in this work" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003808_5.0027306-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003808_5.0027306-Figure1-1.png", + "caption": "FIGURE 1. Sectional view of a centrifugal compressor", + "texts": [ + " The research results are supposed to be used in the mathematical description of the operation of a thrust bearing when exposed to an external dynamic load. DESCRIPTION OF STAND DESIGN To solve this problem, a modern experimental stand was used, installed in the laboratory of the department \u00abCompressor Machines and Installations of KNITU, Kazan, a detailed description of which is given in article [3]. The stand is based on a single stage centrifugal multiplier compressor for compressing air with low performance (Fig. 1). The gas system of the stand consists of a suction pipe 1, a impeller 2, a bladeless diffuser 3, an annular rectangular collecting chamber 4, an expanding outlet pipe 5 and a discharge pipe about six meters Oil and Gas Engineering (OGE-2020) AIP Conf. Proc. 2285, 030016-1\u2013030016-7; https://doi.org/10.1063/5.0027306 Published by AIP Publishing. 978-0-7354-4015-9/$30.00 030016-1 long, divided into two sections. The first section ends with a full bore valve, and the second ends with a manual rotary shutter" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000298_1.1807415-Figure9-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000298_1.1807415-Figure9-1.png", + "caption": "Fig. 9 Pressure distribution under hydrostatic-operating condition: \u201ea\u2026 pre- and post-orifice, \u201eb\u2026 bearing surface", + "texts": [], + "surrounding_texts": [ + "A squeeze film damper ~SFD! is a special kind of fluid-film bearing, which exclusively makes use of only the squeeze effect resulting from the whirl of the journal to influence the vibration of the supported rotor systems. Squeeze film dampers are utilized in Table 9 Comparison of results by two solutions Hydrostatic-operating Hybrid-Operating CFXTASCflow VTEXPRESS % diff CFXTASCflow VTEXPRESS % diff #1 Pocket inlet pressure ~MPa! 2.657 2.650 0.3% 2.671 2.724 1.9% #2 Pocket inlet pressure ~MPa! 4.104 4.165 1.5% 3.723 3.745 0.6% #3 Pocket inlet pressure ~MPa! 4.959 4.975 0.3% 4.980 4.987 0.1% #4 Pocket inlet pressure ~MPa! 4.104 4.165 1.5% 4.413 4.503 2.0% Load capacity ~N! 722.7 638.1 13.2% 727.0 670.9 8.3% Attitude angle ~deg! 0 0 0% 23.4 25.7 8.9% Flow rate ~kg/s! 0.0324 0.0306 5.8% 0.0318 0.0304 4.6% Transactions of the ASME 13 Terms of Use: http://asme.org/terms Downloaded F jet aircraft engines to provide needed damping to control undesirable vibrations. They also are used in many land-based rotating machinery to enhance stability. Example 4. Assume the diameter of a long damper D530 mm ~1.18 in.! with radial clearance h050.03 mm ~1.18 mil!. The dynamic viscosity of lubricating oil is m59.9331023 N s/m2 (1.47 31026 lbf s/in.2). The radius of whirling orbit is e50.02 mm ~0.79 mil! and its circular frequency is v5200 rad/s ~'1910 cpm!. By employing the \u2018\u2018Moving Grid\u2019\u2019 capability and the user\u2019s FORTRAN code, the dynamic simulation of this SFD can be conducted. Some of the transient images of dynamic pressure at different instants are shown in Fig. 12." + ] + }, + { + "image_filename": "designv6_24_0001835_j.mechmachtheory.2012.11.004-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001835_j.mechmachtheory.2012.11.004-Figure2-1.png", + "caption": "Fig. 2. Cam mechanism with flat-faced translating follower and its kinematic inversion.", + "texts": [ + " In the second part of the work, a constant-breadth cam mechanism with an oscillating follower is studied, and a new numerical example is given of the design of a displacement function with the corresponding constant-breadth profile. The software Mathematica\u00a9 is used as a calculating tool and for graphical representation. Cardona and Clos [13] show the parametric expression for obtaining the cam profile that drives a flat-faced translating follower. The cam profile is obtained conceptually by way of a kinematic inversion in which the cam is fixed, Fig. 2. Two coordinate systems are used, one x, y in which the cam is fixed and the other 1, 2 fixed and oriented according to the follower guide. The components of the position vector OP(\u03b8) of the cam-follower contact point P are known in the coordinate system 1, 2, and are expressed in the coordinate system x, y fixed to the cam via the rotation matrix. This expression is: OP \u03b8\u00f0 \u00def g1;2 \u00bc s\u03b8 \u03b8\u00f0 \u00de s \u03b8\u00f0 \u00de 1;2 OP \u03b8\u00f0 \u00def gx;y \u00bc S\u03b8\u00bd \u00b7 OP \u03b8\u00f0 \u00def g1;2; S\u03b8\u00bd \u00bc cos \u03b8 sin \u03b8 \u2212 sin \u03b8 cos \u03b8 : \u00f01\u00de In the constant-breadth cams that drive a parallel flat-faced double translating follower (Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001905_2010-01-0530-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001905_2010-01-0530-Figure7-1.png", + "caption": "Figure 7. Sprungmass deflections during vibration (front view).", + "texts": [ + " In this test a new digital analysis was developed to separate the roll and bounce modes in lab. Two accelerometers are mounted on the right and left sides of the model in order to measure the right and left accelerations ( and ) of the model due to impulse force. A vertical impulse force is applied to the sprung mass center of gravity using a falling rubber ball from a certain height. The test rig used is shown in Figure 6. To obtain the separated roll and bounce accelerations and , the following mathematical processing is applied to the obtained signals using a digital analyzer (Figure 7): Where, 2B is the distance between the two accelerometers. The response of the measured bounce and roll acceleration frequency spectrum are shown in Figures 8a & 8b. These results show a good agreement between the analytical model and the developed lab model. The bounce natural frequency is 5.2 Hz and the roll natural frequency is 1.95 Hz. The turntable test is proposed for the purpose of measuring the variation of the vehicle roll angle \u03a6 with the angular speeds . The VLM was put on the turntable device (Figure 9a)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001425_s1759078719001107-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001425_s1759078719001107-Figure5-1.png", + "caption": "Fig. 5. The Q-slot and the complimentary. (a) Q-slot. (b) The complementary.", + "texts": [ + "93 GHz (black curve in Fig. 4), which is approximately 52% of the 3rd band of 4.7\u20136.8 GHz. The ARBW values shown in Fig. 4 match the above studies about the surface current distribution in Fig. 2 and the LHP\u2013RHP represented in Fig. 3. The Q-slot is almost similar to the elliptical slot. Therefore, its radiation pattern will be likened to the radiating antenna for the elliptical slot. According to the previous study, the complementary antenna for the elliptical slot is the elliptical loop. As it is shown in Fig. 5. In rectangular coordinates, the ellipse can be https://www.cambridge.org/core/terms. https://doi.org/10.1017/S1759078719001107 Downloaded from https://www.cambridge.org/core. The Librarian-Seeley Historical Library, on 20 Aug 2019 at 04:20:51, subject to the Cambridge Core terms of use, available at presented by the following parameters X = a cos u, Y = b sin u, \u2212p \u2264 u \u2264 p, (9) where a, b are the minor and major axes of the ellipse. In terms of rectangular coordinates, the x component of the Hertzian vector potential at point P can be written as [8] Ax = e\u2212ibR i4pRwe \u222bu=p u=\u2212p Io e [ik (x cos \u2205 sin u+ y sin u sin \u2205)] dx, (10) where Io is the current in ds (an element length) = 2\u03c0/ \u03bb" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003381_tia.2004.827476-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003381_tia.2004.827476-Figure1-1.png", + "caption": "Fig. 1. Experimental rotor.", + "texts": [], + "surrounding_texts": [ + "The photograph of an experimental IPM rotor, the cross section of a quarter of the motor, and the demagnetization curve of the NdBFe magnet are shown in Figs. 1\u20133, respectively. A frame size of a 600-W three-phase four-pole Y-connected, 50-Hz 200-V 4-A squirrel-cage induction motor was used for 0093-9994/04$20.00 \u00a9 2004 IEEE testing the experimental rotor. The four-pole magnets arrangement in the rotor is oriented for a high-field-type IPM synchronous motor [2]. The experimentally developed rotor has the following distinctive design features. 1) The fluxes from both sides of the magnet are concentrated effectively in the middle of the magnetic poles of the rotor. 2) The reluctance of the axis is larger than that of the axis, because the -axis flux passes across the magnet with high reluctance. Large reluctance torque can be obtained [9]. 3) The large pull-in torque can be obtained due to deep cage bars in the rotor. 4) The conducting material between the magnet and the rotor core is made from aluminum and has both functions of the flux barrier and cage bar. In particular, the conducting material on the magnet near the air gap decreases the leakage flux in the rotor iron core and causes large reluctance torque. These features are different compared to the configurations of the other line-start PM synchronous motors [10]. Furthermore, the number and configuration of rotor slots have been successfully designed by using the finite-element method so that the waveform of the electromotive force (EMF) due to the PMs was close to the sine waveform and the cogging torque was low." + ] + }, + { + "image_filename": "designv6_24_0002818_mop.32365-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002818_mop.32365-Figure6-1.png", + "caption": "FIGURE 6 The employed model of the proposed slot antennas to derive the relation between the slot voltage and the amplitude of the scattered waves", + "texts": [ + "1,8 However, when the feeding waveguide has nonuniform cross section along the direction of the wave propagation, finding the analytical relation between the slot voltage and the amplitude of the forward and the backward scattered dominant mode off the slot is not an easy task. Instead of the analytical relation, a method has been proposed in References 43-45, which uses only the gathered scattering data of a model of the suggested slot antennas in an available electromagnetic field simulator and performing some mathematical calculations. This procedure has been successfully applied to the suggested slot antennas. To achieve this goal, the model of the suggested slot antennas is fed via a lumped port at its middle as shown in Figure 6. The developed transverse electric field along the slot is a half-cosinusoid distribution. Figure 7 shows the amplitude of the transverse electric fields along the slot when the post offsets were equal to 22 and 28 mm. For the former case, the spacing between the posts and the slot length were equal to 40 and 54 mm, respectively. For the latter case, the spacing between the posts and the slot length were equal to 41 and 50 mm, respectively. For comparative reasons, the half-cosinusoidal function is also added to the figure" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003385_tmag.2013.2238897-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003385_tmag.2013.2238897-Figure2-1.png", + "caption": "Fig. 2. Cutaways of the LRPMA along (a) the linear direction and (b) the rotation direction.", + "texts": [ + " On the tubular mover, there are forty-eight alternately polarized permanent magnets (PMs), with six stacks in the -direction and eight in the -direction. There are eighteen coils, three stacks in the -direction and six in the -direction. The windings are housed in either an air-cored or iron-cored stator as shown in Fig. 1. A knowledge of the magnetic field distribution produced by the tubular PM mover is fundamental to establishing an accurate model of the LRPMA for design optimization and dynamic modeling. Without loss of generality, an air-cored LRPMA is considered in this paper. Fig. 2 shows the cutaways of the aircored actuator along the linear and rotational directions. Region I is air/winding region and region II is the PM region. In the 3-D model, the soft-magnetic parts are considered to be infinitely permeable, 0018-9464/$31.00 \u00a9 2013 IEEE and the boundaries where and are of the Neumann boundary; the end effects are not taken into account; the PMs have a linear demagnetization characteristic, and are fully magnetized in the direction of magnetization. The relationship between the magnetic field intensity vector and flux density vector in the two regions are given by (1) where, is the permeability of free space, is the relative permeability of PM, is the residual magnetization vector of PM", + " 3 can be expressed by (2) is governed by the Laplacian equation in region I and the Possion equation in region II is (3) (4) In the 3-D cylindrical coordinate system, can be decomposed into magnetization components and in the and directions, respectively. In PM region II, of adjacent PMs is opposite and its distribution is a homogeneous harmonic odd function, in both and z directions. Using Fourier series, the distributions of magnetization component in the PM region is (5) where is the pole pairs number in the x direction; is the pole pitch in the z direction. Considering that and , we have (6) where is the remanence of PM; and are the pole-arc coefficients in the and z directions, respectively. The boundary conditions in Fig. 2 can be expressed by (7) where , and . By solving (3) and (4) subject to the distributions of magnetization components (5) and the boundary conditions (7), the Fourier expansion of and directions flux density components in airgap region I in the 3-D cylindrical coordinate system are given by (8) (9) (10) where and and are the modified BESSEL functions of first and second kind and , which are the derivative of and , are expressed approximately by The coefficient is given by the equation shown at the bottom of the page, where To verify the correctness of the analytical solutions, the magnetic field distributions of the prototype are analyzed by 3-D FEM as shown in Fig", + " The electromagnetic force and torque exerted on the rotor, resulting from the interaction between the current in a stator winding and the rotor magnetic field, are given by (11) (12) where denotes the current density vector in the winding. The current density is given by (13) where is the maximum current density in the winding, and are angular frequencies along the and directions, and represent the phase-shifts between the flux linkage and the currents. Each winding comprises a number of circular turns distributed on the cylindrical stator, and occupies an area where (or ) and (or ), as shown in Fig. 2. Considering a single turn with infinitesimal cross-section , the total electromagnetic force and torque produced by a winding can be obtained from the following integration: (14) (15) It can be shown that is dominated by its fundamental component, viz. and . The results of (14) and (15), neglecting high order harmonic, can be rewritten as (16) (17) where and , respectively the amplitude of linear electromagnetic force and rotary electromagnetic torque, can be given by (18) As can be seen from (18) and (19), the amplitudes of the linear electromagnetic force and rotary electromagnetic torque depend upon and as well as the geometrical parameters of the rotor, the stator and the winding" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000740_igarss.2013.6721289-Figure2.1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000740_igarss.2013.6721289-Figure2.1-1.png", + "caption": "Fig 2.1 Mission of the SDS-4", + "texts": [ + " This program is also of key strategic importance for JAXA\u2019s operational and science satellites, enabling the detection of unexpected risks, which could lead to disastrous results, and will pave the way for the adoption of next-generation space technologies. SDS satellites are designed, hand-constructed, tested, and operated by JAXA\u2019s young engineers. This program is also positioned as a part of one of capacity building. The missions of the SDS-4 spacecraft are to demonstrate a Space-based Automatic Identification System Experiment (SPAISE), Quartz Crystal Microbalance (QCM), a Flatshaped heat pipe On-orbit Experiment (FOX), and In-flight experiments related to Space materials using THERME(IST). Each mission component is shown in Figure 2.1 The importance of maritime security has increased. In Japan, maritime security and vessel traffic management are crucial because Japan is a maritime nation with a vast Exclusive Economic Zone (EEZ), and because many ships come from various countries for trade purposes. Traditional shore stations have collected AIS information from ships within a limited range and limited distance (typically about 74km). However, the Japanese EEZ is larger than the AIS area of these coverage. SPAISE is proposed to resolve this problem with satellite based AIS-monitoring" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002990_s0094-114x(02)00037-x-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002990_s0094-114x(02)00037-x-Figure7-1.png", + "caption": "Fig. 7. Superimposed positions of the suspension mechanism solution 1, corresponding to DzN \u00bc 0 and DzN \u00bc 150 mm, viewed from the rear (a) and from above (b).", + "texts": [ + " Beale / Mechanism and Machine Theory 37 (2002) 815\u2013832 827 DY \u00f0zN\u00de \u00bc yS\u00f0zN0\u00de yS\u00f0zN\u00de \u00f020\u00de This parameter describes the fore and aft motion of the wheel during jounce and rebound. However, since it occurs along the direction of car travel, is has a smaller effect upon the car dynamics than the wheel track alteration. The camber angle variation Dd was determined as the projection of the angle between the axes Oz and Nz0 on the vertical transverse plane (Fig. 6a). Similarly, the toe angle alteration Du shown in Fig. 7b was determined as the angle between the axes Ox and Nx0 projected on the horizontal plane. In this case, for 150 mm6DzN6 150 mm the toe angle of solution 1 is slightly larger than that of the existing solution 0, being however compensated by the understeer effect of track widening during jounce. For illustrative purposes, the diagrams of the magnitude of the angular velocity x and angular acceleration e of the wheel carrier have been plotted (Fig. 6) for _zN \u00bc 1:0 m/s and \u20aczN \u00bc 0 using Eqs.(A.1) and (A.2) in Appendix A. The results of the kinematic analysis have been also used in the 3D visualization of the motion of the mechanism. Fig. 7 shows superimposed positions of the suspension mechanism solution 1, corresponding to zN0 and zN0 150 mm, viewed from the front (a) and from above (b). The circle-point-surface and the center-point-surface in Fig. 3 were produced for solution 2. They have been generated as ruled surfaces of the momentary screw axis relative to the chassis 828 P.A. Simionescu, D. Beale / Mechanism and Machine Theory 37 (2002) 815\u2013832 (the circle-point-surface) and to the wheel carrier (the center-point-surface). The inclined position of the screw axis relative to car\u2019s longitudinal axis it is due to the wheel carrier rotation around its own axis, which for solution 2 corresponds to a maximum angle c of 16" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001116_022106-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001116_022106-Figure1-1.png", + "caption": "Figure 1. Visualization of turbine wheel dimensional chain (blade-disc connection).", + "texts": [ + " To get a functional unit subassembly, it is necessary to ensure the commensuration of assembly parameters at the actual assembly process; in particular, they should be in the tolerance range and have the values specified in the technical specification of the drawings for the parts used for such assembly [11]. In case of the turbine wheel, the following values are important: gap between shroud platforms of adjacent blades, gap between attachment platforms of adjacent platforms and draw between the contact surfaces of blade shroud platforms. If assembly parameters deviate from the specifications, the assembly may result in a wedge of 5-7 blades. Let us consider dimensional chains by the example of the draw at blade shroud platforms. Figure 1 shows a dimensional chain with a gap between shroud platform surfaces. The figure demonstrates a part of the disc with two blades. The dimensional chain starts acting from the center of adjacent blade platform contact on the shroud surface towards the attachment plane on one of the blades (V1). Further, the chain is going the same way on the shroud surface towards the attachment end plane (V2). Then it passes to the groove center on the disc rim (V3). At this moment a blade swing is registered. To do this, one uses a vector with zero coordinates and having only a turning angle deviation" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000566_ijhvs.2019.102689-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000566_ijhvs.2019.102689-Figure8-1.png", + "caption": "Figure 8 Stroke amplification mechanism in ADAMS (see online version for colours)", + "texts": [ + " Therefore, an amplification mechanism is designed for enlargement and to convert the vertical vibration into a horizontal movement for easier installation too. The designed stroke amplification mechanism is as shown in Figure 7. The upper part of the rod is connected to the frame with a sliding key, and the lower part takes the hinge connection with the slider. During the slider moving vertically, the crank pin in the curved hole drives the crank, the piston rod being driven to do compression movement. The body can avoid the \u2018dead centre\u2019 effectively. A multi-body simulation model of the mechanism is creating in the ADAMS software, as shown in Figure 8. According to Standard QCT545-1999, the rod is defined as a step movement of 25 mm within 0.15s. The vertical movement of MAKER_MOV1 in the rod, the rotating angle of MAKER_MOV and the horizontal movement of MAKER_MOV2 in the crank\u2019s upper and lower connection points respectively are shown in Figure 9. In the absence of the link, L (MOV1) = 25 mm, L (MOV2) = 33.12 mm, the amplification ratio is 1.32 times. Link with 45\u00b0 angle is added after optimisation, obtaining the horizontal displacement of the link centroid L (cm) = 44" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001394_tmtt.2003.809667-Figure9-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001394_tmtt.2003.809667-Figure9-1.png", + "caption": "Fig. 9. Graph representing relation between the two configuration coupling lengths ( = 2:6).", + "texts": [ + " 6 and 8 and using (17), the two filter bandwidths, we have calculated (18) for the classic SIR filter of Fig. 8(a) and (19) for the semiloop one of Fig. 8(b). The equality between these two formulas gives a simple condition (20) on the return-loss parameters as follows: (18) (19) (20) The two parameters and have been calculated using the topology of Fig. 8 and using an accurate numerical model of coupling areas [7]. Equation (20) between and then gives conditions on and according to the coupling coefficient . An example of the result of this intricate solving is presented in Fig. 9 for a substrate of relative permittivity equal to 2.6. is linked to even and odd impedances by (1). Respecting these conditions, semiloop filters are more compact than classic SIR filters without changing resonant frequencies and bandwidths. The chosen filter structure is represented in Fig. 8(b). In this section, the different steps of a semiloop filter design are detailed. Step 1) The first step is the determination of the input coupling electrical length at . The larger , the smaller the bandwidth", + " At the same time, input coupled lines create a null in the frequency response at a frequency corresponding with . This null must be kept distant from any resonant frequencies. Step 2) The next step is how to obtain the coupling coefficient . The larger , the larger the bandwidth. However, is limited by technological considerations, like minimum spacing between coupled lines. When is set, and are fixed. The impedance of a single line with the same width than coupled lines can also be calculated using a numerical model. Step 3) With Fig. 9, is deduced with and . Then, . This chart must be plotted for the relative permittivity that corresponds to the realization. Step 4) Finally, the resonator constituted by two lines and one line imposes the resonant frequencies. If two frequencies are set, and are given using (3) and (4). In this section, an example of design on Duroid is given. The power divider works at 0.9 and 1.94 GHz. A. Dual-Band Power-Divider Design 1) The zero introduced by the input coupler is set to 1.5 GHz leading to . 2) In order to have the desired bandwidth , is set to 0.3 (0.2 mm between coupled lines), which is squaring with . Thus, the coupled linewidth is 3.7 mm, which fits with a 55- linewidth . 3) Fig. 9 gives and, thus, . 4) Finally, , , and values are introduced in (3) and (4) and give , , and . The corresponding semiloop filter is given in Fig. 10. The method used in Section II-C to transform this filter into a three-port power divider is applied and optimized with the use of circuit simulation software Genesys (Eagleware, Norcross, GA). It gives the final structure represented in Fig. 11. B. Dual-Band Power-Divider Simulated and Measured Performances The power divider has been simulated and realized" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002295_b978-0-08-011216-9.50010-9-Figure7.4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002295_b978-0-08-011216-9.50010-9-Figure7.4-1.png", + "caption": "FIG. 7.4", + "texts": [ + " \u03c3, = \u0399\u0393 [ \u03af 1 / 2 \u038a - sin (\u03c6 - 0) + sin2(0 - 0) + {1 - cos (\u03c6 - 0)} 2 nd3 or \\6WRr . f\u00b1 \u039b\u03bb f ^ ,\u00b1 \u00cc 1 / 2 * i = \u2014 3 - sin (\u03c6 - 0) + 12 - 2 cos (\u03c6 - 0)1 (7.11) The greatest value of \u03c31? from equation (7.11), is found to occur when \u03c6 \u2014 0 = 2\u03c0/3 and 4\u03c0/3. For the case of the bar forming a semi-circle \u03c6 = \u03c0, and there fore aj has its greatest value at 0 = \u03c0 \u2014 2\u03c0/3 or 0 = 60\u00b0. An important application of the bending and torsion theories derived in earlier chapters occurs in the case of the helical spring. The geometry of a spring is shown in Fig. 7.4. The centre-line of the wire forming the spring is a helix on a cylindrical surface, such that the helix angle is \u03b2. A portion of the wire is shown, and on the cross section at O mutually perpendicular axes Ox, Oy and Oz are set up. Perhaps the most common form of loading on a spring is a force W along the central axis UV. Since this force acts at a distance 7?, the coil radius, from the axis of the wire there will be moments set up as follows Mz = WR cos jS (causing torsion in the wire), (7.12) My = WR sin \u00df (causing bending about the y-axis and a change in R), (7", + "18) the force on the spring is w / SGd* 12 x 106 x 0-1254<5 W = 6 ^ = 64 x 12 x 0-6133 = 15'6<5 Ib* \u03b4 Strain in the copper wire = ccT \u2014 \u2014. Li Therefore, force on the wire S = /l7-5 x IO-6 x 50 - ^ \\ - x 0-01252 x 15 x EA \u03b4\\ \u03c0 10e = 1-61 \u2014 36-8(5. For equilibrium, force in wire = force on the spring, therefore 15-65 = 1-61 \u2014 36-85 \u03b4 = ~ = 0-0307 in. Therefore the force in the system = 15-6 x 0-0307 = 0-479 lb. Total strain energy = 2 x \\\u03a8\u03b4 = 0-479 x 0-0307 in. lb \u00ab= 0-0147 in. lb. If a helical spring is subjected to a couple M about its axis UV9 then referring to the geometry of the spring in Fig. 7.4, the resolved components of the couple about the coordinate axes at section O are Mz = M sin \u00df (causing torsion of the wire), (7.20) My = M cos \u00df (causing bending of the coils to a smaller or larger radius), (7.21) Mx = 0. Since in the case of the close-coiled spring \u00df is small, Mz ~ 0, My ~ M9 so that the maximum bending stress in the wire is (My)max 32M The stiffness of the spring when subjected to an axial couple will now be considered using the same approach as in section 7.5 for the stiffness under axial force" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000080_apmc.2008.4958570-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000080_apmc.2008.4958570-Figure3-1.png", + "caption": "Fig. 3. E-plane and H-plane patterns of right (left) and top (right) Yagi arrays at 5.2 GHz.", + "texts": [ + " The beauty of this configuration is that when all four sides are excited conical radiation can be generated that can cover a room with a higher gain than the standard omnidirectional monopoles of wireless routers. The radiation pattern of one microstrip Yagi array in the steerable antenna design becomes an important characteristic to examine because of the presence of the other arrays and the possibly of blockage or feedline radiation. The radiation patterns of the right Yagi array and the top Yagi array are displayed in Fig. 3 taken at 5.2 GHz. In the E-plane of the right Yagi array, the 3-dB beamwidth is around 61\u00b0 from (4.5\u00b0-65.3\u00b0). The 3-dB beamwidth in the Hplane of the same array is around 57\u00b0. For the top array, the 3-dB beamwidths of the E- and H-planes are approximately 46\u00b0 and 57\u00b0, respectively. The H-plane beamwidths of both arrays agree well with each other, but there is a difference of around 15\u00b0 in the E-plane patterns. This difference can possibly be attributed to the position of the diagonal feedline due to the sharp discontinuity in the transition from the diagonal feedline to the feed of the right Yagi array" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002424_iecon.2017.8217456-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002424_iecon.2017.8217456-Figure3-1.png", + "caption": "Fig. 3. Model of a controm moment gyro", + "texts": [], + "surrounding_texts": [ + "Fig. 1 shows that the cargo with translation sway and twist sway. 0 = [x0 y0 z0]T denotes Cartesian coordinates fixed on the earth with x axis pointed upward in perpendicular direction. A = [xA yA zA]T, s= [xs ys zs]T and c= [xc yc zc]T denote those fixed on the supporting point A, on the sensor and on the cargo with axis x pointed downward in perpendicular direction. Here, the relation between 0 and A is A = X( ) 0 because these x axes are set so that these axes point toward the same direction. When roll angle P, pitch angle Q and yaw angle R by the inertial measurement unit were measured, any attitude of the cargo is expressed as the follow equation using ZYX Euler\u2019s angle matrix. s sss sss sss s T P T Q T R MMM MMM MMM XYZ 333231 232221 131211 )()()(A )()(33 )()(32 )(31 )()()()()(23 )()()()()(22 )()(21 )()()()()(13 )()()()()(12 )()(11 PQ s PQ s Q s PQRPR s PQRPR s QR s PQRPR s PQRPR s QR s CCM SCM SM CSSSCM SSSCCM CSM CSCSSM SSCCSM CCM" + ] + }, + { + "image_filename": "designv6_24_0003027_imece2014-38788-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003027_imece2014-38788-Figure2-1.png", + "caption": "Figure 2: A partially-compliant mechanism in its (a) original form and (b) equivalent rigid-body mechanism form.", + "texts": [ + " The largest benefit in the use of PRBM concept comes from considering compliant mechanisms as equivalent rigid-body mechanisms with characteristic compliance (discrete springs). This facilitates the use Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/conferences/asmep/83107/ on 03/30/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use 3 Copyright \u00a9 2014 by ASME of the wealth of existing rigid-body mechanisms analysis and synthesis knowledge to the treatment of compliant mechanisms [8]. Figure 2(a) illustrates a partially-compliant mechanism in its undeformed state. This mechanism consists of one fixed-pinned segment, one curved SLFP, one straight SLFP, two rigid segments, and one revolute joint. Figure 2(b) shows the PRBM of this partially- compliant mechanism. Such transformations enable the use of state- of-the- art rigid-body mechanism synthesis and analysis techniques, along with the energy/torque equations, to design compliant mechanisms for conventional tasks with either specified energy storage or torque/force-deflection characteristics. The PRBM representation of a compliant mechanism facilitates in determining its mobility and energy storage or force/torquedeflection characteristics, henceforth referred to as kinematics and compliance, respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000862_optronix.2019.8862342-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000862_optronix.2019.8862342-Figure4-1.png", + "caption": "Fig. 4. Schematic representation of HMSIW based semi-circular antenna excited with inset feed.", + "texts": [ + "17 GHz as given in Fig. 2. The bandwidth of the impedance varies from 5.19\u20135.25 GHz covering 60 MHz .The radiation pattern obtained after simulation is shown in Fig. 3 where in the co-polarized gain turns out to be 6.5 dBi of the simulated E-plane while it is measured to be of a value of 6.2 dBi. For the same H-plane has the values 5.3 dBi and 5.1 dBi. For both the planes shown below the gain of cross-polarization value is below 30 dBi respectively. The same antenna excited with inset feeding is shown in Fig. 4. The inset feed placed along the centre line excites the fundamental TM010 mode at 5.51 GHz. The fabricated antenna prototypes are shown in Fig. 5. For the excitation of fundamental TM010 mode, the inset feed of width W1 = 0.58 mm is inserted at an optimized distance of L3 = 7.5 mm from the base of the semi-circle to match with 100 ohm impedance. The various measurements of the mentioned antenna are showcased in Table II. ANTENNA WITH INSET FEED (IN MM) The fundamental resonating frequency is 5" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003920_tim.2020.3039641-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003920_tim.2020.3039641-Figure4-1.png", + "caption": "Fig. 4. Illustration of different camera models: a) pinhole camera model in the Air scenario; b) focus adjustment in underwater scenario, the red line is the light-path for virtual camera; c) refractive camera model in underwater scenario.", + "texts": [ + " P can be obtained with image captured in the light-path Air. gv can be used as error function. By following the solution of Eq. (3), stacking multiple Eq. (6) can also yield a linear system. We can thus compute the interface distance d0. To calibrate the laser light plane in the same camera frame, we need to estimate the checkerboard pose via considering medium refraction explicitly. Generally, there are two different camera models for the underwater 3D reconstruction applications (i.e., the pinhole model and the refractive model) as shown in Fig. 4. However, due to the variation of the medium density, the first method changes the camera reference frame and cannot calibrate the sensor parameters in the same camera reference frame as shown in Fig. 4(b) [24]. Therefore, we use refractive model to establish a unified sensor reference frame between water and air. Based on the refractive model, we perform the refractive pose estimation in two scenarios (i.e., AGA, and AGW) as shown in Fig. 2(b). The core of estimating refractive pose is to formulate the forward and back projection equation, where forward projection projects a 3D point from the world reference frame to camera reference frame and backward projection do the opposite projection. More specifically, we first formulate three forward projection equation (i" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002384_1.5031999-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002384_1.5031999-Figure1-1.png", + "caption": "FIGURE 1.Geometry of proposed antenna. (a) Front side radiating element (b) Bottom side view of ground plane", + "texts": [ + " The dual band is obtained by cutting a ring of rectangular shape and a circular slot in the patch and return loss is improved by cutting two rectangular slots in ground. The fabricated to verify the simulated results and it is found that there is a good agreement in simulated and measured result. The antenna proposed can be use for the application of 2.4/5.2/5.5/5.8 GHz WLAN application. The antenna proposed has better performance characteristics than other existing antenna which is shown in table 1. The presented low profile antenna has small volume and wider bandwidth than other antennas. Figure 1(a), (b) shows the geometry of front side and bottom side of antenna proposed with all dimensions. The antenna consists of a rectangular patch of dimension Wp\u00d7Lp with a rectangular ring of width g and a circular slot of radius R on one side of substrate with a truncated ground plane on other side. The ground plane having dimension Wsub\u00d7Lp with two rectangular slots of B\u00d7C. The antenna is fabricated on the substrate FR4_epoxy relative t and easily available substrate. The thickness of . The optimized dimensions of antenna obtained are shown in table 2" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000863_bf02169812-Figure6.1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000863_bf02169812-Figure6.1-1.png", + "caption": "Figure 6.1. A two-link manipulator.", + "texts": [ + " Then, for each (eo, so, ~1o, ~3o, 0o) 6 A N x 1~ N \u00d7~2 \u00d7 I~ p, and each to ~ ~ , (1) there exists at least one solution (e(t), s(t), ~1 (t), ~3(t), O(t) ), each solution having an indefinite extension; (2) each solution is bounded and ( e ( t ), s(t)) converges to zero; (3) trim ~i (t), i = 1, 3, exists. Proof: The Lyapunov function employed is 1 1 1 _ ~.)2, V= sTH(q)s+'~(O--o*)TF-I(o--o*)+-~gl (~l--~f)2+~g3(~3 3 where ~ and ~ are as in (3.8), and 0* is as in (P--4). The rest of the proof is similar to the proof of Th. 5.1 and is omitted. Ill 6. Numerical Example 6.1. Manipulator description We apply the continuous control law of Section 4 to a two-link manipulator operating in the vertical plane, shown in Fig. 6.1. There, m i are the masses of each link, m3 is a point-mass at the tip of link 2, Ii are the moments of inertia about the centers of mass located at lc,, and ROBOT ROBUST PATH TRACKING 349 Ji are the moments of inertia of the motors. The numerical values are: m l = m2 = 0 .8Kg, m3 ----- either 0 or 0.47 Kg , 11 = 12 = 0 .4m, lc~ = lc2 = 0 .2m, 11 = Iz = 0.0106Kg m 2 , J1 = J2 = 0.025Kg me. The two links are powered by two identical permanent magnet DC motors located in the manipulator base. Note that both ql and qz are defined with respect to the horizon" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000770_pierb13030101-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000770_pierb13030101-Figure5-1.png", + "caption": "Figure 5. (a) EBG structure in vicinity of microstrip line and, (b) equivalent resonant circuit of (a).", + "texts": [ + " Further, the stop band can also be predicted by the plot of variation of S21 with frequency by the use of a microstrip line with EBG structure which is shown in Figure 4. It is observed that the S21 < \u221210 dB ranging from 4.68 GHz to 6.18 GHz which indicate the stop band behavior of the structure and also in good agreement with the dispersion diagram. F re q u en cy ( G H z) A microstrip line based approach given in [17, 20] is applied to study the resonant frequency characteristics of EBG which is the approach applied to design UWB band notched antennas. As shown in the Figure 5, the complementary L-slot loaded EBG patch is located in vicinity of 50-\u2126 microstrip line with gap \u2018Gp\u2019. An equivalent circuit model is designed based on LC resonator to explain the mechanism of the EBG structure coupled to the microstrip feed line. In this model, C0 denote the capacitance between the EBG patch and microstrip line and C1 is capacitance between EBG patch and ground plane. The inductance L1 is due to current flow through the via and L2 is due to current flow around the L-slots. Then based on the LC-parameters the resonant frequency is given by, fr = 1 2\u03c0 \u221a (L1 + L2)(C1 + C0) (1) To create band notched characteristics for WLAN (IEEE 802", + " When Le1 decreases the center frequency of notched band increases with constant bandwidth. Figure 7(b) presents the variation of VSWR with coupling gap (Gp) between EBG patch and feed line. It is observed that with increase of \u2018Gp\u2019, peak value of VSWR decreases and the bandwidth at notched band increases while upper frequency of notched band kept constant. This is due to the reduction of capacitive coupling between EBG and feedline. The radius (r) of via of EBG structure is also an important parameter which include an inductance parameter in equivalent circuit of Figure 5. The effect of variation of \u2018r\u2019 on VSWR plot is shown in Figure 7(c). It is clear that as the radius of the via decreases, the center frequency of notched band also shifted to lower frequency range with some decrease in bandwidth, it is due to the fact that when the radius of via decreases the inductance related to via increases. Figure 7(e) presents the effect of variation of distance (G2) between two L-shaped slots of EBG patch on VSWR pattern. Due to variation of distance the current path on EBG patch increases which add the more inductance value, due to which the bandwidth and center frequency of notched band increases" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002699_icorr.2015.7281298-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002699_icorr.2015.7281298-Figure7-1.png", + "caption": "Fig. 7. Force model of a powered wheelchair on a slope", + "texts": [ + " By the way, with different virtual mass coefficient M (therefore \u03c4 is changed), the different dynamic corresponds of the system will be obtained. B. External Force Estimation In the above admittance control, the force the attendant applies to the handle of the wheelchair is usually measured by an expensive multi-axis force/torque sensor. To make our PAU affordable for most of users, we do not use such expensive sensor. In stead, we develop a disturbance observer to estimate the applied force of the attendant. As shown in Fig.7, Fxs is the propel force the attendant applies to the handle of the wheelchair; Frs represents the rolling friction between the drive wheel and the ground; Fts is the traction force of the drive wheel of PAU; m is the total mass including the wheelchair, the PAU, and the seat occupant. On a slope with a angle \u03b8 about to the ground, Frs = c \u00b7mg \u00b7 cos \u03b8 (7) where, c is the coefficient of rolling friction. Fts is caused from the motor\u2019s electromagnetic torque and it drives the drive wheel of the PAU by the following equation: Fts = gr r (K\u03c4 I\u2212J d\u03c9m dt \u2212Dm\u03c9m)\u2212mg \u00b7 sin \u03b8 (8) where, K\u03c4 , J , and Dm are the motor\u2019s torque coefficient, motor\u2019s initial moment and damping, respectively. On the other hand, as shown in Fig.7, the resultant force F along the slope is F = Fxs + Fts \u2212 Frs = mr d\u03c9 dt +DW r\u03c9 (9) where, DW is assumed to be the damping of the powered wheelchair. In this way, the applied force Fxs of the attendant is estimated by the following formulation: Fxs = F + Frs \u2212 Fts = mr d\u03c9 dt +DW r\u03c9 + cmg \u00b7 cos \u03b8 \u2212 gr r (K\u03c4 I\u2212J d\u03c9m dt \u2212Dm\u03c9m)+mg \u00b7 sin \u03b8 (10) This equation is a disturbance observer to force-sensorlessly estimate the applied force of the attendant. The whole system model including speed observer, disturbance observer, admittance control, and speed control is shown in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002566_eurocon.2011.5929228-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002566_eurocon.2011.5929228-Figure1-1.png", + "caption": "Fig. 1. (a) Encoded nanowire, (b) Schematic of a NWFET device with high- dielectric layer.", + "texts": [ + " Two types of nanowires are currently produced, called undifferentiated or uniform nanowires, and Differentiated or encoded nanowires. The uniform nanowires are grown identically and are differentiated after assembly. The encoded nanowires, on the other hand, are grown with different encodings in advance. Dopant molecules are added to a gaseous mixture as the nanowires grow. As a result, heavily and lightly doped regions form along the nanowire lengths, depending on the exposure time as shown in Fig. 1(a). These two types of nanowires can be used as active devices in different ways. Consider a mesowire (MW) at the top of a uniform nanowire. Depositing impurities such as gold particles or depositing a highdielectric at the contact between the mesowire and nanowire can lead to a controllable junction, and preventing the deposition of such impurities makes the junction uncontrollable. Applying an electric field on the mesowire can control conductance of the nanowire. A schematic view of such device [13] is shown in Fig. 1(b). In the encoded or differentiated silicon nanowires, the resistance of a lightly doped region of the nanowire can be controlled by an electric field. When a mesowire has a crosspoint with a lightly doped region of a nanowire, the nanowire only conducts when the mesowire does not carry an electric field. In this way, the mesowire forms a fieldeffect transistor (FET) with the nanowire. Controlling NWs with mesowires is a fundamental challenge, which can be satisfied using nanowire decoders. A nanowire decoder consists of a set of MWs and NWs such that applying an electric field on a subset of the MWs causes one or more NWs to conduct" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001436_icmic.2016.7804205-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001436_icmic.2016.7804205-Figure1-1.png", + "caption": "Fig. 1. Geometrical structure of a high Tc superconducting circular disc microstrip antenna.", + "texts": [ + " In the formulation, finite conductivity and finite thickness of the superconducting strip have been accounted for by considering surface impedance, which is modeled according to the Gorter- Casimir two-fluid model and London\u2019s equation [12], [13]. In order to validate our proposed approach, we compare in subsection III.A our numerical results with the theoretical and experimental data of Richard et al. [9]. In subsection III.B, numerical results for the effect of the temperature on the resonant frequency and half-power bandwidth of the high Tc superconducting circular disc microstrip antenna are also given. Finally, concluding remarks are summarized in section IV. II. THEORY The problem to be solved is illustrated in Fig. 1. The high Tc superconducting microstrip patch antenna considered in this work was obtained by depositing a superconducting circular disc of radius a and thickness e on a dielectric layer. The dielectric layer of thickness d is characterized by the free-space permeability \u00b50 and the permittivity \u03b50 \u03b5r.Let J(,) J ( ,) J ( ,)(where T implies transpose) be the surface current density on the circular disc. Also, let E( , , d ) E ( , , d ) E (,, d ) be the value of the transverse electric field at the plane of the circular disc. Owing to the revolution symmetry of the multilayered medium of Fig.1 around the z-axis, when the Helmholtz equations for the longitudinal field components Ez and Hz are solved in cylindrical coordinates inside dielectric layer, it turns out that the dependence of Ez and Hz on the coordinate is of type ein (where n is an integer), as a consequence, J(,) and E( , , d ) can be written as )(),( n n ne JJ i (1) ),(),,( ded n n n EE i (2) Following a mathematical reasoning very similar to that shown in [6], we obtain a relation among J(,) and E( , , d ) in the vector Hankel transform domain given by )(", + " Boundary conditions require that the transverse electric field given in equation (13) vanishes on the area of the superconducting circular disc and the surface current vanishes off the disk, to give the following set of vector dual integral equations: ,)( )( )(),( 0 0jZGHE kkkkdkd nsn o n a (14) akkkdk nnn ,)( )()( 0 0jHJ (15) Now, that we have include the effect of the superconductivity of the circular disc in the integral equation formulation, the well-known Galerkin procedure of the moment method can be easily applied to equation (14) and (15) to obtain the resonant frequencies and the half-power bandwidths of the resonant modes of the high cT superconducting circular disc microstrip antenna. III. NUMERICAL RESULTS AND DISCUSSION The choice of the basis functions is of prime importance as it conditions the stability and convergence of the moment method. The basis functions used in the approximation of the current density on the high Tc superconducting circular disc shown in Fig. 1 are very similar to those used in [6] in the approximation of the current density on the perfectly conducting disc. They are formed by the complete orthogonal set of TM and TE modes of a cylindrical cavity of radius a with magnetic side walls and electric top and bottom walls. These currents modes which are none-zero only on the patch are given by )( i )( )( ' aJ an aJ pnn pn pnn pn \u03a8 , )( )( i- )( ' aJ aJ an qnn qnn qn qn \u03a6 (16) For n 0, 1, 2,... . )(pn\u03a8 ( p 1, 2," + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000533_978-1-4757-3701-1_6-Figure6.4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000533_978-1-4757-3701-1_6-Figure6.4-1.png", + "caption": "Figure 6.4 Compensated SC Integrator", + "texts": [ + "3 illustrates the effect of finite opamp gain to the performance of a SC integrator [MAR 81, GRE 86]. Obviously, due to finite opamp gain, a residual error is imposed at the input terminal of opamp at every integration event of a SC integrator and thus limits the accuracy. This in turns limits the resolution of the L~ modulator. On the other hand, if the output signal can be effectively predicted, a battery that is pre-charged with the amount of gain error can be used to make the charge transferring node a perfect virtual ground as shown in Fig. 6.4. Theoretically, the opamp DC gain can be compensated to make it infinite with a perfect prediction scheme. In practice, a few number of possible compensation techniques have been proposed [KI 91,95, NAG 97] in recent years to significantly reduce the finite-opamp-gain error to achieve good circuit performances. Specifically, for bandpass applications, the double sampling finite-gain-compensation (DSFGC) technique proposed in [NAG 97] can be employed to achieve a compensated opamp gain up to A,2, where Ao is the actual opamp DC gain" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001605_imece2011-63545-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001605_imece2011-63545-Figure7-1.png", + "caption": "Fig 7. Nonlinear FEM (ABAQUS) on rocker section", + "texts": [ + "org/about-asme/terms-of-use 3 Copyright \u00a9 2011 by ASME Next we compared moment capacity about Z axis for four sections from a typical midsize vehicle rocker. Figure 6 shows a comparison between, a) & b) Elastic and plastic load capacity from CATIA using traditional section analysis. c) Effective width method (see figure 3) d) Linear Buckling load from FSM. (FEM and FSM are in good agreement as seen in Table 1, therefore only FSM buckling is shown in Fig 6) e) Plastic buckling of prismatic structure with same section using nonlinear FEM ABAQUS (see Fig 7). The sections were constrained at one end and a pure moment was applied about the Z axis at the other end. f) Actual load in vehicle for SUV side impact from full vehicle CAE, i.e. acting moment or target. Section Number (see Figure 6) Normalized Moment 1 2 3 4 FSM Buckle Moment .897 .97 .986 .981 FEM Buckle moment .927 1.005 1.017 1.012 Difference (%) 3.24 2.79 3.05 3.06 Table 1. Comparison of FSM v/s FEM for linear buckling analysis. FSM is generally softer due to selection of natural deformation modes We observe followings from the comparison, 1) FSM and FEM estimate of buckling load are in agreement (Table 1)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001418_isemc.2018.8394077-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001418_isemc.2018.8394077-Figure1-1.png", + "caption": "Fig. 1. complement Split ring resonator(CSRR) unit", + "texts": [ + " In this paper, we present a CSRR(complementary split-ring resonator) array for electromagnetic energy harvesting and electromagnetic energy received by each unit collect into an output port using a matching circuit. Then we design a rectifier circuit which transform RF to DC and achieve electromagnetic power harvesting on the load. II. METASURFACE ARRAY DESIGN Metamaterial which is an artificial medium, by producing negative permittivity and magnetic permeability, makes the impedance of the medium match with the free space to achieve full absorption of the incident wave. The CSRR designed in this work as array cell to maintain polarization stability is shown in Fig. 1. CSRR are placed on one side of a t = 0.5mm thickness Rogers F4B substrate with a dielectric constant of 2.65 and a loss tangent of 0.001. The geometric dimensions of the cell are optimized to achieve maximum harvesting efficiency at 5.16 GHz, L = 8mm, S = 4mm, w = 0.55mm, and copper thickness is 35 m. For CSRR, the energy conversion efficiency in receive mode is different from the typical radiation efficiency[2].Therefor,the energy conversion define as: incident received P P=\u03b7 Where receivedP is the maximum available time-average RF power received by the CSRR measured at the feed" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000289_bf00772952-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000289_bf00772952-Figure1-1.png", + "caption": "Fig. 1. Test scheme for annular thermomechanical fatigue specimens with parabolic surfaces: 1) specimen ; 2, 3) pressing and rolling rolls; 4) limiter; 5) motor.", + "texts": [ + " They cannot adequately model the material state for many structures, for example those having wedge-shaped zones, different narrowing of the cross section, and also other temperature and stress concentrators. As necessary an increase in thermal stresses is arrived at by an increase in their dimensions. Considerable forces are necessary in order to create mechanical stresses. In annular specimens with mechanical loading by compressing a specimen between rotating rolls according to the schemes presented in [1, 2] and in Fig. 1 an increase in mechanical forces leads to an increase in contact stresses. Failure of the majority of materials under these conditions due to the action of contact stresses occurs sooner than due to the action of bending stresses in the *This work was carried out as a result of finance for the scientific and technical project \"Substantiation of the possibility of using new high-temperature materials, including those with coatings, with the aim of improving the economics for aviation and ship engines,\" (theme 5" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000005_lawp.2007.914115-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000005_lawp.2007.914115-Figure4-1.png", + "caption": "Fig. 4. Simulated and measured plane radiation patterns at 2.7 GHz.", + "texts": [], + "surrounding_texts": [ + "The slot PICA comprises a leaf-shaped metal structure in one plane, and a larger leaf-shaped slot in a second metal plane. A microstrip feed line connects to the metal structure in the first plane. The second plane acts as ground for the microstrip line, cf. Fig. 1. The leaf-shaped slot in the ground plane results in strong electromagnetic coupling to the feeding structure. The antenna impedance can thereby be controlled by adjusting the slot and feed. The area of the feeding structure is approximately a quarter of the slot area. The distance between the bottom edges of the leaf-shaped slot and the feed is crucial for the impedance matching, in particular at the lower end the frequency range and at the highest frequencies. For characterization, an antenna according to the following is built: The leaf-shaped slot and feed line are etched on a 60 mm 60 mm RO3003 substrate with a thickness , a relative permittivity and loss factor . Dimensions are given in Fig. 1. An SMA connector soldered to the 50- microstrip line serves as antenna port. No external matching circuit is included in the presented design." + ] + }, + { + "image_filename": "designv6_24_0000316_mms.2014.7088943-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000316_mms.2014.7088943-Figure1-1.png", + "caption": "Figure 1. The geometry of the proposed RFID Tag antenna", + "texts": [ + " In this work, a new modified T-shaped antenna for RFID-UHF Tag Applications is proposed. To have a good conjugate matching between the tag antenna and the ship, the method of adding four half-circle shaped patches at the both ends of each T monopoles was applied. The measured results show that our new tag antenna design deliver good read range regardless of to which surface the tag is mounted (Metal plate or paper box). Simulation and measurements results are presented, discussed and compared. The proposed tag antenna is illustrated in Figure 1. The structure of this proposed design is very straightforward and has a dimension 80\u00d750 mm2, with two T-monopoles, and four half-circle shaped patches. These patches are connected to each both ends of T-monopoles. Furthermore, the two T-monopoles are linked together by a tag chip. By properly selecting the radius of the four half-circle shaped patches, good conjugate impedance matching can be achieved. According to the geometrical parameters given above, a prototype of the proposed tag antenna has been realized (see Figure 2) using an LPKF Protomat S100 mill/drill" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002026_tap.2015.2477093-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002026_tap.2015.2477093-Figure1-1.png", + "caption": "Fig. 1. Geometry of (a) proposed open-loop square-ring antenna for CP operation and (b) previous open-loop square-ring antenna in [20].", + "texts": [ + " The first antenna is loaded with lumped components, and the second antenna is loaded with distributed elements. The proposed antenna uses an interdigital capacitor and meander-line capacitor and inductor. The details of the proposed antenna design and the CP performance measurement results are presented later. A detailed simulation is performed to understand the behavior of the proposed antenna and optimize it, as described in Section III. The geometry of the proposed square-ring antenna for CP operation is shown in Fig. 1(a). An open-loop square-ring resonator is introduced as a simple structure for producing CP. The proposed antenna is fabricated on a substrate with a thickness h and relative permittivity \u03b5r. The outer and inner side lengths of the square-ring resonator are denoted by L1 and L2, respectively. To maintain the structural symmetry of the antenna, the top, the bottom, and the left- and right-hand sides have the same line width w. The dotted line represents a slot on the opposite side of the substrate", + " The antenna is fed by a microstrip line via proximity (electromagnetic) coupling, and the gap for the feed structure is denoted by g. For a simple design, the width of the feed line wf corresponds to the characteristic impedance of 50 \u03a9. The simulation results show that the dimensions and position of the T-shaped feed structure can be adjusted to enhance the impedance matching at resonance. The simulation is performed using HFSS software. The proposed structure is designed on the basis of the design in [20]. As shown in Fig. 1(b), the proposed open-loop squarering antenna without a lumped element is similar to that in [20], except for the feeding structure. As mentioned in [20], an open-loop square-ring antenna exhibits similar broadside radiation patterns at the three resonant frequencies. The three resonance modes include one unloaded mode (fU ) and two loaded modes (fL1 and fL2). The unloaded mode originates from the fundamental mode (TM11 mode) of a closed-loop square-ring antenna (a square-ring antenna without a gap)", + " It is slightly shifted to a higher frequency compared to the unloaded mode frequency of the closed-loop square-ring antenna. When a gap is introduced within the square ring, two more resonances, namely, the loaded modes, appear at fL1 and fL2 as follows: fL1 < fU < fL2. (1) However, only fU and fL2 are considered in this study because an open-loop square-ring antenna has a low gain (approximately \u221210 dBi) at fL1. Fig. 2 shows the simulated reflection coefficient of the openloop square-ring antenna in Fig. 1. The lumped element is not considered in the simulation. For the simulation, the design parameters of the square-ring slot are L1 = 18mm, L2 = 12mm, w = 3mm, s = 1mm, g = 0.2mm, d1 = 1mm, d2 = 0mm, wt = 0.2mm, wf = 3mm, \u03b5r = 4.4, and h = 1.6mm. As can be seen, there are two resonance frequencies (f1 and f2). The first resonant frequency f1 is the unloaded resonance frequency (fU ), and the second resonant frequency f2 is the loaded resonance frequency (fL2). Another loaded resonant frequency (fL1), which is not shown in this graph, exists at 1", + " To verify the CP mechanism, which requires modes of equal magnitude that are phase orthogonal, the simulated surface electric-field distributions viewed from the top are shown in Fig. 8. In this simulation, a 2.2-pF lumped-chip capacitor is loaded. The simulated frequency is 2.72 GHz, which is the frequency with the minimum AR within the impedance BW. The red region represents the peak electric field, whereas the blue area has almost no electric field. Both plots are normalized to the minimum and maximum values. When the capacitor is loaded, we can see from the electric-field flow directions that the polarization of the designs in Fig. 1 is RHCP. When the proposed antenna is loaded with an inductor at the same location, LHCP is achieved. IV. PARAMETRIC STUDY In order to design a high-performance circularly polarized antenna, a detailed parametric study of the antenna is conducted. Fig. 9 shows the simulated reflection coefficient and AR of the proposed antenna for different widths (w). A 2.2-pF lumped-chip capacitor is loaded into the open gap. The simulation results show that the first and second resonant frequencies decrease as w increases, whereas the matching characteristics and minimum ARs increase" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002224_3527608117-Figure5.19-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002224_3527608117-Figure5.19-1.png", + "caption": "Fig. 5.19 Principle of plasma hot wire cladding (ISAF, TU Clausthal)", + "texts": [ + " Finally, the process is characterized by the possibility of producing defectfree coatings and interfaces with dilution of only 5%. The main application field of this process is coating of components which are subject to high corrosive and/or wear loads. Typical application areas are armouring valves in engine construction, highly stressed zones of tools in the plastic processing industry as well or protective layers on sealing surfaces of fitting parts in the (petro)chemical industry. 5.3.3.8 Plasma Hot Wire Cladding Plasma hot wire cladding (Fig. 5.19) offers a separate adjustment for melting the base material surface by a plasma burner with transferred arc and melting the additive materials by the hot wire technology. The weld additive is usually supplied to the weld pool in the form of two crossing wires. The contact of the wires and the short-circuit over the weld pool causes a current flow of the contact tubes over the free wire ends, and thus a resistance heating to the solidus temperature of the weld additive metal. To minimize the magnetic field influence, alternating current sources are used, which are equipped with voltage limitation to avoid arc discharges (flashover)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000770_pierb13030101-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000770_pierb13030101-Figure8-1.png", + "caption": "Figure 8. Different cases for the variation of number of unit cell of EBG structure.", + "texts": [ + " Due to variation of distance the current path on EBG patch increases which add the more inductance value, due to which the bandwidth and center frequency of notched band increases. The variation of VSWR for different width (Wm) of L-slots on EBG patch is shown in Figure 7(f). It is seen that the upper notched frequency decreases with the increase of slot width, which can be used to tune the band precisely. Further, investigations of the number of unit cell of the proposed EBG structure is carried out on the performance of the antenna. Four different cases are considered as shown in Figure 8. Figure 9(a) shows the variation of VSWR as a function of frequency with the variation of number of elements of EBG structure. In case 1, the value of VSWR and bandwidth of band notched decreases in comparison to case 2 (proposed Antenna 2). This is due to the reduction of capacitive coupling between feed line and EBG structure. When we consider case 3 and case 4, dual band notched characteristics is observed. Therefore, wide notched band is achieved. In case 3, since more unit cells of EBG is directly coupled to feed line, therefore the peak of VSWR value is more than the case 4" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001778_2017-01-1563-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001778_2017-01-1563-Figure5-1.png", + "caption": "Figure 5(b). Block diagram of Circuit", + "texts": [ + " It is evident from this plot that, the absence of air drag destabilises the wobble mode as compared to baseline configuration. Curve H is the result of a reduction in rake angle. It highly destabilises the wobble mode at lower speed range. The aim of the work was to compare numerical analysis with experimental results obtained from test runs that were carried out in a two wheeler. For this purpose, a commercially available motorcycle, equipped with measurement devices was considered for conducting the experiment. Figure 5(a) shows the circuit diagram of the potentiometer that was connected to the Arduino. A memory card was further connected to output port of Arduino to store the data (the code of which is given in Appendix B). The other sensors were connected in the same manner to Arduino. A data logger was fixed to the rear rack, which stored the measured sensor data in a memory chip by the help of Arduino. A metallic toothed wheel made of 14 teeth (wheel speed sensor) was mounted on rear wheel. A Hall Effect sensor was used to find the angular velocity and thereby, the slip produced", + " It was quite a huge challenge to revert back to the same wobble speed, when tests were repeated after some time. Thus the rider was instructed to assure a firm connection with the saddle, by maintain the upright posture andinitiate the steering oscillation by a lateral hand hit on the end of the handlebar, for initialising the disturbance. The main focus of the test runs was to find out the wobble speedand the frequencies of the wobble mode (from the measured steering angle through circular potentiometer) for different speeds. The block diagram of circuit is given in Figure 5(b). Figure 6(a), 6(b) and 6(c) illustrates, an example of a particular test run, with steering angle \u03b4 and roll angle \u03d5 and longitudinal velocity u. The initial speed is about 5.4 m/s, as observed from Figure 6(c). It can be observed from the Figure 6(a), that the roll angle starts continuously varying when rider takes off his hands, starting from 28.2 s upto 30.3s, in order to counter balance the lateral thrust. An increasing wobble oscillation shows up after the initial lateral hit on the handlebar" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002153_amm.813-814.915-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002153_amm.813-814.915-Figure1-1.png", + "caption": "Figure 1:Two wheeler wheel designs", + "texts": [ + " (#72606402, University of Newcastle, Callaghan, Australia-28/12/16,09:39:58) [9] performed the static and modal analysis on four wheeler rim with aluminum and steel and suggested steel is better for manufacturing. To observe the best design and material for automobile two wheeler wheels, three designs and three materials are chosen which are existing in present days. Both static structural and free vibrational analyses are performed on the three wheel designs each containing 3 and 5 spokes made up of steel, aluminum and magnesium metals as shown in Fig 1. Comparison is made between these designs and concluded the best design which contains lower von mises stress, higher specific structural stiffness, and higher structural stiffness. Problem Modeling. The designs are modeled in Ansys workbench. The dimension of the wheel except spokes is taken from the reference [2]. Material Properties. The materials generally used for manufacturing of wheels of vehicle are Table1: Material properties Material Density (Kg/m^3) Young\u2019s Modulus E (Pa) Poisson\u2019s Ratio (\u028b) Shear modulus G (Pa) Steel 7850 2e11 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001770_tim.2014.2364699-Figure10-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001770_tim.2014.2364699-Figure10-1.png", + "caption": "Fig. 10. Diagram of orthogonal coordinate transformation.", + "texts": [ + " For the rotating angle \u03b4 can be arbitrary, if \u03b4 = 0 is chosen, (6) can be represented by \u23a7\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23a8 \u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23a9 La0 = k\u2211 n=1 an [ cos(n\u03c9t) ( 1 \u2212 cos 2n\u03c0 3 )] Lb0 = k\u2211 n=1 an [ cos n ( \u03c9t\u2212 \u03c0 3 ) \u2212cos n ( \u03c9t+ \u03c0 3 ) cos (n\u03c0 3 )] Lc0 = k\u2211 n=1 an [ cos n ( \u03c9t + 2\u03c0 3 ) \u2212 cos n ( \u03c9t \u2212 \u03c0 3 ) cos (n\u03c0 3 )] (21) and (2) can be represented by \u23a7\u23aa\u23aa\u23aa\u23aa\u23aa\u23a8 \u23aa\u23aa\u23aa\u23aa\u23aa\u23a9 La0 = La \u2212 1 2 (Lb + Lc) Lb0 = Lb \u2212 1 2 (La + Lc) Lc0 = Lc \u2212 1 2 (La + Lb). (22) The corresponding normalized vectors can be obtained as \u23a7\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23a8 \u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23a9 L\u0304a0 = La0\u221a 2 3 [ L2 a0 + L2 b0 + L2 c0 ] L\u0304b0 = Lb0\u221a 2 3 [ L2 a0 + L2 b0 + L2 c0 ] L\u0304c0 = Lc0\u221a 2 3 [ L2 a0 + L2 b0 + L2 c0 ] . (23) As shown in Fig. 10, the vector L\u0304b0 (\u03b4 = 0) is perpendicular to L\u0304a\u03b4 (\u03b4 = \u03c0/6). Furthermore, combining (9) and (23), it can be found that (L\u0304a\u03b4) 2 + (L\u0304b0) 2 = 1. (24) Substituting (13) into (24), the L\u0304b0 can be obtained L\u0304b0 = \u221a 1 \u2212 cos2 ( \u03c9t \u2212 7\u03c0 6 ) = sin ( \u03c9t \u2212 7\u03c0 6 ) . (25) Thus, L\u0304b0 is also an unit sine function. As shown in Fig. 10, the angle \u03d1 can be estimated based on the geometrical relationship \u03d1 = arctan L\u0304b0 L\u0304a\u03b4 = \u03c9t \u2212 7\u03c0 6 . (26) Thus, the rotor position can also be estimated \u03b8 = 22.5 \u03c0 [ 7\u03c0 6 + arctan L\u0304b0 L\u0304a\u03b4 ] . (27) As discussed in the previous sections, the proposed coordinate transformation-based method is dependent on the unsaturated full-cycle inductance. Under high-load drive running conditions, the method is not applicable. Thus, to solve this problem, the phase current slope difference comparisonbased method is adopted for high-load driving cases" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003069_pcicon.2016.7589207-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003069_pcicon.2016.7589207-Figure6-1.png", + "caption": "Fig. 6 (b) Flexible", + "texts": [ + " The outer surface is covered with silicone rubber. Heat can be directly where it is needed to improve heat transfer, speed up warming and decrease wattage requirements, which are all essential features of a flexible-type heating element. Fig. 4 (c) Tubular Fig. 4 (d) Finned tubular heater Fig. 4 (e) Flanged mount tubular heater Fig. 5 (c) Strip heater - connection Fiberglass-reinforced silicone rubber provides dimensional stability without sacrificing flexibility. Because insulating layers are very thin, heat transfer is rapid and efficient. Fig. 6(a) shows the construction with a wire-wound element or with an etched foil element. Thin construction of the heater allows it to fit into applications where space is limited. Fig. 6(b) shows the finished product. 4) Explosion Proof: Explosion-proof heaters are suitable for hazardous area such as Class 1 Division 1 or 2. The surface temperature of the heater dictates the temperature code for the hazardous area. In this type of heater, the terminals are enclosed in a NEMA type 7 or 9 enclosure with a threaded entry for the end users to connect. The conduit, fittings and wirings have to be suitable for the area requirement and approved by the relevant authorities. The construction has to be suitable and strong for the environment and include over-temperature protection" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003746_rem49740.2020.9313869-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003746_rem49740.2020.9313869-Figure7-1.png", + "caption": "Fig. 7: Hardware for controller implementation and validation.", + "texts": [ + " In this case a well damped system behavior is reached by setting TN as a multiple of Tmech ' Therefore, TN is selected to 80 ms. V. VALIDATION For the validation of the discussed controller the hybrid actuator test bench in Fig. 6 is used. It consists of the hybrid actuator, an iron yoke, an air gap sensor and an iron weight of 6 kg. The actuator can oscillate freely within an adjustable range, set to 0.5 mm - 1.5 mm during operation. The control algorithm is implemeted on a ST N ucleo F767ZI microcontroller (MCV) (Fig. 7a). For air gap and current measurement a shield board (Fig. 7b) is de signed. The shield is mounted on the MCV and equipped with a 16-channel 16-Bit dual simultaneous sampling ADC (Analog Devices AD7616) exchanging data with the MCV via SPI. Apart from that, the power supply and PWM outputs as well as measurement inputs for six H-bridges are part of the shield board. The H-bridge in Fig. 7c is equipped with two half bridges of type International Rectifier IRSM808 204MH and a LEM LAX 100 - NP for current mea surement. Additionally, electronics for air gap measurement and power supply of the air gap amplifier are part of the board. Air gap measurement is realized with an eddy cur rent sensor and amplifier of type Emerson P R6423 and Emerson GON021. Simulation is performed numerically in Simulink. For this purpose a model is implemented following the structure in Fig. 5. The hybrid actuator is modeled according to the mechanical differential equation (19)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001095_20.280999-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001095_20.280999-Figure1-1.png", + "caption": "Fig. 1 3D finite element model of tooth profile.", + "texts": [ + " INTRODUCTION It is known that some step motors generate unacceptable audible noise levels during operation. Recent investigations have revealed that the origin of acoustic noise in step motors is the radial vibration on the stator caused by the variation of radial magnetic force inside the air-gap acting on the stator due to rotor movement [1][2][3]. In this paper, we study the effect of tooth shapes, rotor tooth to valley ratio and saturation on radial forces in step motors based on a typical step motor tooth profile using 3D finite element method. 11. 3D FINITE ELEMENT MODEL Fig. 1 shows the 3D meshed model of the tooth profile used for radial force calculation. For the computation, it is assumed that the stator and rotor bases are at constant potentials with a positive potential assigned to the stator and a negative potential assigned to the rotor [4]. The value of the potential was chosen to give a flux density inside the air-gap to be equal to a typical value for a stepping motor running at high stepping rate. It is because for step motor the problem of acoustic noise is most pronounced at high stepping rate", + " By integrating the stresses along a closed path, C, surrounding a component, the total force on the component is computed. The Maxwell stress equations are expressed as 181: where f is force per unit area, B is the flux density, n denotes the component normal to C, t the component tangential to C and The radial magnetic forces inside the air-gap acting on the stator are calculated by integrating the stresses over a closed area inside the air-gap through the central layer of air-gap elements as shown in Fig. 1 in order to get the most accurate results [6][7][8]. The mesh as shown in Fig. 1 is selected for the force calculation because any further increase in the number of element results in only a 5 % change in the radial force value. The radial forces are calculated for full alignment, 3/4 alignment, 1/2 alignment, 1/4 alignment and is the permeability of free space. 0018-9464/93$03.00 1993 IEEE 2414 mis-alignment positions between stator and rotor teeth for various tooth shapes, tooth to valley ratios and potentials. The d t of radial force density distribution for full alignment between stator and rotor teeth with rectangular tooth shape and stator and rotor tooth to valley ratio (TVR) of 1:l is shown in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000265_icra.2018.8463165-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000265_icra.2018.8463165-Figure4-1.png", + "caption": "Fig. 4. Reconstructed 3D maps consisting of line segments: (a) results of Line-SLAM [3]; (b) results of our Struct-PL-SLAM (color red, green and blue represent 3 dominant directions).", + "texts": [ + " Around the fence, 800 cameras are generated along a circle which is the intersection of the sphere \u03a6 and a horizontal plane passing through the centroid of interval \u03a9. We synthesize an image sequence by projecting 3D features to cameras. We compare our Struct-PL-SLAM with the non-structural line-based system [3] denoted as Line-SLAM. As shown in Fig. 3, the error accumulation in absolute pose of LineSLAM is significant over time. In contrast, the error of our Struct-PL-SLAM remains much lower, demonstrating the advantages of structural features. Fig. 4 shows a comparison of the reconstructed 3D line segments. Line-SLAM reconstructs a disordered map, while our Struct-PL-SLAM can generate a more accurate and structured map, thanks to precise camera pose and 3D map optimization strategy leveraging the structural constraints. We evaluate the proposed system on the HRBB4 dataset [26]. The image sequence is recorded in a typical corridor scene by a monocular camera mounted on a moving robot and contains 12,000 frames of 640\u00d7 320 pixels. The 1http://ceres-solver" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001248_tmech.2012.2215048-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001248_tmech.2012.2215048-Figure3-1.png", + "caption": "Fig. 3. Lateral compliance variation of 110.0-mm postbuckled strips at the point of maximum deflection: (a) experimental settings and (b) compliance variation graph. The error bars were determined at 95% confidence taken from a sample set of 5 for each interval (t = 2.571).", + "texts": [ + "0, and 3.6 mm and a frequency range of 0.1\u20134.0 Hz. This frequency range has been extended from that related to the cardiac application (heart rate) to understand the wider frequency response of the system. For all experiments, a strip of 110.0-mm length was used. The system output was defined as the lateral displacement of the strip at ymax quantified using a LVDT. Initial measurements established that the dynamic effect of the LVDT on the actuation response was insignificant. An instrument [see Fig. 3(a)] was devised to quantify the strips static compliance. This instrument measured the force needed to laterally displace the actuator 2.54 mm at the position of ymax when the linear actuator was still. A single strip of 110.0 mm was used throughout and the boundary attachment method and linear displacement varied. Compliance was calculated as displacement/normal force. Fig. 2 shows images of the strips for different strip neutral lengths and boundary conditions. Also shown on the figure are the extracted shapes using the visual profilometry method described earlier", + " There is some variation with input amplitude between different attachment methods, with the addition of fixed end attachments reducing the excitation amplitude. When both ends of the strips are free to rotate, there is a shift in magnitude response of approximately 0.2 dB with change in input amplitude. When both ends are fixed, this shift is reduced and there is almost no change in the frequency response with input amplitude. This indicates that the fixed supports stiffen the actuation structure. Over this frequency range, there is no indication of resonance or other undesirable dynamic behavior. Fig. 3(b) shows the change in compliance of the actuator with variation in end displacement and attachment condition. The most compliant system was found when both ends of the actuator were free to rotate. With the addition of one clamped end and then two clamped ends, the compliance reduced progressively. Going from both ends free to both ends clamped produced around a 2.5 times decrease in compliance at any one separation distance indicating that clamping the ends of the actuator is a viable means of altering actuator compliance", + "46 mm) showed a cubic growth of stiffness for all three boundary attachments. Modifying the material used for the strips, does not change the buckled shape of the strips and only changes the stiffness (Table I). This means that a modified compliance can be implemented without having to change shape design or position control. For the cardiac simulator application, the measured compliance on an explanted ovine heart using the same apparatus as above gave a compliance of 3\u00d710\u22123 m/N. This was within the range measured for the metal strips shown Fig. 3(b). For an active heart, the compliance would decrease from this level as the contraction of the heart muscle decreases compliance. Tactile heart stiffness approximately doubles between contraction and relaxation [25], and this stays within the variation found here. Further tuning of compliance to other tissues could be obtained by altering material properties and geometry at build time. It is important, however, that the \u201cnormal\u201d compliance is matched, as most published compliance values tend to represent Young\u2019s modulus type stiffness and this material property is not the same as the response of the material to a normal load" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002960_iembs.2009.5332882-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002960_iembs.2009.5332882-Figure8-1.png", + "caption": "Fig. 8. Axis-configuration of the Foot-Boards.", + "texts": [ + "5816 \ud835\udc60\ud835\udc53 (5) The angle and vertical elevation of the foot boards over time are two must-to-know information for developing the motion algorithm of foot-boards. From the raw coordinate data provided by [7] we have determined this information. A. Foot-Boards Movement Trajectory For a sagittal plane, from the gait study of Section II, it can be concluded that each foot trajectory can be denoted by a vector \ud835\udc47\ud835\udc53 = [\ud835\udc4b\ud835\udc53 \ud835\udc61 ,\ud835\udc4d\ud835\udc53 \ud835\udc61 ,\ud835\udf03\ud835\udc53 \ud835\udc61 ] where (\ud835\udc4b\ud835\udc53 \ud835\udc61 ,\ud835\udc4d\ud835\udc53 \ud835\udc61 ) is the coordinate of the ankle position and \ud835\udf03\ud835\udc53 \ud835\udc61 denotes the angle of the foot. For simplicity in robotic movement design as shown in Fig. 8 we have adopted the trajectory vector as \ud835\udc47\ud835\udc39 = [\ud835\udc4b\ud835\udc39 \ud835\udc61 ,\ud835\udc4d1\ud835\udc39 \ud835\udc61 ,\ud835\udc4d2\ud835\udc39 \ud835\udc61 ] which supports all the components of Tf. Now from experimental data of [8], at heel strike or initial contact, the feet are in 250 dorsiflexion with the toes up, followed by a total contact with the ground at the end of the loading response- the toes drop towards neutral alignment and maintain this position throughout mid stance. With heel rise in terminal stance, the foot dorsiflexes up to 200. This motion continually increases throughout pre-swing to a final position of 700 extension", + " The following table summarizes the time division of horizontal movement of the foot-boards All the vertical movements of both the forwarding and backwarding footboards have to be completed within the time tstep. Now following the %duration of different phases of a complete walking cycle suggested in [8], time division for movements along two vertical axes is shown in Tables II and III: From the above information we can easily derive the vertical velocity along Z1 and Z2 direction over time as shown in Fig. 11 (Z1 and Z2 were described in Fig. 8). On the basis of the time divisions for horizontal and vertical movements along X, Z1 and Z2 axes shown in Table I, II and III, we have simulated our proposed person specific gait algorithm in MATLAB. Fig. 12 depicts motion of the foot-boards during backward and forward movements. In this figure we have also shown the co-ordinate points for ankle, knee, hip and base rib joints derived from sample data of a normal subject, as according to our first hypothesis we are expecting that the foot-boards movement will also help the patients to follow an ideal trajectory and angle for these joints" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002482_9783527631506-Figure8.13-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002482_9783527631506-Figure8.13-1.png", + "caption": "Figure 8.13 (a) An infinite-gain OA and (b) its transpose, a COA of infinite gain.", + "texts": [], + "surrounding_texts": [ + "We now illustrate the method of obtaining CM circuits from OA-based VM circuits. We first take the case of a filter that employs a finite gain VA (which can be realized using an OA), and then two filters that employ infinite-gain OAs. 8.5.1.1 CM Biquads Derived from VM Biquads Employing Finite Gain Amplifiers Consider the LP Sallen and Key VM structure of Figure 5.4, shown as Figure 8.14a for convenience. We know from Eq. (5.13) that its VTF is given by Vo Vi = K(G1G3) C2C4 s2 + s {( G1 C2 ) + ( G3 C2 ) + (1 \u2212 K) ( G3 C4 )} + G1G3 C2C4 (8.22) One can easily obtain the transpose by replacing the OA of gain K by a COA of gain K with its ports reversed and leaving the other elements intact. The resulting" + ] + }, + { + "image_filename": "designv6_24_0001472_9781118354179.auto089-Figure11-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001472_9781118354179.auto089-Figure11-1.png", + "caption": "Figure 11. Twin clutch for an agricultural vehicle. (Janic\u0301ijevic\u0301, Jankovic\u0301, and Todorovic\u0301, 1987. Reproduced by permission of Faculty of Mechanical Engineering, University of Belgrade.)", + "texts": [ + " As a result, dry twin clutch designs for twin clutch gearboxes (as shown in Figure 10) are only suitable for smaller vehicles and engines, typically under 350 Nm maximum engine torque. Clutches for engines with higher torque require wet clutches to satisfy increased heat dissipation requirements and ensure acceptable life of friction pair components. In agricultural and \u201cimplement carrier\u201d vehicles clutches have an additional function in separating the power flow from the engine flywheel toward two outputs shafts. Indicated by arrows in Figure 11 (Janic\u0301ijevic\u0301, Jankovic\u0301, Encyclopedia of Automotive Engineering, Online \u00a9 2014 John Wiley & Sons, Ltd. This article is \u00a9 2014 John Wiley & Sons, Ltd. DOI: 10.1002/9781118354179.auto089 Also published in the Encyclopedia of Automotive Engineering (print edition) ISBN: 978-0-470-97402-5 and Todorovic\u0301, 1987), the power entering the clutch is split between one (central, solid) shaft, which is entering the gearbox, with the other (outside, hollow) shaft driving power take off (PTO). Such a design enables independent control of the vehicle and machine (implement) propulsion. The normal force (Figure 11) is provided by circumferentially positioned coil springs (only one is shown), installed between two central pressure plates. The pressure plate on the left is pressing one friction disc (on the left), splined to the central shaft, against the engine flywheel. The pressure plate on the right pushes the second friction disc (on the right), splined to the hollow shaft, against the (heavy) clutch cover. Typically, there is a single clutch pedal, with the control levers and mechanism designed in such a manner that there is, as the pedal travel progresses, sequential power interruption (and establishment) to the two output shafts" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003824_012049-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003824_012049-Figure1-1.png", + "caption": "Figure 1. CAD\u2013drawing of friction type roller screw mechanism: 1 - screw, 2 - housing, 3 - roller, 4 - nut, 5 - annular gear, 6 - bearing base.", + "texts": [ + "ontent from this work may be used under the terms of the Creative Commons Attribution 3.0 licence. Any further distribution of this work must maintain attribution to the author(s) and the title of the work, journal citation and DOI. Published under licence by IOP Publishing Ltd During the computer-aided design of a roller screw mechanism it is necessary to predict its functional characteristics [1]. In some cases, especially for friction roller screw mechanism [2] (figure 1), a preliminary estimation of their efficiency is necessary if they are made of different materials, with different coatings and lubricants. In this paper, the influence degree of the material, the type of lubricant and the type of coating on the coefficient of friction and efficiency has been theoretically established. 6th AMMSE 2019 IOP Conf. Series: Materials Science and Engineering 739 (2020) 012049 IOP Publishing doi:10.1088/1757-899X/739/1/012049 The friction coefficients for uncoated surfaces has been determined from the reference data" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001469_s12541-013-0180-1-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001469_s12541-013-0180-1-Figure1-1.png", + "caption": "Fig. 1 Physical model of a 3D overhead crane", + "texts": [ + " This paper is organized with five sections. The 3D overhead crane dynamics is constituted in section 2. Adaptive robust control design is represented in section 3 composed of proposing an adaptive SMC law, constructing an adaptation parameter mechanism, and analyzing the system stability. Section 4 implements both numerical simulation and experiment study with real-time crane system. Finally, some conclusions and suggestions are mentioned in section 5. The crane system is physically modeled as on Fig. 1. The bridge is a distributed beam whose mass mb is focused on the bridge center. ml denotes the equivalent mass of hoist mechanism, mt and mc are masses of trolley and cargo, respectively. The overhead crane has five degrees of freedom corresponding to five generalized coordinates: x(t) is trolley displacement, z(t) indicates bridge movement along Oz axis, and the payload position is described by three generalized coordinates (l, \u03b8, \u03d5). Hence, the positions of system masses are characterized by q = [z x l \u03d5 \u03b8]T" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000621_asemd.2009.5306701-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000621_asemd.2009.5306701-Figure1-1.png", + "caption": "Figure 1. Sand-pile model of a HTS bulk magnet.", + "texts": [ + " The control strategy based on voltage Space Vector PWM (SVPWM) is applied to the running of the HTS-LSM, and the SVPWM speed-adjusted system simulation model based on Matlab/Simulink is also built. This work will give a great help for the further researching of the technologies of HTS-LSM. II. NUMERICAL MODEL AND EQUIVALENT MAGNETIC CIRCUIT The flux density distribution produced by a HTS bulk magnet can be determined numerically by the sandpile model in combination with the Biot-Savart law [9-11]. The sandpile model of the cylindrical HTS bulk magnet is schematically shown in Fig. 1. The cross section of each current loop is given by the height h and width w. The complete volume of the HTS bulk magnet is divided into NL layers with NC current loops in each layer. The current I in one current loop is given by cI J h w= \u0394 \u0394 (1) where Jc is the critical current density of HTS bulk magnet, and it is estimated as ( )trap 0 1 z c B r J r\u03bc \u2202 = \u2212 \u2202 (2) where 0 is the permeability of vacuum, r the radial distance from the center of the bulk magnet, and trap zB the z component of the trapped-magnetic field" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003611_is3c.2014.314-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003611_is3c.2014.314-Figure3-1.png", + "caption": "Fig. 3 Distribution of the PM poles array on the rotor", + "texts": [ + " )1( From the Helmholtz\u2019s Theorem, there must be a magnetic vector potential field A that can satisfy the equation of 0)( A . Hence, the magnetic flux intensity B can be expressed as AB . )2( Because the actuator has two layers of PM poles distributed in 3D space, the magnetic circuit of the motor is complicated. Without loss of generosity, the magnetic field modeling is conducted on a single pole first, and then the superposition principle is used for all poles in the spherical coordinate system to obtain the complete magnetic flux density distribution in the actuator. 2) Layout and parameterization of PM poles Fig.3 illustrates the distribution of the two layers of PM poles on the inner and outer rotors. To reduce the system cost, cylindrical PM poles are employed for the actuator design. The parameters of a single pole are illustrated in Fig.4, where M is the magnetic intensity vector of the magnetic medium, r and 0h are the radius and the height of the PM poles, respectively, 1h is the distance from the center of the motor to the lower surface of the PM poles. For any magnetic medium in the steady magnetic field, there are equivalent volume magnetization current and equivalent surface magnetization current, i" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003809_0921-5093(94)09732-1-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003809_0921-5093(94)09732-1-Figure1-1.png", + "caption": "Fig. 1. Schematic presentation with maximum elastic shear distortion along the z axis.", + "texts": [ + " Though these findings constitute a contradictory fracture mode, the difference is phenomenological rather than substantial as the background of the deformation mode is the same and originated from the shear stress field developed around the crack tip. An indication of this difference is given in Fig. 13 and it is shown schematically in Fig. 14. The stress developed and the surface deformation observed exhibit a pattern of plastically deformed material with little or no evidence of shearing. This is because the observation was made along the Z-direction which is also the direction of deformation. Fig. 1 I. Higher magnification of the area indicated by the series of arrows in Fig. 10. Cracks associated with regions of high deformation are shown. The conclusions drawn are: (1) The theoretically determined plastic zone range depends only on the pre-crack e and has value 0.6~. (2) The experimentally measured plastic deformation falls in the range 0.2~ 0.6~, which is in very good agreement with the theoretical predictions. (3) The plastic zone determined from stress intensity factor (K m) calculations established results fitting well to both the theoretical and experimental determinations" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003929_rast.2013.6581342-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003929_rast.2013.6581342-Figure7-1.png", + "caption": "Fig. 7. Satellite Quetzal V.l", + "texts": [ + " The satellite bus consists of the following subsystems, which will be developed under requirements and constraints of the selected payload: navigation, attitude determination and control subsystem (NADCS), power supply subsystem (PSS), structural subsystem (SS), thermal control subsystems (TCS), telecommunications subsystem (TS), data handling and processing subsystem (DHPS), telemetry subsystem (TMS) and propulsion subsystem (PS). For the correct integration stage, it has been considered the interface subsystem. Also will be performed the reliability and electromagnetic compatibility analysis, and as part of our recent research field, the sustainability analysis. Preliminary work on the Quetzal satellite is shown in the following figures (Fig. 5, Fig. 6 and Fig. 7). The first proposals are currently used for mechanical, thermal and vibration analysis, which will be verified later on mock ups, while other systems are developed in a engineering model stage with COTS and low cost elements by the student groups at UNAM. Once the electronic design and software has the first round of test, we will evaluate if groups from other institutions will be invited for the integration of the full engineering model, and definition would be made for the flight model based on reliability, cost and execution" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001924_sp.j.2096-5796.2019.0016-Figure9-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001924_sp.j.2096-5796.2019.0016-Figure9-1.png", + "caption": "Figure 9 A large-range six-axis F/T sensor with dumbbell grooves[31].", + "texts": [ + " The above mentioned mechanical structures are classical structures of elastic beams. In recent years, many researchers presented some novel mechanical structures of elastic beam on the basis of the classical structures to improve the performance of multi-dimensional force sensor. A large-range six-axis force sensor made by machining dumbbell grooves in cross beams is developed by Changchun Institute of Optics, which has high sensitivity while ensuring a large measurement range[31]. Schematic diagram of the elastic beams of the sensor is depicted in Figure 9. Mastinu et al. designed a six-axis F /T sensor with a novel mechanical structure, seen in Figure 10a. The sensing element of this sensor is a quasi-statically determined three spoke structure constrained by virtue of elastic sliding spherical joints, seen in Figure 10b, which is designed to avoid friction[32]. Experimental results demonstrated that the sensor have a good performance in linearity, crosstalk and dynamic behavior. Liang et al. developed a novel miniature fourdimensional force sensor with elastic elements consisted of circular diaphragm and cantilever beam, seen in Figure 11[33]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003231_tap.1980.1142339-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003231_tap.1980.1142339-Figure3-1.png", + "caption": "Fig. 3. Geometry of arbitrarily shaped wire antenna located over lossy half-space.", + "texts": [ + " r2' is chosen to be large enough so that RCM expressions are valid at 0'. Therefore, vector potential values along interval 0'0 are obtained by using initial value at 0' and integrating down along z axis. 111. WIRE ANTENNAS RADIATING OVER A LOSSY HALF-SPACE The approximate field solution to the current element problem radiating over a lossy half-space, developed in the previous section, in conjunction with the method of moments [ 11 can be employed to analyze a wide variety of thin-wire antenna problems radiating over a lossy half-space. Fig. 3 depicts the geometry of an arbitrarily shaped wire antenna over a lossy half-space, with (rQ, e,, GQ) defining a point on the antenna axis. For simplicity it is assumed that the antenna is entirely in the xz plane (GQ = 0). Assuming that the antenna is excited by 400 IEEE TRANSACTIONS ON ANTENNAS AND PROPAGATION, VOL AP-28, N0 .3 , MAY 1980 the field pxc&), and having a loading function A(%) n/m, one can write a general integral equation enforcing the total E-field along the antenna equal to that induced by the possible loading function, that is, I(;,) ", + " Here pulse-basis and delta-matching functions are chosen since they eliminate the need for integrating the kernels G', and c'h. The number of unknown patches on the antenna N should be large enough so that the patch-length A is at most one-sixth of the wavelength. The approximated current along the antenna is therefore represented as N f = IJn, (25 ) n = 1 for which I , is an unknown constant value over the nth patch and zero outside of it; also in is a known unit vector tangent to the antenna at the center of the nth patch (see Fig. 3). Substituting (25) into (1 7) and letting subscripts n = I , 2 , 3 , .-. denote \"valuation at the center of the nth patch,\" and letting [I] and [Eexc] be column vectors containing the current and the tangential excitation field values at successive patches, one finally arrives at [Eexc] = - [Zimp][I] + [A][I], (26) where [ A ] is a diagonal matrix with elements A,, A2, -, An ; and [Z1mp] is an n x n square matrix with its ith row and jth column element def ied as ZijimP = Af(7,i) . [i . f@,,)6u@,i, TUj) +; " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000885_j.apm.2014.04.032-Figure9-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000885_j.apm.2014.04.032-Figure9-1.png", + "caption": "Fig. 9. The angular capability angles for the ball joints M (a), N (b) and C (c).", + "texts": [ + "032 in the local reference frame x1y1z1 of the lower suspension arm; Please degree rN \u00bc \\ ns3 ; uN \u00bc \\ \u00f0 ns3 \u00de3; \u00f0 uN\u00de3 \u00bc arccos nx3 s3 ux3 N \u00fe ny3 s3 uy3 N \u00fe nz3 s3 uz3 N ; \u00f017\u00de in the local reference frame x3y3z3 of the upper suspension arm; rC \u00bc \\ ns4 ; uC \u00bc \\ \u00f0 ns4 \u00de4; \u00f0 uC\u00de4 \u00bc arccos nx4 s4 ux4 C \u00fe ny4 s4 uy4 C \u00fe nz4 s4 uz4 C \u00f018\u00de in the local reference frame x4y4z4 of the tie-rod. The angles s are expressed in the reference frames of the casings planes, respectively in \u00f0xs1 ys1 \u00de for the joint casing M, \u00f0xs3 ys3 \u00de \u2013 for N, and \u00f0xs4 ys4 \u00de \u2013 for C (Fig. 9), having the form: sM \u00bc arccos u ys1 Mffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi u xs1 M 2 \u00fe u ys1 M 2 r sgn \u00f0 xs1 \u00de; \u00f019\u00de sN \u00bc arccos u ys3 Nffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi u xs3 N 2 \u00fe u ys3 N 2 r sgn xs3 \u00f020\u00de sC \u00bc arccos u ys4 Cffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi u xs4 C 2 \u00fe u ys4 C 2 r sgn \u00f0 xs4 \u00de: \u00f021\u00de The computation of the angular capability angles, including the socket size, involves to know the current positions of the kinematic elements" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003381_tia.2004.827476-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003381_tia.2004.827476-Figure4-1.png", + "caption": "Fig. 4. Circuit of three-phase IPM synchronous motor.", + "texts": [ + " The equivalent resistance for rotor bars including the rotor end rings can be given below if the bars are distributed at equal intervals in the rotor [5] (4) where is the resistance of a bar, is the resistance of the end rings, is the number of rotor slots, and is the pole pair number. Therefore, is given by (5) This coefficient is found effective to take into account the rotor-bar current for the fundamental space harmonic. Moreover, the agreement between computed and measured results of the load performance characteristics at synchronous speed is good as described later in Section IV. Therefore, it is considered that the design use of the is acceptable, even if the higher space harmonics exist [6]. The value of the coefficient is 0.55 in this paper. Fig. 4 shows the circuit of the experimental motor. It has three stator phase windings, which are star connected with a neutral. The voltage and current equations of the IPM synchronous motor are given as (6) (7) (8) (9) where , and are the phase voltages, subscripts , and represent stator quantities in lines , and , respectively. is the potential of the neutral , when the potential of the neutral of the supply source is zero, , and are the line currents, and and are the resistance and end-winding leakage inductance of the stator winding per phase, respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002224_3527608117-Figure11.9-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002224_3527608117-Figure11.9-1.png", + "caption": "Fig. 11.9 Spraying chamber (schematically shown).", + "texts": [], + "surrounding_texts": [ + "For large volume production it is recommended to use specially designed spray forming units for a specific alloy. The disadvantage of this is that if there was an unexpected alloy change at the plant this would lead to an extensive service processing time. The manufacturing plant in St. Avold, France consists of a 1.2 t inductive melting furnace, a 2.5 t holding furnace with a flanged fore-hearth and the actual spraying chamber. The melting furnace is filled from a container by a tilting apparatus. From there the molten metal is transferred via a launder into the holding furnace. The complete unit is situated on load cells to record and control continuously the data related to the metal. By controlled pressure increase in the holding furnace the molten metal is forced into the fore-hearth. The PEAK Werkstoff GmbH runs two serial spray forming units and one F & E spray forming unit. For spray forming the holding furnace is moved in the direction of the spray chamber and the fore-hearth is docked onto the spray chamber. The pressure of the fore-hearth on the spray chamber can be adjusted by load cells in the hydraulic cylinders, therefore absolute impermeability can be guaranteed and the sealing elements are not mechanically destroyed. The most important feature of the holding furnace is the pressure control. Using that, it is possible to keep the bath level constant during spraying, independent of the furnace filling level. A constant bath level is a mandatory condition to control the rate of metal flow accurately. This happens by a continuous increase in the overpressure within the holding furnace at a speed which corresponds exactly with the rate of flow. The measured bath level, taken by a float lever in the fore-hearth, is used as the actual value for control. At the end of the spraying process the pres- sure is reduced only as far as to ensure that the applied foam filter in the forehearth is just still wetted by the molten metal. This is important for a long service life of the filter. Furthermore, the cleaning effort of the fore-hearth is thereby considerably reduced. 11.3.4" + ] + }, + { + "image_filename": "designv6_24_0001255_3297097.3297110-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001255_3297097.3297110-Figure1-1.png", + "caption": "Figure 1. Free body diagram of the foot and leg for knee", + "texts": [ + " ),( 2 1 mrII iCoM n i a (2) where p is the number of objects rotating around a axis, aI is the moment of inertia around a axis, and jr the distancebetween axes a and CoM of object j . In order to define a kinematic model for the leg, it was necessary to develop a free-body diagram of each section. In this paper only the knee and ankle were considered and the Drillis and Contini human body representation was used[12], [13]. That representation parametrizes and expresses each body segment length as a fraction of total body height as its shown in Figure 1. rmwFN 2 )()sin()( 2211 2 21 LmLmwmmgAN (3) marFT )()cos()( 221121 LmLmaAmmg T (4) actuation In Figure 1, the lengths 1L , 2L and 3L , are the distances between the CoM of the body segment and the knee engines axis of rotation. Masses 1m and 2m are the leg mass and foot mass respectively. TA and NA are the tangential and normal components of the force reaction at the motor shaft and MT the motors torque. Equations (3)\u2013(5) are the movement equations applicable to the leg and foot actuated on the knee [14]. Taking into account equation (5) and Figure 2, we can obtain the relationship between the torque and speed of the system as: ,)()sin()cos()( 2 2 1 322211 a N NILgmLmLmgMM TM (6) where a is the acceleration in motor shaft, M is friction torque, MM is motor torque, 1N and 2N are the proportion of the torque between the motor drive and the stick, respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000352_cp:19991054-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000352_cp:19991054-Figure5-1.png", + "caption": "Figure 5: nux plots of the experimental motor with d-axis (a) and q-axis @) excitations", + "texts": [], + "surrounding_texts": [ + "smooth bore. The sinusoidally distributed surface current density produces a stator magnetic potential distribution m(9) split into the d-axis distribution m&), symmetrical with respect to the d-axis, and q(S), symmetrical with respect to the q-axis. Moreover, the thickness of the laminations are considered so thin that a continuous iron density can be awmed, allowing an analysis on iniiniteshd increments. Let a(4) indicates the ratio between insulation thickness to insulation and lamination thickness, that is a(S)=td(ti+tl). For the sake of convenience, this ratio is expressed as function of the angular position 9. It implies that an infinitesimal thickness dt consists of a magnetic thickness (l-a(S))dt and a nonmagnetic thickness a(9)dt.\nAnalysis of the d a i s inductance. Considering the stator magnetic potential distribution in the d-axis position, the flux paths are in the direction of lamination. Only the iron permeances are considered, neglecting the interlaminar permeances. The given stator magnetic potential has to be divided between the airgap and the rotor lamioation. The latter can be neglected only if no iron saturation occws. Let us indicate the infinitesimal airgap and lamination permeances at the angular position 9 as\nwith L the motor length, D the bore diameter, g(9) and l~(9) the airgap and lamination length distributions respectively (see Appendix B). Then, from the network shown in Fig.2a, the flux corresponding to the intinitesimal angle d9 centred at the angular position 9 can be computed from\n-- dqd(9) -,,,(.g)$ P LD d4 I, (9)\ng(9)+ pfe (1 -a(a))cos(\\y - 4) (3)\nso that the flux density distribution is\nAppendix A. For a balanced comparison, each nonuniform a(9) has sinusoidal form and the same average value equal to aeO.3716. The corresponding d-axis flux linkage is reported in the caption of Fig.3. From this analysis, it results as follows:\nin linear condition (low current) the computed & is practically the same, in saturation region (high current), the motor with non-uniform a(9) presents a Ad higher, by up to 5-8 Yo.\nAnalysis of the q-axis inductance Considering the stator magnetic potential distribution in the q-axis position %(a), each lamination assumes a magnetic potential M9). Therefore the flux density in the airgap depends on the difference of magnetic potential between the stator and rotor surface. In this case, since the main path is into the nonmagnetic sheets, the iron saturation can be neglected. This is confirmed by experimental data [9-lo]. The study of the q-axis airgap flux density distribution can be canid out by refemng to the network shown in Fig.Zb, in which ra(9)d9 is the interlaminar relnctance at the angular position 4. From this network, one can write\nBy substituting ( 5 ) in the derivative of (6), one obtains\n(7) = r(s)Pg(s)+(9)\nThis is a second order ditferential equation with nonconstant coefficient 171. It can be solved by a numerical method, thw M4) is achieved. Then the airgap q-axis flux density Bq(4) can be derived. Some results are reported in Fig.4 with I=7& and a(9)=0.3716.", + "378\n2D analysis may be applied. La, and inductances are predicted separately, considering two orientations of the rotor with respect to the stator current distribution, thus only a half pole of motor need to be simulated. The phase currents are properly selected so that only d-axis or qaxis current component are obtained. Since the rotor and stator field symmetries, a symmeuy of flnx density exists in the two positions. When the rotor is in d-axis position, the vector magnetic potential along the d-axis is zero. Homogeneous Dirichlet boundary condition is given along the d-axis, while homogeneous Neumann boundary condition is given along q-axis, where flux density presents only normal component. In the same way, when rotor is in the q-axis position, Dirichlet and Neumann conditions are given along q-axis and d-axis respectively. The FEA has been firstly applied to an experimental machine (whose data are repofid in Appendix A). The flnx plots with d-axis and q-axis excitations are shown in Fig.S(a) and @) respectively.\nMeasurements and FEA result5 For the experimental motor, the measured inductances have been compared with the FEA computed ones. The good agreement in Fig.6 confums the FEA accuracy. Then, FEA was adopted to investigate the effect on the motor parameters of the non-dorm distribution of the laminations in the rotor as well as of the magnetic material, cold-sterling or grain oriented steel.\nFINITE ELEMENT ANALYSIS\nA FEA was successfully used to predict the magnetising inductance b. and L, of an ALAREL motor in [9-101, 117-201. In this paper the FEA is used to quantify the effect of the distribution of the lamination on the two axes inductances.\nDescription of the FEA on the ALAREL motor. By neglecting flux-density components in axial direction, a", + "EFFECT OF DISTRIBUTION OF LAMINATIONS\nSome results of FEA are reprted in Figs.7-9. The motors are characterised by the same stator, winding and rotor arc-pitch as the experimental machine, which is labelled A in Figs.6 and 7 and is characterised by a constant a(9)=0.3716. With the purpose to increase the saliency ratio 5, motor D has a constant a(9) higher than that of motor A. The other motors (B and C) are characterised by different non-nniform distributions of the laminations in the rotor, but with the same average value d . 3 7 1 6 . For each motor non-oriented grain steel has been used. Fig.7 shows the d q axis flux l i g e s vs. d q axis current. Fig.8 shows, in the (id, 14) plane, the constant torque curves of motors B and D. They are characterised by unsaturated 5=14 and 17.5 respectively. It is worth noticing that, even if motor D presents higher saliency ratio, motor B requires lower rated current, to achieve the same torque of 50 Nm.\nEFFECT OF TEE LAMINATION MATERIAL\nSince the stator iron is widely saturated, a further comparison has been carried out with a redesigned stator, characterised by wider stator tooth and back iron, so that the effect of the lamination steel on the motor parameters is emphasized. Figs.9-IO refa to the ALAREL motor with the redesigned stator and transformer grain-oriented steel laminations in the rotor. Motor labelled A' has uniform distribution a(5)=0.3716 (as motor A), while motors B' (corresponding to B) and E have non-uniform a@). Fig.9 reports the d q axis flux linkages showing the higher & for motors B' and E. One can observe that Ad assumes values up to 30% higher than the corresponding values of Fig.7. Fig.10 shows, in the (id, is) plane the constant torque curves of motors B and D; it highlights that motors B' and E require lower rated current than motor A' to achieve the same torque.\nCONCLUSIONS\nThe &ea of the distribution of rotor laminations of an ALAREL motor has been investigated Quality of the steel used for laminations and iron saturation have been taken into account. Both analytical and FEA calculafions have been presented. A comparison bemeen measured" + ] + }, + { + "image_filename": "designv6_24_0003350_6.1976-912-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003350_6.1976-912-Figure1-1.png", + "caption": "Figure 1 LWF F i g h t e r Simulated With R i t z Method", + "texts": [], + "surrounding_texts": [ + "AN INTEGRATED CAPABILITY FOR THE PRELIMIWRY DESIGN OF AEROEIASTICALLY TA1U)RED WINGS+\nR. W. Lynch* W. A. Rogers*\nW. W . Braymen+ General Dynamics Corporat ion\nF o r t Worth, Texas\nA b s t r a c t\nAn automated des ign procedure c a l l e d t h e Wing A e r o e l a s t i c Synthes is Procedure has been developed t o y i e l d optimum wing lamina t e s w h i l e s a t i s f y i n g a v a r i e t y of design c o n s t r a i n t s . The a lgo r i thm cons ide r s on ly the wing sk in th i ckness d i s t r i b u t i o n and o r i e n t a t i o n s f o r composites and balance masses as des ign v a r i a b l e s . It i s assumed t h a t t h e e x t e r n a l l i n e s of t h e wing have been def ined dur ing conf igu ra t ion syn thes i s and t h a t t h e s u b s t r u c t u r e will be designed a f t e r t h e s k i n s a r e def ined . The s t i f f n e s s and mass mat r ices f o r t h e wing s t r u c t u r a l box, t h e l ead ing edge, and t h e t r a i l i n g edge a r e genera ted us ing a d i r e c t RayleighR i t z energy formula t ion , which provides t h e necessary a n a l y s i s speed. The l ead ing and t r a i l i n g edge R i t z models are coupled\nv t o t h e s t r u c t u r a l box w i t h moment s p r i n g s . The opt imiza t ion scheme uses the FiaccoMcCormick non l inea r programming technique. During t h e des ign process , t h e a c t i v e parameters i nc lude f l u t t e r speed, divergence s p e e d , a i r c r a f t a n g l e of a t t a c k (which inc ludes wing f l e x i b i l i t y e f f e c t s and requi red t a i l t r i m f o r a s p e c i f i e d load f a c t o r ) , s t r a i n s and r e l a t i v e s t r a i n margins f o r t h a t load f a c t o r , f l e x i b l e l i f t , f l e x i b l e - t o - r i g i d l i f t r a t i o , r o l l e f f e c - t i v e n e s s , fundamental frequency, and s t r u c - t u r a l weight. These parameters can be included e i t h e r as measures of m e r i t t o be a d j u s t e d , as c o n s t r a i n t s , o r as both.\nAn aerodynamic a n a l y s i s provides p a r t i a l leading-edge s u c t i o n drag po la r s i nc lud ing predic ted po la r break l i f t c o e f f i c i e n t . Untrimmed and trimmed l i f t curves and drag p o l a r s a r e c a l c u l a t e d f o r the r i g i d and a e r o e l a s t i c cases .\nAn a p p l i c a t i o n of t h e procedure t o t h e des ign of an a e r o e l a s t i c a l l y t a i l o r e d wing i s presented here .\nY\n+A por t ion of t h i s r e sea rch was sponsored\nI. In t roduc t ion\nWing des ign i s a complex problem because of t h e many des ign v a r i a b l e s and c r i t e r i a involved. depth , a s p e c t r a t i o , sweep, a r e a , s p a r l o c a t i o n , s p a r spac ing , s k i n th ickness d i s t r i b u t i o n , f l a p s i z e , a c t u a t i o n system, and j i g shape o f f e r innumerable poss i - b i l i t i e s . I d e a l l y , t h e des igne r would l i k e t o t r y a l l t h e p o s s i b i l i t i e s and s e l e c t t h e b e s t performer. I n s t e a d , he must r e s t r i c t h i s a n a l y s i s t o a few of^ t h e p o s s i b i l i t i e s because of l a p s e time and manpower r equ i r ed t o mathematical ly model the conf igu ra t ion a s w e l l a s t h e high c o s t of computer t i m e . A s a r e s u l t , t h e des igne r becomes f r u s t r a t e d because he be l i eves he can develop b e t t e r des igns i f he had t ime and budget t o pursue them. Also , i t i s u s u a l l y imposs ib le t o assemble a l a r g e team of s p e c i a l i s t s a t t h e p re - l iminary des ign phase of a s tudy ; t h e r e - f o r e , t h e a v a i l a b i l i t y of a des ign t o o l that can be used e f f e c t i v e l y by one o r two i n d i v i d u a l s becomes extremely va luab le .\nI n r ecen t y e a r s , t h e r e has been g r e a t\nDesign parameters such a s span,\ni n t e r e s t i n automating t h e des ign process . The b e n e f i t s t o be der ived from automating t h e aerospace v e h i c l e des ign rocess have The des ign of a i r c r a f t s t r u c t u r e s t o s a t i s f y a e r o e l a s - t i c requirements has been the s u b j e c t of c u r r e n t r e sea rch a s demonstrated by t h e reviews o been surveyed by Heldenfe ls . ( ! )\nStroud. ( 3 f Haviland and Cooley(2) and\nThe pre l iminary wing des ign procedure descr ibed he re has ev,olved over t h e pas t n ine years a t General Dynamics. An a l g o r i - thm us ing t h e R i t z method i n conjunct ion wi th beam mode shapes f o r ana lyz ing an i so - t r o p i c p l a t e s was developed f o r t h e A i r Force Materials Laboratory i n 1967.(4)\nAn i n i t i a l p l a t e ana lys i s c a p a b i l i t y f o r r ec t angu la r p l a t e s was included i n a\nby t h e A i r Force F l i g h t Dynamics i a b o r a t o r y . *Senior S t r u c t u r e s Engineer , Member AIAA ++Senior Aerodynamics Engineer advanced composite covers . The . synthes is s y n t h e s i s prbcedure f o r t h e des ign of geom e t r i c a l l y s imple wing box s e c t i o n s w i t h\n1", + "'\nprocedure(5) , which was a l s o sponsored by t h e A i r Force Mate r i a l s Laboratory, appear - ed i n 1969. Analys is c a p a b i l i t y f o r t r a p e - z o i d a l p l a t e s was developed i n a 1970 proo,ram(6) a s a n ex tens ion of t h e i n i t i a l\n- ? l a t e a n a l y s i s a lgor i thm.\nExperience i n t r e a t i n g t h e p l a t e s t a b i l i t y problem suggested t h e analogy of p l a t e buckl ing wi th t h e s t a t i c a e r o e l a s t i c behavior of a wing. R i t z (DRR) method formed t h e b a s i s of a procedure t o be used f o r conduct ing a prel iminary a n a l y s i s of wing and empennage s t r u c t u r a l response. This procedure was developed a s a par t of the AFFDL Transonic A i r c r a f t Technology Program(7). In a d d i - t i o n , a f u l l y automated a e r o e l a s t i c synthes is procedure was formulated f o r s t a t i c a e r o e l a s t i c c o n s t r a i n t s , and t h e s t r u c t u r a l s y n t h e s i s problem f o r f l u t t e r c o n s t r a i n t s was i n v e s t i g a t e d i n pre l iminary f a sh ion . The d i r e c t Rayleigh-\nThe p i l o t ve r s ion of t h e Wing Aeroe las - t i c Synthes is Procedure (Code name TSO) was developed under c o n t r a c t w i th t h e A i r Force F l i g h t Dynamics Laboratory i n 1972.(8,9) A n a l y t i c a l c a p a b i l i t i e s f o r s t r e n g t h , s t a t i c a e r o e l a s t i c i t y , divergence, and f l u t t e r were i n t e g r a t e d wi th a non l inea r programming opt imiza t ion module. The procedure has been used t o exp lo re advanced composite wing des igns wi th va r ious ae ro - e l a s t i c c h a r a c t e r i s t i c s b S h i r k and Gr i f f in ( l0 , l1 ) and Krone(P2). v\n.- However, t h e a n a l y s i s i n t h e p i l o t procedure lacked f l e x i b l e c o n t r o l s u r f a c e s i n t h e wing s imula t ion and a provis ion f o r measuring t h e consequence of a e r o e l a s t i c phenomena on drag. Under a r e c e n t cont r a c t wi th t h e A i r Force F l i g h t Dynamics Labora to ry ( l3 ) , t h e p i l o t procedure was extended t o inc lude f l e x i b l e c o n t r o l s u r - f a c e s a n d a drag a n a l y s i s c a p a b i l i t y wi thou t compromising t h e c a p a b i l i t i e s of t h e o r i g i n a l formula t ion . The r e s u l t of these a d d i t i o n s i s an e f f i c i e n t pre l iminary des i g n procedure i n t e g r a t i n g numerous wing a n a l y s i s c a p a b i l i t i e s i n t o one computer program t h a t r e q u i r e s minimal i npu t da t a and i s economical i n computer c o s t . T h i s i s accomplished without s a c r i f i c i n g t h e s t r u c t u r a l and aerodynamic d e t a i l t h a t make t h e r e s u l t s s e n s i t i v e t o small a e r o e l a s t i c changes i n geometry, such a s camber and t w i s t . The b a s i c i dea of t h e i n t e g r a t e d c a p a b i l i t y i s t o keep t h e des igne r i n cont r o l w i t h a t i m e l y and a f f o r d a b l e a n a l y s i s of h i s des ign i n t u i t i o n . who has a working knowledge i n many d i s - c i p l i n e s provides balance wi thout overemphasizing a p a r t i c u l a r d i s c i p l i n e . T h i s i s e s p e c i a l l y t r u e i n t h e pre l iminary s t a g e s of a design. A s i n g l e des igne r\n2\nThis paper desc r ibes t h e s t r u c t u r a l and aerodynamic methodology conta ined i n the Wing A e r o e l a s t i c Syn thes i s Procedure and p r e s e n t s i l l u s t r a t i v e examples of i t s a p p l i c a t i o n .\n11. S t r u c t u r a l Simulat ion\nThe d e t a i l e d l i gh twe igh t f i g h t e r wing s imula t ion shown i n F igu re 1 c o n s i s t s of t h r e e R i t z p l a t e s w i th f r e e boundary cond i t i o n s coupled toge the r by po in t r o t a - t i o n a l sp r ings r ep resen t ing a c t u a t o r s .\nThe d i r e c t Rayleigh-Ritz formula t ion i s used t o develop s t i f f n e s s and mass ma t r i ces f o r t h e s imula t ion i n a gene ra l i zed coord i n a t e system, box is expressed as The d e f l e c t i o n of t h e wing\nwhere t h e c o e f f i c i e n t s of t h e series (a,) a r e t h e gene ra l i zed coord ina te s . e m ( t ) and '&n (3 a r e LeGendre polynomials r ep re - s e n t i n g t h e chordwise and spanwise shapes, r e s p e c t i v e l y . A l t e r n a t i v e bases have been explored, and t h e most r ap id convergence was experienced f o r t h e LeGendre polynomials.\nA l l computations ( s t r u c t u r a l , s teady a e r o e l a s t i c , and unsteady a e r o e l a s t i c ) are transformed i n t o t h e s t r u c t u r a l r e f e r - ence system of assumed modes. no i n t e r p o l a t i o n i s r equ i r ed f o r any system s t a t e eva lua t ions . This r a t h e r s imple development i s t h e key t o t h e s ign - i f i c a n t speed advantage over convent iona l f i .n i te -e lement discretization/interpolat i o n methods. Absolu te ly", + "Each p l a t e i s def ined a s a t rapezoid , w i th i t s o v e r a l l depth expressed i n polynomial form as a func t ion of t h e nondimensional t r a p e z o i d a l coord ina te s 6 and\nr l .\nThe s k i n th i ckness i s expressed i n polynomial form\ne - a1 + a2 F + a3 c 2 + a4ri + ... + (3)\nand i s sub t r ac t ed from t h e o v e r a l l depth f o r t h e computation of t h e bending s t i f f - ness .\nThe p l a t e bending s t i f f n e s s mat r ix i s then t h e polynomial\nD - 1 6 [ 3d2 - 6 t d + 4t21 [ c q e ] ( 4 )\ni n 6 and r ) over the s t r u c t u r a l box where i s t h e Hooke's L a w mat r ix .\nThe d e f l e c t i o n assumptions a r e s u r f a c e shapes der ived from t h e c r o s s products of LeGendre polynomials a s shown i n F igu re 2 .\nLfGCNDRt POLYNOMIALS FOR CONlROL SURfACiS\nv\n.--\n'I3 ' 'I' LiGiNDRt POLYNOMIALS FOR WING BOX In . I t ( -11 iq,\nF igu re 2 R i t z Method Employing LeGendre Polynomials\nThe depth , sk in th i ckness , m a t e r i a l prop e r t i e s , and t h e LeGendre polynomials a r e combined mathematical ly t o account f o r t h e s t r a i n energy of t h e s u r f a c e ,\nAs shown i n F igure 2 , each c o n t r o l s u r - f a c e i s represented by two degrees of freedom t o s imula te t h e spanwise t w i s t i n g and r igid-body f l app ing of the c o n t r o l su r f ace .\nThe s t r a i n energy of t h e c o n t r o l s u r f a c e i s uncoupled wi th t h e s t r a i n energy of t h e wing box. The s t r a i n energy ma t r ix of t h e moment sp r ings provides coupl ing terms between t h e wing box and the c o n t r o l s u r - f a c e . such a s a c t u a t o r s i s achieved by spec i fy - i n g t h e at tachment l o c a t i o n of a moment sp r ing t h a t is compatible i n r o t a t i o n w i t h t h e box on one end and wi th the c o n t r o l s u r - f a c e on t h e o t h e r end. of t h e wing box a t t h e moment s p r i n g i s proper ly transformed t o accommodate a swept hinge l i n e and i s expressed a s S t r a i n energy of moment connect ions The edge r o t a t i o n\nend t h e r o t a t i o n of t h e c o n t r o l s u r f a c e i s\nwhere X , Y i s t h e l o c a t i o n of t h e moment sp r ing i n d g i i s t h e unknown shape coe f f i c i e n t f o r which t h e equat ion i s be ing solved. Then, t h e s t r a i n energy of t h e moment sp r ing becomes\nThe spanwise bending s t i f f n e s s of t h e cont r o l s u r f a c e i s r e l a t i v e l y unimportant i n the dynamics of most wings and i s ignored i n these d e r i v a t i o n s . Spanwise bending s t r a i n energy, which could r e p r e s e n t t h e d i s t r i b u t i o n of loads t r a n s f e r r e d through a redundant hinge system, could be ca l cu - l a t e d . However, t h i s would r e q u i r e a n excess ive number of spanwise degrees of freedom and would provide a r e l a t i v e l y smal l improvement i n accuracy. In l i e u of c a l c u l a t i n g t h e s t r a i n energy of t h e spanwise bending of t h e c o n t r o l s u r f a c e , t h e l a t e r a l d e f l e c t i o n o f . t h e wing box edge i s superimposed wi th t h e t w i s t i n g and f l app ing motions of t h e c o n t r o l s u r f a c e f o r c a l c u l a t i n g t h e e x t e r n a l work, Q , of po in t loads and t h e k i n e t i c energy, T, of t h e lumped masses. The k i n e t i c energy of t h e c o n t r o l s u r f a c e i s c a l c u l a t e d f o r lumped masses only. by t w i s t i n g and f l app ing of t h e c o n t r o l s u r f a c e i s The d e f l e c t i o n caused\n3" + ] + }, + { + "image_filename": "designv6_24_0001451_0954411916648988-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001451_0954411916648988-Figure6-1.png", + "caption": "Figure 6. Actuation setup and experimental setup: (a) actuation system consisting of two linear actuators and one rotational actuator and (b) experimental setup for a gelatin phantom.", + "texts": [ + " We performed these experiments using artificial tissue phantoms (gelatin) and real cow livers; the following sections will explain the results of these experiments in detail. Experiments with gelatin phantom In order to test the behavior of the needle, we first conducted experiments in gelatin phantom. The actuation system is manufactured similar to that used in other literature4,5,8 except the existence of an additional linear actuator. The system consisted of two linear actuators and one rotational actuator, as shown in Figure 6(a). One of the linear actuators was in charge of inserting the whole needle toward the soft tissue, while one end of the linear actuator held the outside of the proximal part of the needle cannula. The second linear actuator and the rotational actuator were in charge of steering the control offset and changing the bevel\u2019s direction, respectively. The rotational actuator was placed on the linear actuator, and the proximal part of the stylet was fixed to the axis of the rotational actuator. Thus, it changes the bevel\u2019s direction directly. The second linear actuator moved the rotational actuator as well as the stylet, and thus, it changes the control offset. A steel sheath close to insertion point was used to prevent needle buckle outside the phantom. As shown in Figure 6(b), each actuator was actuated by an individual DC motor (IG-32GM, 09 type; D&J WITH Co., Ltd, Seoul, South Korea), and each motor was driven by a digital positioning controller (EPOS2 24/5; Maxon Motor AG, Sachseln, Switzerland). The controlling program was developed using Microsoft Visual Studio 2010. The phantom tissue was prepared using gelatin (Edentown F&B, Incheon, South Korea). A calibrated camera (Logitech HD Pro C920) was used to capture needle images. A pattern of 10 3 10 squares in black and white was placed at the same height as the needle in order to use it as a reference for the image processing algorithm" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002123_16.766889-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002123_16.766889-Figure1-1.png", + "caption": "Fig. 1. Schematic cross section of the oxide-planarized, self-aligned double mesa Si/SiGe-HBT with a minimum feature size of 1 m.", + "texts": [ + " After the formation of a second oxide spacer to avoid short circuits between base and emitter, TiSi and Ti(SiGe) -contacts were formed by annealing a sputtered Ti layer for 5 min at 650 C [5]. Then, a low-resistivity contact metal was patterned, reaching from the collector-base mesa onto the planarized oxide. After another SiO -layer was sputter deposited, the collector, base, and emitter contact windows were opened and the first AlSiCu-metallization level was deposited. The base contact-metal was contacted on the planarized oxide at the head of the emitter-mesa, to reduce the base-collector area of the double-mesa HBT. Fig. 1 shows a cross section of the processed transistor. For circuit fabrication this process was extended by a two-level metallization and tungsten silicide resistors. 0018\u20139383/99$10.00 1999 IEEE As shown in Fig. 2 the differential circuit configuration of the new DEMUX is comparable to the DEMUX presented in [3]. One emitter-follower pair is used in the data input, two in the clock-buffers each, allowing on-chip matching and DCcoupling. Three reference voltage sources supply the master and the slave latches and the output buffer" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001310_6.1995-1532-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001310_6.1995-1532-Figure5-1.png", + "caption": "Fig. 5: PS integration into the skin", + "texts": [ + " BRS integration into the booster skirt The parachute system is integrated into the skirt during the final assembly of the booster in Kourou. This operation is carried out by using a special jig to create the 12 degrees necessary position of the PS into the skirt. Two removable guiding rails prevent any contact of the PS with the booster during installation and adjust the position of the threaded holes for connection. The two brackets are fitted with their counter parts on the top ring. After pyro chain connections, the nose cone will close the booster (fig. 5). 19 American Institule of Aamautics and Astrmadcs A Functional description (fig.6) At the altitude for parachute extraction, the radioaltimeter generates the electric command of the nose cone release and separation device. The nose cone extracts the auxiliary parachute which pulls the cluster of three drogues. After a time delay, the drogues are separated and the mainladditional combination is extracted (ref.5). The explosive bolts for the drogue release and the reefing line cutters are both delivered by SRIPC and its suppliers" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003808_5.0027306-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003808_5.0027306-Figure2-1.png", + "caption": "FIGURE 2. Working surface of the thrust bearing (a) and mounting location for Keller pressure sensor series 2MI PA 110 (b)", + "texts": [ + " The impeller 2 of the compressor is installed closed type with low performance, the diameters of the input and output sections, respectively, are D0 = 110 mm, D2 = 240 mm. The channel width at the impeller exit is b2 = 6 mm, the number of blades Z = 21. The thrust bearing 8 is separated from the rest of the centrifugal compressor by a rubber ring 12 with a spring mounted on the high-speed rotor 6. To fix the high-speed rotor and perceive axial loads, an experimental double-sided thrust bearing with fixed pads is used (Fig. 2). The removable thrust bearing on each side of the disk has bevels of fixed pads made parallel to the inter-pad channel [3], with the following dimensions: inner and outer diameters of 70 mm and 115 mm; the number of pads Z = 8; angular length of the pad \u03b8pad = 38.8\u00b0; bevel width and depth hk = 20 mm and \u03b4bevel = h1-h2 = 0.05 mm; thrust disk width Hd = 25 mm. The value of the total axial clearance of the thrust bearing (axial movement of the rotor) was hs = 0.2...0.26 mm. The runout of the thrust disc when measured by an indicator clock was 0", + "07 mm on the outer diameter. The working surfaces of the thrust bearings are filled with B-83 GOST1320-74 babbitt. The lubricant is supplied to the bearing through the holes on each side of the thrust disk in the TB housing from the center to the periphery. In the lubrication system, Tp-22S TU 38.101821- 83 turbine oil was used. 030016-2 STAND MEASUREMENT SYSTEM To measure the pressure of the lubricating layer at four points on the average radius of the pad on each of the working and non-working sides of the TB (Fig. 2), Keller absolute pressure sensors of the 2MI PA 110 series were used. The set of sensors is located on the working and non-working sides of the thrust bearing mirror relative to the thrust disk. The sensitive element of the sensor is a piezoresistive chip with a silicon micro-layer placed in a 316L stainless steel case. The power and the output signal from the sensor were output from the end of the sensor case. Sensors were inserted into the thrust bearing slots, specially bored for them, pressed against the stop and sealed" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003053_ao.28.001036-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003053_ao.28.001036-Figure8-1.png", + "caption": "Fig. 8. Speckle pattern of a laser illuminated, seeded flow is recorded in multiple exposure photographs and processed to extract data", + "texts": [ + " This technique is suited for the study of vortical flows like those about helicopter rotor blades or airplane wings at high angles of attack. In laser speckle velocimetry, a pulsed or chopped laser beam is expanded in one dimension by a cylindrical lens to illuminate a thin fan-shaped region of the flow to be measured. The flow is seeded by small particles. A lens with an optical axis perpendicular to the illuminating beam forms an image of the illuminated particles on a photographic plate (see Fig. 8). If the seed particles are distributed densely enough, the image is a random speckle pattern caused by the interference of light reflected from the seed particles. In effect, the on the velocity field. speckles form a random grid embedded in the image of the illuminated region. If the laser is pulsed more than once, and if the motion of the fluid between pulses is large enough to cause a shift of at least one average speckle diameter but not large enough to destroy the coherence between the speckle patterns at the different pulse times, the resulting multiple exposure speckle photograph (specklegram) contains laterally shifted versions of nearly identical speckle patterns" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002458_tvt.2012.2186991-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002458_tvt.2012.2186991-Figure1-1.png", + "caption": "Fig. 1. Schematic description of the car-like autonomous vehicle.", + "texts": [ + " In Section II, a new transformation to a chain form is proposed under the assumption of pure rolling. In Section III, a more realistic kinematical model is presented, where sliding effects are modeled as unknown disturbances, and a new path-following controller, using disturbance attenuation principals, is developed. Section IV presents numerical results and demonstrates the effectiveness and feasibility of the method. Section V concludes this paper. OF AUTONOMOUS VEHICLES We consider a car-like autonomous vehicle with rear driving wheels and front steering, as presented in Fig. 1. For the vehicle in Fig. 1, the bicycle model (1) is commonly used for control system design x\u0307 = cos \u03b8v y\u0307 = sin \u03b8v \u03b8\u0307 = tan \u03c6 L v \u03c6\u0307 =w. (1) This model consists of two virtual wheels (rear and front) and represents the vehicle kinematics. The rear virtual wheel is located at the center of the rear axle (i.e., at the point (x, y)), and the front virtual wheel is placed at the front axle midpoint. L is the distance between the two wheels (i.e., car length), \u03b8 is the vehicle heading, \u03c6 is the steering angle (which is equal to the mean value of the two actual front wheel steering angles), v is the vehicle velocity, and w is the steering angle speed" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001905_2010-01-0530-Figure9-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001905_2010-01-0530-Figure9-1.png", + "caption": "Figure 9. Turntable test rig.", + "texts": [ + " To obtain the separated roll and bounce accelerations and , the following mathematical processing is applied to the obtained signals using a digital analyzer (Figure 7): Where, 2B is the distance between the two accelerometers. The response of the measured bounce and roll acceleration frequency spectrum are shown in Figures 8a & 8b. These results show a good agreement between the analytical model and the developed lab model. The bounce natural frequency is 5.2 Hz and the roll natural frequency is 1.95 Hz. The turntable test is proposed for the purpose of measuring the variation of the vehicle roll angle \u03a6 with the angular speeds . The VLM was put on the turntable device (Figure 9a). The turntable is rotated at different speeds; the change in turntable angular speed simulates the lateral acceleration change during cornering. The roll angle is measured using a sensor mounted on the VLM sprungmass (Figure 9b). ON VEHICLE RESPONSE The design parameters studied in this paper include the following: 1. The vehicle C.G. height. 2. The VLM suspension width. 3. The VLM suspension springs rate. The results obtained are divided into two categories according to the test. Figures 10a & 10b show the impulse response of bounce and roll accelerations respectively of the VLM when the C.G. height is increased from H=0.00688 m to H=0.01029 m (about 49.5% increase). The bounce natural frequency does not change (5.2 Hz), while the roll natural frequency is decreased from 1" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003322_1.4767233-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003322_1.4767233-Figure3-1.png", + "caption": "FIG. 3. (Color online) Schematic of the setup for ion implantation with scanning probe alignment with an in situ noncontact scanning force microscope image (3 3 lm2, 512 512 pixel) of a readout device with a 60 nm 300 nm channel.", + "texts": [ + "5 The implantation chamber consists of a 10 in. cube, mounted on an air table and connected to the beam line with a soft bellow for vibration isolation. We have implemented an amplitude modulated scanning force microscope that operates in dynamic noncontact mode for noninvasive in situ imaging of to-be-implanted devices and for alignment of the ion beam to regions of interest. Compared to our earlier contact SFM, the noncontact SFM enables faster imaging with higher resolution and minimal tip wear.14 Figure 3 shows a schematic of the SFM setup and an example of an in situ noncontact SFM image of a 100 nm scale silicon nanowire device. In the noncontact-SFM, cantilevers are actuated using the thermally actuated bimetal (bimorph) effect with an AC bias near the cantilever resonance frequency of 70 kHz. The cantilever-sensor integrates bimetal AC- and DC-actuation of the cantilever motion and detection of the interaction of the cantilever with the sample through a piezoresistive read-out. This technology allows for high speed noncontact imaging and also enables the ultimate miniaturization of SFMs for operation in vacuum, air, and in liquids since neither a laser nor a photodetector nor provisions for their mechanical alignment are required" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000673_tie.2018.2842736-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000673_tie.2018.2842736-Figure1-1.png", + "caption": "Fig. 1 FE model of 8pole/9slot and 8pole/24slot MFM-BDRMs: (a) 8pole/9slot MFM-BDRM; (b) 8pole/24slot MFM-BDRM", + "texts": [ + " From (15) and (16), we can see that no matter how to select sp and mp , PM and armature modulated magnetic fields always produce a lot of such harmonic magnetic fields that their harmonic orders are not the integer times of stator slots or the even times of phase number. Therefore, the FSCW of MFM-BDRM would be induced a lot of harmonic back EMF in the stator windings. III. ANALYSIS OF THE SPECIFIC FSCW OF MFM-BDRM A. Specific FSCW of MFM-BDRM with 2 1sQ p To validate the above analysis, a specific FSCW of MFM-BDRM with 2 1sQ p is designed. The stator pole pair is 4, and the corresponding stator slots are 9. Its finite element model is shown in Fig. 1(a). To comprehensively understand the performance of specific FSCW of MFM-BDRM with 2 1sQ p , a integer-slot winding of MFM-BDRM with the same size is design to compare their electromagnetic performances, as shown in Fig. 1(b). Specific parameters of two schemes are listed in Table I. To guarantee fair comparison of two schemes in analysis process, the two schemes have the same current, PM pole pairs and magnetic blocks. According to (1) and (3), when 9Q , 3N . According to (5) and (8), the stator winding coefficient can be expressed as 2\u03c0 ( \u03c0) 9sin 3\u03c0 2sin 2\u03c09 ( \u03c0) 93sin 2 dp p dk k k (17) Based on (17), the calculated absolute value of winding coefficient of any order of harmonic magnetic field is listed in Table II" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001212_6.2008-5955-Figure19-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001212_6.2008-5955-Figure19-1.png", + "caption": "Figure 19. Long Range Commercial Jet Parametric Geometry", + "texts": [ + " The vehicle was chosen to perform the same mission and have the same configuration characteristics as the Boeing 747-100 with a mission profile as shown on Figure 16. Thrust loading and wing loading of the vehicle are set to equal 0.25 and 132 lb/ft2 respectively. 17 of 23 As described before, aircraft geometric information is hierarchically modelled with support from the parametric geometry modeller that provides a unified geometric description to all disciplines. The aircraft geometry used in the validation analyses as developed by the parametric modeller is shown on Figure 17, Figure 18 and Figure 19 respectively. Comparison of the primary sizing results from the new design environment and the real aircraft data for the evaluated examples is shown on Table 1, Table 2 and Table 3. The implemented design environment correlated well with key aircraft parameters, not only in terms of weight but also in estimated performance such as takeoff and landing field lengths. 18 of 23 19 of 23 Note as well in Table 3 how the different aerodynamic methods used for drag buildup provide good correlation with published aerodynamic data" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000492_icarcv.2006.345263-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000492_icarcv.2006.345263-Figure7-1.png", + "caption": "Fig. 7 explains the principle of angular subdivision. Assume the current intensity of phase A or B is I, while the phase A or B conducts respectively. To ensure the angle displacement of the rotor from location a to b, we can supply phase A with a current 11 cos\u03b8II A = and supply phase B with a current 11 sin\u03b8IIB = , respectively. If we need the rotor", + "texts": [], + "surrounding_texts": [ + "Under the speed invariable control of gas engines, the characteristic parameters of the engine changes with the load changing or the other interfering factors. Self-adapted control model applies modern control theory to distinguish the characteristic parameters of the target and change the control strategy in real time, which can keep the quality of the control system in the best range. But the adjustment effect of a selfadapted control model depends upon the accuracy of the object identification, which is difficult to a complicated system, so the traditional PID algorithm still used widely in many cases. There are many adjustment methods of PID parameters, but the most of methods are based on characteristics of the target. With the development of computer technology, we can make use of the adjustment experience of experts as knowledge, which can be pre-restored in a computer by the way of Artificial Intelligence. In this kind of intelligence system, the system adjusts PID parameter according to the situation of the environment automatically. This kind of controller combines the conventional PID control with advanced expert system to achieve the best control, so it is called as an intellectual PID controller. Under this situation, control rules are formed by the experience of experts, and then they can be used to realize the best adjustment of PID parameters by fuzzy inference. Because it is difficult to describe the experience of experts accurately and it is not easy to describe the evaluation in quantity, the fuzzy logic becomes a valid way of solving the problem. The point of fuzzy logic is to map the error space of speed to the throttle position space, and the primary mechanism for doing this is a list of if-then statements called rules. All rules are evaluated in parallel, and the order of the rules is unimportant. The rules themselves are useful because they refer to speed error and difference of the error that describe the speed. In one word, the concept of fuzzy inference is a method that interprets the values in the speed vector and, based on some set of rules, assigns values to the throttle position vector. In our system, this interpretation is indirect, that is, we take fuzzy inference to adjust the parameters of PID, and the adjustment of throttle position is achieved by PID algorithm. The mechanism is called fuzzy adaptive PID control. The parameters setting of a PID controller is a primary matter in control system design. It confirms the magnitude of proportional coefficient, integral time and difference coefficient of the PID controller. The parameter setting methods of traditional PID controller include critical proportion method, response curve method and attenuation method. The properties of control model will no longer change in the system working period. It makes governors unable to have a best performance in a wide range. Thus we choose fuzzy inference for parameter setting. Using the fuzzy inference, it can achieve the optimum adjustment of PID parameter automatically; thereby the governor will work well in an extent working region. Adaptive fuzzy PID controller regards speed error e and error change ec as the input, which can satisfy the requirement to PID parameter self-setting at different moments. Adaptive fuzzy PID controller uses fuzzy control rules to modify PID parameter automatically, and its structure is shown in Fig. 3. The fuzzy self-adjustment of PID parameter is to find the fuzzy relationship among three PID parameters, e and ec . With the e and ec of the current time, we can modify three PID parameters according to fuzzy rules so that they can satisfy Step Motor Fig.2 Block chart of the system hardware SCI A/D GPIO TM S3 20 F2 81 2 CAP EV M od ul e Speed display Speed setting AT89S51 Active Filter Throttle actuator Speed detection Step Motor Driver Throttle position sensor Engine Fig.3 Adaptive fuzzy PID controller structure Actual speed Given speed + Engine - Error Fuzzy Inference PID Regulator pk ik dk Throttle dt de the different request to the control parameter, then make the speed control of gas engine have a good performance in different working environment. Fuzzy self-setting PID is based upon PID algorithm. It carries out the fuzzy inference using fuzzy rules by operating the error e and error change ec currently, which achieves the parameter adjustment by inquiring about the fuzzy matrix. Based on the expert experience, the fuzzy inference rules of the gas engine can be summarized as below: When e is bigger, pK should be bigger and iK should be smaller; it will increase the response speed of the speed control system. When e is middle, pK should be smaller and iK should be middle, it will make the control system have a smaller overadjustment. When e is smaller, pK and iK should be bigger, it will make the control system have a perfect steady performance. When e\u2206 is bigger, pK should be bigger and iK should be smaller, it will make the control system have a smaller overadjustment too. When e\u2206 is smaller, pK should be smaller and iK should be bigger to reduce the static error of adjustment. According to the above fuzzy rules, we can build a fuzzy relationship as below: At the running time, we establish the fuzzy inference table in DSP memory. With look-up-table, a fuzzy inference mechanism can be achieved easily. After look-up-table, increment PID parameter pk , ik are put into the following formula: { } { }iiiii piipp ecekk ecekk , , ' ' += += (1) Then the control system achieves the self-adjustment of PID parameters. The dynamic curve of the throttle fuzzy PI regulator is shown in Fig.4. We can see from the figure, the regulator has advantages of response fast, over-adjustment small, and no oscillation. IV. SPEED DETECTION OF THE GAS ENGINE Speed detection is achieved by a pulse-encoding disk, which has several holes around. When the disk turns up, the optical coupler fixed beside the disk detects the infrared light leaked through the holes. Because the light leaked from the holes changes gradually, the detected signal has a waveform of sine wave approximately. A simple pulse reshaping circuit is designed to change the sine waveform into rectangle waveform. After the waveform is changed to rectangle waveform, the speed signal is connected to the capture unit of TMS320F2812. The speed calculation is achieved in a pulse capture interrupt service. TMS320F2812 has 6 independent capture units, and all capture inputs are synchronized with the CPU clock. The capture unit provides a logging function for different events or transitions. The values of the selected General-Purpose timer counter is captured and stored in the two-level-deep FIFO stacks when selected transitions are detected on capture input pins, CAPX (X = 1, 2, or 3 for EVA; and X = 4, 5, or 6 for EVB). The capture unit consists of three capture circuits. The capture pins can also be used as generalpurpose interrupt pins, if they are not used for the capture function. There are three general methods for speed detection: M method, T method and M/T method. M method is based on the pulse count in a certain time interval and the point of M method is measuring the frequency of the pulse source. The higher the speed is, the better is the precision. T method is measuring the time period of two adjacent pulse or two transitions detected on the capture pins, and it is suitable for a lower revolution measurement. M/T method is a combination of M method and T method. Because the gas engine works in a situation of lower speed range, we choose T method for speed measuring, shown as Fig. 5. Assume clock frequency is f, P is the aperture count of the encoder, and the counting per minute obtained from Fig.5 is: PT fN 60= (2) The relative error of the measurement is: Fig.4 Dynamic adjustment curve of the inner-loop fuzzy PI regulator Pulse from Encoder CPU clock I Fig.5 Speed measuring with T method )1( \u00b1=\u2206 \u2206\u2248 \u2206+ \u2206=\u2206= T T T TT T N Ner (3) Obviously, when the speed of the engine is lower, T is bigger and the measuring error will decrease. V. FINE STEP CONTROL OF THROTTLE OPENING In the research, we take a step motor as the throttle actuator element to achieve the throttle control of the engine. For the reason of increasing the step control precision, the step motor works in a single-double beat mode. The driver of the step motor employs current-invariant chopped-wave to control angular subdivision, which can fractionize a mechanical motor step elaborately, reduce the step-lost, and improve the speed of dynamic response of the step motor. Step motors have many kinds of drive circuits, such as single voltage drive, high-low voltage drive and currentinvariant chopped-wave drive. Because the current-invariant chopped-wave drive can work under a high voltage, it can accelerate the rising and dropping speed of the current greatly. At the same time, it makes the currents in the step motor coils keep the same, which ensures the motor working steadily in low frequency situation, and increasing the output torque in high frequency area. This character improves the frequency characteristics of angular moment. Current-invariant choppedwave mode does not make the energy accumulated, which can restrict low frequency resonance effectively. For these reasons, we choose current-invariant chopped-wave drive mode in the driver of throttle actuator, the principle chart and current waveform are shown in Fig. 6. When iu high, the drive transistor VT1 and VT2 conduct is at the same time, power supply provides current to the coil of the step motor. Because of the restraint function of the inductance XL , the current rises gradually, and the sampling voltage su on resistance sR also increases slowly. Comparing su with the given voltage ru , a gate control signal will be produced. When su is bigger than ru , the output of comparator will be low, then the AND gate outputs low, VT1 is turned off. After VT1 opens, the current in the coil keeps flowing through VT2, sR and VD2; then voltage su on sR gets down. When su is lower than ru , comparator outputs an enable signal, VT1 conducts again. The power supply charges up the coil again. The procedure will repeat, and the current will be kept on a certain given value, until the control pin has a stop command. Angular subdivision is also called as mini-step distance control. The subdivision can reduce the step angle of the step motor, which will make the step motor run smoothly and increase the precision of control. The direction of the synthetic magnetic field can be changed with the change of current magnitude in two coils of a step motor, and then the balance position of the rotor will be changed with the change of magnetic field. This is the principle of the step distance angular subdivision. + - D/A Control Pin R su ru VD1 VD2 VT1 VT2 Lx R iu 1bu (a) Principle chart of a permanent current chop drive C R (b) Current waveform Fig.6 Principle chart and the current waveform +V angular displacement change from b to c, then just supply phase A and B with 22 cos \u03b8II A = , 22 sin\u03b8IIB = . This is the principle of the angular subdivision of a step motor. In this research, 18 sub-steps are adopted. This will separate the original step angle 1.8 degree into 0.1 degree. According to the angular subdivision steps, the characteristics of DAC and the sampling resistance sR , DSP inquires upon the look-uptable to achieve the operating of angular subdivision in real time. VI. THE MAIN FLOWING CHART OF THE SYSTEM The flowing chart of main control program is shown in Fig. 8. The flowing chart of the speed detection of the gas engine is shown in Fig.9. VII. CONCLUSION A 32 bit DSP TMS320F2812 was adopted to improve the efficiency of gas engine governors. The governor was designed with a double CPU structure, the DSP completed control rules, and a microcontroller managed human machine interface. The self-adjusted fuzzy PID algorithm made the governor have a wider working region. To increase the precision of the throttle control, an angular subdivision mechanism of the step motor was adopted. The governor has an extent working region than traditional governors and has an advantage of small excess adjustment and short transition time when the load changes." + ] + }, + { + "image_filename": "designv6_24_0001981_6.2009-6752-Figure10-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001981_6.2009-6752-Figure10-1.png", + "caption": "Figure 10. View Showing Hatch Size Relative to CBM", + "texts": [ + " This is important to compensate for potential misalignment between elements to be attached. Another observation during consideration of the CBM is that lunar surface dust is a major concern for interface seal surfaces and interface mechanisms as well as CBM Powered bolts would be susceptible to fouling with dust and CBM Capture Latches may be susceptible to jamming with dust Another observation during study of interfaces is that the Lunar Surface System Hatch Size is compatible with the use of a CBM. Please see Figure 10. III. Summary of Assembly Observations The ISS assembly operations are performed either by Docking, which is flying two elements together or by Berthing, where a robotic arm is used to position the elements in close proximity so that an interface mechanism can then structurally attach the elements. The ISS has performed many assembly operations and in all of these, knowing the relative position of the two elements is an important part of the process. On the ISS, a variety of techniques are used to convey that information to the robotic arm operator" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003372_sirf.2012.6160119-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003372_sirf.2012.6160119-Figure1-1.png", + "caption": "Fig. 1. Bi-directional MOS resistor principle", + "texts": [ + " The analysis shows comparable noise to a polysilicon resistor with the same resistance value and high linearity. Furthermore, a strategy to extend the tuning range is introduced which achieves a tuning ratio of about 19. The paper presents the proposed bi-directional resistor principle in Section II. The circuit implementation including tuning range extension is presented in Section III. Section IV presents the testchip design and measurement results. Finally a conclusion is presented in Section V. A bi-directional MOS resistor with 2nd order nonlinearity cancellation is shown in Fig. 1 [4]. Assuming transistors M1 and M2 are biased in triode region and both have same dimensions, the operating conditions are restriction to (1) where VFB is the flat-band voltage and \u03a6F is the Fermi level. If VA is higher than VB (as shown in Fig. 1), the current I flowing from node A to B can be calculated as: (2) where \u03b3 is the body-effect constant and KP is the gain ( ) ( ) 3 3 2 2 2 2 2 3 F A F BV V\u03b3 \u03c6 \u03c6 \u23ab\u23a1 \u23a4\u2212 + \u2212 + \u23ac\u23a2 \u23a5\u23a3 \u23a6\u23ad ( )2FAB FC BVV V \u03c6\u2212 \u2212\u2264 ( ){1,2 1,2 2 2P C FB F AB W I K V V V L \u03c6= \u2212 \u2212 978-1-4577-1318-7/12/$26.00 \u00a9 2012 IEEE SiRF 201237 factor of the transistors. The main part of second order non-linearity of a single MOS transistor is canceled in this equation, so we get a resistance almost independent on the signal voltage VAB", + " The noise characteristic of this active resistor is mainly defined by the noise of the MOS transistors. The noise spectral density of a transistor in triode region with zero drain-source voltage is given by (8) where T is absolute temperature and k is Boltzmann\u2019s constant. This expression shows the noise of a MOS resistor is the same as for a passive resistor with the same resistance value. In reality the thermal gate resistance noise will also contribute but this is not the dominating part. A critical design issue for the active resistor shown in Fig. 1 is the generation of control voltage VC. A common method is using a unity gain source follower structure to transfer the signal to the gate and creating a DC bias VC [4]. The disadvantages of this solution are an increased noise, high power consumption and limited tuning range, which make it not very efficient. As an alternative, the transistor gate can be DC biased by a constant voltage Vb while the signal is transferred to the gate by an AC coupling capacitor for non-linearity compensation as shown in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002575_125906-Figure13-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002575_125906-Figure13-1.png", + "caption": "Figure 13. Measurement of the ball screw\u2019s axial contact rigidity performance.", + "texts": [ + "p ( )= (2) Usually, the value of Fp is determined by certain application requirement. When the working current is a sine waveform with amplitude 3 A and frequency 1 Hz, the adjustable range of pre-tightening force Fp is as shown in figure\u00a012. Here, the adjustable ranges of the force are respectively 1141 N\u20131184 N, 1598 N\u20131632 N and 1962 N\u20132048 N in figures\u00a012(a)\u2013(c). According to the needs of practical application, we can adjust the amplitude and frequency of the working current to realize timely adjustment of the pre-tightening force using the GMM structures. As is shown in figure\u00a0 13, nut A is fixed to the flange, and the flange is fixed to the baseplate. Nut B is not fixed. When the pre-tightening force Fp is applied to the ball screw by the GMM structures, the axial elastic deformation \u03b4 between the ball and the roller path will change, which is measured by a CCD laser displacement sensor. The output signal of the laser displacement sensor is in the form of a voltage U. Relation between \u03b4 and U is as follows: U1000 m .( )\u03b4 \u00b5= (3) when Fp = 406 N, \u03b41 = \u221240 \u00b5m, as shown in figure\u00a014(a)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000083_ir-03-2018-0042-Figure10-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000083_ir-03-2018-0042-Figure10-1.png", + "caption": "Figure 10 Wearable vehicle leg with reference frames assigned", + "texts": [ + " To control the lower limb, the relationship between joint angles and position of the foot must be identified. The wearable vehicle leg is considered to be a RotationalRotational-Rotational series planar robot due to its movement on the sagittal plane only. Forward kinematics is applied to find the position of the foot if values for the joint angles are given. The joint angles is simply identified from human gait analysis (Linskell, 2019; Winter, 2006; Majernik, 2013; Kirtley, 2014). At the beginning, a reference frame is assigned to the hip, knee and ankle joint as shown in Figure 10. All-terrains wearable vehicle B.M. Sayed et al. Industrial Robot: the international journal of robotics research and application Volume 46 \u00b7 Number 6 \u00b7 2019 \u00b7 740\u2013762 To simplify the model and calculate the joint moment at the sagittal plane, the D-H parameters are assigned as given in Table II. By using the frame assigned, the Denavit\u2013Hartenberg (DH) table that represents the translational and rotational relationship between links are constructed as shown in Figure 11. Ti 1 i \u00bc cosu i sinu icosai sinu isinai aicosu i sinu i cosu icosai 0 cosu i sinai aisinu i 0 sinai cosai di 0 0 0 1 0 BB@ 1 CCA (1) The parameters of the left leg-exoskeleton can be substituted in coordinate transformationmatrix" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003855_ji-3a-1.1946.0008-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003855_ji-3a-1.1946.0008-Figure2-1.png", + "caption": "Fig. 2.\u2014Modes in field distribution between parallel plates. Left\u2014Polarization perpendicular to plates.", + "texts": [ + " To radiate maximum power the input impedance of the array at resonance must be matched to the characteristic impedance of the line. The most serious limitation of this type of array is its small frequency band-width. In the non-resonant array the frequency band-width has been made very wide by spacing the radiators at electrical separations slightly different from \\ A (guide), e.g. 200\u00b0. However, this results in the beam emerging at an angle to the normal, this angle varying with frequency (see Fig. 2). This array is terminated in a matched load to avoid reflections from the end. Emergent beam \\ Power : Sink for unused P o w e r \u20228 in here A Fig. 2.\u2014Illustrating emergence of beam at an angle with non-resonant array. to -4 i 2 0 1 1-2 1-4 1-6 1-8 2 A guide/X air Since this type of array is matched approximately along its whole length, the slot-conductance pattern required is quite different from that required for a resonant array; e.g. half-way down the array, if the power is reduced to half its initial value, double the slot conductance will be required to maintain uniform power radiated. Arrays of this type have been constructed up to 100 A long, both of shunt-displaced and shunt-inclined slots" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000745_2000-01-0498-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000745_2000-01-0498-Figure5-1.png", + "caption": "Figure 5. Torsion is led past the bellows in the shortest way possible", + "texts": [], + "surrounding_texts": [ + "The function of a decoupling element is to connect the \"moving\" engine section of the exhaust gas system with the \u201crigid\u201d area of the vehicle. By doing this, the aim is primarily to compensate for the relative (tilting) movements of the engine which occur especially in the case of transverse fittings. With the decoupling element, the gas-conducting components such as the manifold and the down pipe are relieved from mechanical tension to the extent that the principle of lightweight construction can be implemented here. The generally increasing requirements with regard to driving comfort concern not just a reduction in the level of vibration in the vehicle but also a drop in the level of noise. By employing a decoupling element, the transmission of structure-borne noise (e.g. noises generated by the charger and the engine) to subsequent areas of the exhaust gas system is reduced. This decrease in the medium- to high-frequency radiation of the components results in improved acoustics within the passenger section. Thus costly encapsulations or shieldings in these areas can be dispensed with. In today's applications, decoupling elements are installed at various positions of the exhaust gas system separate from the converter and silencer arrangement. Fig. 1 shows existing arrangements of decoupling elements. Figure 1. Arrangement of decoupling elements in the exhaust gas system" + ] + }, + { + "image_filename": "designv6_24_0001186_105011-Figure12-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001186_105011-Figure12-1.png", + "caption": "Figure 12. Passive actuators developed in this study and tested using the present experimental model structure with corresponding FDV plots: (A) friction damper; (B) negative spring with damping prototype device uses a commercial airpot damper.", + "texts": [ + " The first device is a simple friction block that slides on the platform of the present test model building (friction force = 0.2 N). The second one is a \u2018negative spring\u2019 (spring constant = \u22124.3 N cm\u22121) plus a damper (damping coefficient = 0.117 N s cm\u22121) mounted on the test model building. The negative spring with damping prototype device uses a commercial airpot damper (the black cylinder) together with a novel element that uses three magnets. Two magnets on a linear slide are attracted to a stationary magnet that is held in the arched structure mounted on the base. Figure 12 shows a comparison between theoretical predictions and experimental measurements using force\u2013displacement\u2013 velocity plots for the two sets of prototype device\u2014friction damper and negative spring damper. The friction damper is employed here because it is a simple and common device. The negative spring damper has the desired property indicated in figure 5, and the proposed bio-inspired passive actuator is inspired by this kind of damper. Furthermore, the structural response of third floor displacement corresponding to the comparison in figure 12 is shown in figure 13. The agreements in the two sets of the time-history displacement data for the first 10 s are convincing. Such comparison may indicate that reliable results can be obtained in the subsequent numerical simulations and validations of the presently suggested concept of bio-inspired passive actuator. Inspired and challenged by the simplicity and enormous capability of actuators present in bio-organisms, the authors have successfully built passive actuators which compare favorably with state-of-the-art semi-active actuators, and with potential in the direction of bio-inspired research in the future" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003378_bf02765177-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003378_bf02765177-Figure4-1.png", + "caption": "Fig. 4. Partial copy of plan of mining operations in the Gorelyi area from crosscut No. 17,", + "texts": [ + " The support consists of sections that are kinematically connected and that move frontally along the dip of the bed; these are interconnected along the strike by rigid and flexible connections along the floor and roof, respectively. The main purpose in conducting semicommercial tests was verification of the operability of the unit and the possibility of its use with unstable coal and wall rock in working thick, steep beds. For the tests on the ShchK6 shield unit, we selected a section in the Gorelyi bed from crosscut No. 17 got. + 40 shakh. Ziminka (Fig. 4). The thickness of the bed is 5.5-7.5 m, dip angle 64-74 \u00b0, coal density 1.35 tons/m 3, natural moisture content 6%. The coal of the bed has semilustrous cleavage, Grade KO. The coal mass is fissured. The lower band of the coal, with a thickness of 2 m, is extremely unstable. The strength coefficient of the coal f = 0.6-0.8. Encountered in the bed are sandstone lenses with inclusions of pyrite. There are interlayers of argillite, crumpled and unstable. The strength coefficient of the interlayer rock is f = 2-3" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000459_jae-2010-1266-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000459_jae-2010-1266-Figure2-1.png", + "caption": "Fig. 2. Photo of magnetic gear experiment device.", + "texts": [ + " Because the narrow distance between the magnets on the surface of the york gets better performance of torque transmission, the distance between the magnets is set as narrow as possible. Proposed cylindrical magnetic gear system consists of the input and output sides of magnetic gears. The driven motor is attached on the input side. On the other hand, a device for measuring the output torque is installed on the output side. After here, the input and output magnetic gear are referred as magnetic gear 1 and 2 respectively. A Fig. 2 shows the photos of experiment device from the top and the side views. A Fig. 3 shows both the gear 1 and gear 2 that are used on the device. The magnetic gear 1 uses 10 units of neodymium magnet of 20 mm\u00d7 10 mm\u00d7 4 mm while magnetic gear 2 uses 20 units of neodymium magnet with 40 mm \u00d7 11 mm \u00d7 5 mm. Poles of these magnets are counterbraced, that N pole and S pole are arranged alternately. For gear 1 and gear 2, the numbers of teeth are 10 and 20 respectively. Also the diameters of imaginary pitch circle are designed at 37 mm and 74 mm" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003391_ijvnv.2012.046175-Figure13-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003391_ijvnv.2012.046175-Figure13-1.png", + "caption": "Figure 13 Phase portrait of the ZMP in the lateral plane (experiment) (see online version for colours)", + "texts": [ + " We fix the gain of the PD controller using PD-gain tuning algorithm as explained in Section 2. The tuning result is shown in Figure 12, where we increased linearly the PD gain while evaluating the oscillation of the upper body. The optimal gain corresponds to the minimum upper body oscillation. Second, we calculated the rolling motion parameters by investigating the dynamics of the robot during stepping motion, using the phase portrait of the ZMP position in the lateral plane. The phase portrait of the ZMP position during stepping motion is shown in Figure 13, which is similar to the simulation result in Figure 9. Note that the more we increase the derivative gain the better convergence we will have, however, we are limited by the high frequencies resonance that can be triggered at high feedback gain as shown in Figure 12. To demonstrate the effect of the virtual damper-spring system of equation (37), which is implemented to work as illustrated in Figure 11. Hence, Figure 14 demonstrates how the robot uses gravity for landing and the spring energy for lifting" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000781_j.jmatprotec.2016.09.016-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000781_j.jmatprotec.2016.09.016-Figure1-1.png", + "caption": "Fig. 1. The schematic diagram of FAB-SAW.", + "texts": [ + " Introduction Welding on the large engineering structures such as ocean latforms, large-diameter pipelines and shipbuilding are usually erformed in cold and wet environments. With the development of arine construction industry and shipbuilding industrial clusters, mproving large heat-input welding technologies is rapidly becomng a critical subject. Flux aided backing submerged arc welding FAB-SAW) can easily achieve single-pass welding on the steel lates with a thickness up to 25 mm and can satisfy the requireents of the applications above, as reported by Liu (1986). As the iagram shown in Fig. 1, FAB-SAW is a simple and practical method hich can be used in cold and wet environments. In this method, he alloy powders are directly added into the weld groove which ill alloy with the molten weld metal, while a ceramic backing is nstalled on the reverse side of steels to provide refractory conainment for the alloy powders and restrict the molten weld bath ispersion. By using FAB-SAW method, two surfaces weld forming an be achieved by a one surface welding operation. However, the high FAB-SAW process heat-input of over 60 kJ/cm esults in grain coarsening, generating large quantities of both grain \u2217 Corresponding author", + " The number, size and distribution of nclusions were analyzed. The microstructure evolution, tensile and mpact properties of weld metal were studied. The relationships mong the characteristics of inclusions, microstructure evolution nd mechanical properties of weld metals were discussed. . Experimental procedures A commercial grade D32 steel plate with a thickness of 20 mm as used as the base metal. AWS 5.17 F7A4-EH14 welding wire as used with a diameter of 4.8 mm. AWS 5.17 F7A0-EH14 flux was ried at 350 \u25e6C for 2 h before use. As depicted in Fig. 1, a single V ype groove with an angle of 40\u25e6 was machined into the base metal nd the root gap of the base metal was maintained between 1 and mm during pre-assembly. The V type weld groove was filled with lloy powders. TG-B1 ceramic backing was installed on the weld roove. Run-on and run-off plates were used at two ends of the ase metal. To investigate the effect of CeO2 content on the microstrucure and mechanical properties of weld metal, alloy powders with ve different concentrations of CeO2 (0, 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001814_mfi.2014.6997693-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001814_mfi.2014.6997693-Figure2-1.png", + "caption": "Fig. 2 the sensor board", + "texts": [ + "1 Skeleton-based model of the human hand We use the inertial and magnetic sensors to estimate the orientations of palm and each finger to determine the pose of the hand, hence the size of the sensors must be considered to ensure that they are enough small to be equipped on the knuckles. With advances in MEMS technology, the sensors are becoming smaller and integrated, so that they are suitable for measuring the hand pose. The nine-axis MEMS sensor named MPU-9250 is adopted in the paper. It is a 9-axis Motion-Tracking device that combines a 3-axis gyroscope, a 3-axis accelerometer, and 3- axis magnetometer in a small 3\u00d73\u00d71mm package. We use this sensor to make the sensor board, the size is 10\u00d715\u00d72.6mm, as shown in the Fig.2. Then the sensor boards are deployed on the each section of the hand. One sensor board is put on the palm, and each finger is deployed three sensor boards. There are 16 sensor boards are used to measure the hand pose, and the measurements of the sensor are sent to the processor to computer the pose. The sketch of the design are shown in the Fig.3. The green represents the sensor board, and the red represents the processor board. Thumb Index Middle Ring Pinky Palm Proximal phalanx Middle phalanx Distal phalanx x z y According to the three kinds of sensors, there are two independent ways to determine the attitude and heading" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002482_9783527631506-Figure2.4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002482_9783527631506-Figure2.4-1.png", + "caption": "Figure 2.4 (a) Circuit with a voltage source for nodal analysis. (b) Transformed network for using nodal analysis using matrix formulation.", + "texts": [ + "3b shows the equivalent circuit redrawn after taking Laplace transformation and applying source transformation to the VCCS of value gmV(s). The currents I1 and I2 are the loop currents. Applying the suggested steps, one gets[ R1 + R2 0 0 R3 ] [ I1(s) I2(s) ] = [ Vs(s) \u2212gmV(s)R3 ] (2.1) But, V(s) = I1(s)R2. On substituting and bringing it to the left side, one gets[ R1 + R2 0 gmR2R3 R3 ] [ I1(s) I2(s) ] = [ Vs(s) 0 ] (2.2) Equation 2.2 is the loop matrix formulation for Figure 2.3a. Example 2.2. Figure 2.4a shows a circuit with a voltage source. Develop the nodal matrix formulation for the network. Figure 2.4b shows the equivalent circuit redrawn after applying source transformation to vS(t) and transforming the network using the representation for a capacitor as shown in Figure 2.2b. Since admittances are to be used, letting G = 1/R for the conductance, one can write[ G1 + G2 + sC 0 0 G3 ][ V1(s) V2(s) ] = [ Is(s) \u2212gmV1(s) ] (2.3) where Is(s) = G1VS(s). On substituting and bringing V1(s) to the left side, the final formulation becomes[ G1 + G2 + sC 0 gm G3 ][ V1(s) V2(s) ] = [ Vs(s)G1 0 ] (2.4) It is observed that Eq" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001212_6.2008-5955-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001212_6.2008-5955-Figure6-1.png", + "caption": "Figure 6. Surface generation", + "texts": [ + " In the current implementation a watertight surface which is used as the unified outer-mold-line for the different analyses is generated using the pySurface class. Different approaches can be used to intersected and loft together the different surface components, to maintain consistency and extensibility in the surface generation they are integrated through the use of a generic Surface abstract class as shown in Figure 5. Currently, AeroSurf (see e.g. 27) is used to generate a fine structured watertight surface which is used for high fidelity aerodynamic and structural analysis as shown in Figure 6(b). Capabilities are also provided to simplify the surface model as shown in Figure 6(c) which can be used by medium fidelity panel methods and/or equivalent beam structural analysis. Another surface approach using the GNU Triangulated Surface Library 28 to provide a fine unstructure surface is currently being integrated. In a similar way as the aircraft geometry implementation, the mission is hierarchically decomposed. This decomposition makes it possible for designers to generate and/or modify any mission by combining mission segments accordingly. When decomposing a mission into elementary objects, two characteristics need to be carefully considered for flexibility: the conceptual boundaries, and the relationships of these elementary mission objects", + " 12 of 23 As shown before, the aerodynamic module provides different levels of analysis complexity to choose. The simplest level provides parametric aerodynamic analyses such as those present in other state-of-practice programs. Aircraft lift, drag and stability derivatives are calculated based on parametric and semi-empirical formulations augmented with a potential aerodynamics flow non-planar multiple lifting surfaces vortex method developed in-house as shown on Figure 10 which uses the simplified geometric model shown in Figure 6(c). Drag calculations include Lift-induced, parasite, and transonic wave drag effects. The induced drag is calculated based on the vortex method Trefftz plane downwash. The Oswald efficiency factor and downwash effects are obtained from the vortex method and corrected with parametric technology models.32 Parasite drag is calculated using a detailed aircraft components build-up,33 taking into consideration viscous separation and mutual interference effects, with a skin friction formulas modeled by Sommer & Short formulation34)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003952_978-3-642-33457-3_7-Figure7.7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003952_978-3-642-33457-3_7-Figure7.7-1.png", + "caption": "Fig. 7.7 Spindle motor drawing: a top view, b lateral view and c a lateral cross-section", + "texts": [ + "5, negative motor stiffness, which results from the unbalanced magnetic force, must be, at least, an order of magnitude lower than the stiffness of the radial bearings and the unbalanced force must be lower than the bearing force capacity. The stiffness limit was critical for the case of AMB (estimated in order of 105 N/m) while the force limit was critical for the aerostatic bearings. 7. Structural robustness of the rotor was an equally important requirement for the design. Proper retaining of the magnet was crucial, particularly for a high-speed rotor with a high diameter to length ratio. All these requirements affected the design of the motor whose conceptual design is depicted in drawings in Fig. 7.7 The motor concept is explained in the rest of this section (Fig. 7.8). A laminated, slotless stator core has protrusions corresponding to the axial direction for good thermal contact with the housing. Advantages of slotless machines for very-high-speed operation were discussed in Chap. 2. Exclusion of stator teeth removes slotting-effect harmonics from the PM field while, at the same time, reduces impact of armature-field harmonics in the PM rotor. As a whole, a slotless motor is prone to be more efficient and less susceptible to rotor overheating than its slotted counterpart" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001170_s13272-011-0023-7-Figure11-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001170_s13272-011-0023-7-Figure11-1.png", + "caption": "Fig. 11 Time-averaged mass fractions examined by SAS\u2013 SST simulation", + "texts": [ + " At every outlet except outlet 1, a reduction of the absolute relative deviation from the stationary SST to the transient SAS\u2013SST and LES simulations can be observed. Especially for lower outlets, high improvements of the absolute relative deviation are present. It is tempting to speculate that this is due to the fact that SAS\u2013SST and LES are scale resolving (and thus predict a better turbulent mixing), but SST is not. Finally, a reduction of the overall absolute relative deviation from 13.8% for the stationary SST calculation to 5.35% for the time-averaged SAS\u2013SST simulation and 5.56% for the LES simulation is observed. In Fig. 11, isovolumes of the time-averaged mass frac- tions from the SAS\u2013SST simulation are shown. The investigations described in this paper show that the mixing quality of the MU at the various outlets can be determined successfully with the tracer gas method. An advantage of the tracer gas method is, besides low expenses and high flexibility (no additional integration of devices in the test bench is necessary), that there is no need of optical access to the MU which would be necessary, for instance, for LDA measurements" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003219_b978-0-08-011192-6.50018-9-Figure17-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003219_b978-0-08-011192-6.50018-9-Figure17-1.png", + "caption": "FIG. 17.", + "texts": [ + " To demonstrate such a problem it is instructive to consider the rotating disc and seal assembly illustrated on Fig. 16. This consists of a rotating seal rigidly attached to a relatively stiff disc resulting in zero slope and axial deflection at a ring of taper bolts, and constrained axially at the inner ring. By using the powerful matrix techniques, it is convenient to break the system down into small elements, i.e. cones and rings, equate conditions at junction points and derive a completely compatible solution. The second diagram, Fig. 17, which exaggerates the deflected shape shows the resulting stress analysis throughout the systems in terms of local bending stress and hoop stress. Examination of the system showed that an immediate reduction in bending stress at the junction between A and X resulted from reduction in size of the disc hub; this, of course, reduces the total weight and results in a more uniform stress distribution by increasing the minimum stresses without increasing significantly any other stresses. Methods such as these are being extended into all regions of aero engine design with the aim of producing the lightest yet most reliable configuration possible" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003351_0022-4898(87)90004-8-Figure9-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003351_0022-4898(87)90004-8-Figure9-1.png", + "caption": "FIG. 9. Cutting path of a crankshaft digger.", + "texts": [], + "surrounding_texts": [ + "VEHICLE AND MACHINERY DESIGN 149\npossible with the use of appropriate gauges. The authors point out that there is a need for proper classification of soils according to the degree of (operational) difficulty and also to predict the soil condition on the basis of continuous identification (by means of electrical resistance) and by dynamic penetrometer readings.\n4.1. Soil tillage (papers 2, 10 and 14) Under this heading, soil deformations (paper 2), energy requirement under field conditions (paper 10), and wear aspects of tool components (paper 14) are discussed. The phenomenon of soil cutting is characterized by large deformations of the soil. A generally accepted, satisfactory analytical or experimental technique to clarify the mechanism of large deformations does not exist. In paper 2, the authors acknowledge techniques such as the \"Soil Pressure Theory\" (soil mechanics), \"Limit Equilibrium Theory\" (plastic mechanics) and FEM-analysis. Although these techniques offer an effective means for estimation of, for instance, cutting resistance, they cannot be used to explain the mechanism of large deformations. The experiments described in paper 2 are conducted using a flat edge with a relatively large cutting angle (ranging from 45-75 degrees) and a small working depth--blade height ratio. The relative displacement was determined with the help of tracers. Next (cumulative) strain has been calculated.\n- - Typical cutting phenomena are described in Fig. 8 (paper 2, Fig. 10); a soil wedge tends to be formed and \"intake by shear plain failure\" can be distinguished. - - The strain analysis in the failure zone shows that (among others) volumetric expansion (dilatancy) occurs.\nInstruments. A three-component load cell is used to measure the forces on the cutting blade. Lead markers (2 mm in diameter) proved to be adequate tracers for soil movement determination. This was checked by the use of lead-oxide powder. The X-ray radiography method (here applied in a vertical plane) will, in the near future, also be applied to three-dimensional phenomena.\nA practical method to predict the energy requirement of soil tillage tools roughly is described in paper 10. Specific ploughing resistance is used as the exclusive input parameter for soil type and condition. Tool parameters are only soil-cutting intensity and tool velocities (relevant for p.t.o.-powered machinery). The tool parameters can be derived from tool design and operational data. Recently, this method has been successfully applied to existing", + "150 U.D. PERDOK\nA typical outcome for this machine is:\nE= 1.23 A + 18.7 kJ /m 3.\nE is the specific energy required; A is the same for mouldboard ploughing (= 70 kJ /m 3 for medium soil).\n- - Typical bite dimensions are a length, width and depth of 0.25 m, a tool-shaft speed of 2.25 rev/s and a forward speed of 0.5 m/s. - - From specific energy, other relevant data such as p.t.o, and drawbar power, tractor fuel consumption and task times can be derived in a relatively simple way. Estimated absolute values appeared to be about 20% higher than those measured. The effect of changes in operation (speed, r.p.m.) are reasonably well explained.", + "VEHICLE AND MACHINERY DESIGN 151\nWear of tool components (Zs) as a function of relative speed (v) and load intensity (p) (Zs = tip,v)) is the research subject of paper 14. This relation is determined empirically. Figure 10 (paper 14, Fig. 1) shows some results. The following explanations are given for wear behaviour in relation to load intensity and speed.\n- - Soil particle behaviour in the soil-tool contact area (in particular rolling versus sliding) is seen as the essential factor determining the rate of wear. Rolling particles cause less wear than sliding ones. - - Increase of wear with load intensity is explained by the fact that particles have more resistance to rolling in a denser soil matrix (caused by higher load intensities). - - Increase of speed appeared to stimulate the rolling action of artificial soil material i.e. gravel and marked ceramic cylinders (about 6 mm in diameter). A decrease in wear rate may therefore be expected at an increase of speed. - - In a particular situation, the tendency of soil particles to roll instead ofslide over the tool surface will also depend on tool surface roughness, particle geometry, soil moisture content etc.\nPAPERS PRESENTED IN SESSION IV: VEHICLE AND MACHINERY DESIGN; IMPLEMENTS\nl 1] A. FEKETE, Tractor engine operation under full-load conditions. Hungarian Institute of Agricultural Engineering, Hungary. [2] S. ICHmA, K. HYODO and Y. OOISHI, Visualization of the cutting mechanism of soils by the X_ray radiography. Mitsubishi Heavy Industries Ltd, Japan. [3] D. KORNICER, Application of morphological analysis for reconstruction and design of logging tractors. Industrija Masina i Traktora, Yugoslavia. [4] P. LINARES, Land mechanics and its influence on agricultural vehicles. Polytechnical University of Madrid, Spain. 15] S. MILIDRAG, M. GAVRIC, M. DAUTOVIC and R. HERBEZ, Optimisation selection of a system solution of hydromechanical transmission with a hydrostatic transformer for operating agricultural tracked tractors. Mechanical faculty, Sarajevo. SOUR \"BNT\", Pucarevo. Yugoslavia. [6] T. MURO, Excavating performance of bulldozer for a layered rock mass. Department of Ocean Engineering, Faculty of Engineering, Ehime University, Japan. [7] N.R. MURPHY, JR, A. S. LESSEM and B. E. REED, Experimental and analytical determination of structural and thermal behavior of tank tracks. U.S. Army Engineering Waterways Experiment Station, U.S.A. lSl K. OHMIYA, Analysis of the rotational vibrations of a small tractor. Faculty of Agriculture, Hokkaido University, Japan." + ] + }, + { + "image_filename": "designv6_24_0001026_tap.2015.2487512-Figure11-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001026_tap.2015.2487512-Figure11-1.png", + "caption": "Fig. 11. Geometry of the coaxial probe to excite and terminate the dielectrically loaded PPW feedline (Figure 3) with Dielectric - TMM13i (\u03b5r = 12.85, tan \u03b4 = 0.0019), C = 10mils, SP = 100mils, H1 = 30mils, H2 = 270mils, and H3 = 50mils.", + "texts": [ + " Figure 10 shows the desirable design properties for the cavity-backed slot antenna element. That is, its radiation performance is similar for most g values (corresponding to various scan directions). This is true for g \u2265 150mil as implied in Figure 9. Additionally, the range of coupling, through controlling the slot width, is quite large from as low as \u221217dB to \u22123.4dB. This is desirable for designing a suitable array aperture excitation. Below, g = 200mils is used to design the array as this is a good design midpoint. A coaxial probe feed (semi-rigid UT-85) is employed (see Figure 11) to excite the array. A second coax port was also placed at the far end of the PPW to measure the remaining/unradiated signal with the PPW (viz S21). Initially, this is of a width SR = 30mil and then tapers to 120mil, the width of the PPW. The other Figure parameters are LR = 50mils, LE = 100mils, and LT = 50mils. We remark that, the feed exhibits only 1 to 1.5dB of insertion loss, implying proper operation. Combining the dielectrically loaded PPW feedline, nonrectangular cavity-backed slot, and coaxial probe feedline excitation, a 20 element linear array was created (Figure 1)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003869_851385-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003869_851385-Figure5-1.png", + "caption": "Figure 5. Current ET TPS Design", + "texts": [ + " Currently this same data is being used to further reduce cost by further reducing the coverage of the ablator based on the actual temperatures, not the higher predicted heating. SUMI4ARY Since the start of the project, the External Tank has progressed from development to production. The thermal protection system has changed with changing requirements and improvements in materials and processes (Table 2), however, it is still true to the original concept. Today's changes represent an upgrading and maturing of an already efficient design. Figure 5 shows the current TPS configuration. As the production rate increases, most changes are directed toward producibility improvements and elimination of sole-source dependencies. Increased use of robotics and net-molding techniques will greatly reduce touch labor and material usage. The qualification of NCFI 22-65 for use on areas other than the aft dome would provide an alternative to CPR-488 and could potentially reduce the amount of ablator used. Other efforts are underway to develop cheaper/lighter ablators and/or cheaper foams with improved thermal capability to replace SLA-56l ablator completely" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002824_ias.2007.4347776-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002824_ias.2007.4347776-Figure5-1.png", + "caption": "Fig. 5. Equi-potential distribution.", + "texts": [ + " By the analytical method of cogging torque, the position of magnetization dead zone, , and the angle of magnetization dead zone, , for reducing harmonic component of cogging torque are shown in Table . In this paper, the dead zone positions, 1\u03b1 and 2\u03b1 , are selected as design parameters for elimination of cogging torque. Fig. 4 shows the initial model and a shape designed model for elimination of the fundamental and third harmonics frequency of cogging torque. The IPM motor, however, has an intensive saturation in flux barrier of rotor core. Therefore, a numerical analysis such as FEM should be required for a precise analysis of cogging torque. Fig. 5 shows Equi-potential distribution of two models by using 2-D FEM and flux density distribution on airgap is shown in Fig. 6. In shape design model, flux density distribution is large distorted but the cogging torque is much reduced by slot combination. Fig. 7 shows the cogging torque characteristics of a shape design model by an analytical prediction comparison with the result of initial model. The peak-peak value of cogging torque of initial model is 5.84 (kgcm), but the shape design model is generated to peak-peak value of 3" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003953_tia.1986.4504787-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003953_tia.1986.4504787-Figure7-1.png", + "caption": "Fig. 7. Torque ripple minimization in two-phase-fed BLDCM. (a) Theoretical voltage and current waveforms. (b) Simplified structure of trapezoidalEMF PM motor.", + "texts": [ + ": MINIMIZATION OF TORQUE RIPPLE Trapezoidal EMF and sinusoidal current. (b) Arbitrary EMF and current waveforms. minimization of torque ripple is applied to two-phase and three-phase feeding systems. Two-Phase Feeding In the torque expression given in (10), the dc input current Id is constant so the torque will be constant if the term [ea(wt) - ea(wt - 2ir/3)] is made constant for 7r/6 < wft < 7r/2. This condition is obtained with trapezoidal EMF's having a flat top equal to 1200 as illustrated in Fig. 7. This result can also be explained by referring to the simplified structure of a trapezoidal-EMF motor shown in Fig. 7(b). Over 600 of rotation, the conductors of phases a and b are fed by a constant current Id while remaining in a constant magnetic field. As a result, the produced torque is constant. This phenomenon is repeated every 60\u00b0. This approach has been adopted by several authors to obtain smooth torque with cylindrical rotor BLDCM using two-phase feeding scheme [4], [5]. In disk-type PM motors, it is difficult to put magnets side by side without supporting material so that the EMF's cannot be made constant over 1200" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002605_973233-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002605_973233-Figure2-1.png", + "caption": "Fig. 2. Before and After Assembly", + "texts": [ + " The original suspension bushing rates are used as the initial design for optimization. The original bushing rates and their design windows are displayed in Table 3. It took MARS 11 iterations to find the optimal bushing rates. The optimal bushing rates are shown in Table 3. The performance indices for both original and optimal bushing rates are shown in Table 4. The objective function is reduced from 10.68 to 9.74, which is an 8.8% improvement. A RETAINER CLIP PROBLEM In this example, a retainer clip (see Fig. 2) used to position the valve body of an automatic transmission is studied. The performance indices for the clip are the maximum insertion and removal forces. The target insertion force is a maximum of 14 N, while maintaining a minimum removal force of 18N. The geometrical parameters for defining the clip's shape are shown in Fig. 3. Due to symmetry only half of the clip is needed for the force analysis and calculations. As shown in Fig. 3, the clip is a curved beam composed of five straight and four circular segments" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001046_062024-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001046_062024-Figure5-1.png", + "caption": "Figure 5. Rear sensor module. (a) Overall appearance drawing. (b) Sectional view of assembly force. (c) Sectional view of applying preload. (d) Exploded view.", + "texts": [ + " In figure 4(d), before the middle module is assembled, press the buttons on both sides, the slider is squeezed to slide to both sides and shrink under the outer casing, and the rear module extends from the central rectangular area. In figure 4(e), after the rear module is assembled in place, the button is released, and the slider rebounds to the initial position by the action of the spring and extends into the slot of the rear module to fix the two modules. ISCME 2020 Journal of Physics: Conference Series 1748 (2021) 062024 IOP Publishing doi:10.1088/1742-6596/1748/6/062024 The rear sensor module is used to detect the preload force exerted on the patient. In figure 5(a), the rear module is connected to the middle module by the cooperation of the slot and the slider. When assembling, the sliding block at the tail of the middle module extends into the slot to fix the relative position of the middle module and the rear module. In figure 5(b), the spring is pre-compressed during the assembly process of the rear module. At this time, the force detected by the sensor 1 is the precompression force F1. In figure 5(c), when detecting a patient, the Putter is forced to move to the left to squeeze the sensor 1. At this time, the detected force F2 minus the pre-compression force F1 is the pre-load force applied to the patient. The appearance and dimensions of each part are shown in figure 5(d). ISCME 2020 Journal of Physics: Conference Series 1748 (2021) 062024 IOP Publishing doi:10.1088/1742-6596/1748/6/062024 This paper designs a modularized mechanical system of the vertebral detection and therapy instrument. The mechanical structure of the vertebral detection and therapy instrument relies on applying an impact force to the patient and detecting the reaction force generated by the human body to detect muscle stiffness. And by transmitting the pulse force generated by the electromagnet to the human body to generate resonance to achieve the purpose of treatment" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001573_j.apm.2014.11.058-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001573_j.apm.2014.11.058-Figure8-1.png", + "caption": "Fig. 8. Waveforms for radial components of magnetic flux densities within the inner and outer air gaps obtained using the proposed 2-D equivalent magnetic circuit network model.", + "texts": [ + " Therefore, the magnetic flux densities within the inner air gap and outer air gap can be calculated by finding the magnetic fluxes of all nodes at layer 5 and layer 3 divided by the areas of inner and outer air gaps, respectively. For numerical calculations, geometric parameters and the magnet\u2019s material properties of a coaxial magnetic gear mechanism, given in Table 1, are taken as an example to calculate the air gap flux densities. NdFeB magnets are used to increase the magnetic flux densities within the air gaps and the transmitted torque of the mechanism. The cross-section and geometric parameters of this coaxial magnetic gear mechanism are illustrated in Fig. 7. Fig. 8(a) and (b), respectively, present the cite this article in press as: Y.-C. Wu, B.-S. Jian, Magnetic field analysis of a coaxial magnetic gear mechanism by two-dimensional lent magnetic circuit network method and finite-element method, Appl. Math. Modell. (2014), http://dx.doi.org/10.1016/ 2014.11.058 waveforms for radial components of magnetic flux densities within the inner air gap and the outer air gap of the coaxial magnetic gear mechanism, when N is set as 3000, i.e. the coaxial magnetic gear mechanism is divided into 3000 parts in the circumferential direction", + " Because the high-speed inner rotor is equipped with three magnet pole-pairs, the period of magnetic flux density within the inner air gap is 360 /3 = 120 . Furthermore, the number of flux density pulsations within a period is governed by the number of steel pole-pieces and permanent magnets mounted on the high-speed inner rotor. For this coaxial magnetic gear mechanism, the numbers of steel pole-pieces and permanent magnets mounted on the high-speed inner rotor are 25 and 6, respectively. The number of flux density pulsations of the inner air gap for half of a period is 25/6, as shown in Fig. 8(a). It is noted that the largest space harmonic is successfully modulated by steel pole-pieces from 3 magnet pole-pairs in the inner air gap to 22 magnet pole-pairs in the outer air gap, which are respectively depicted in Fig. 8(a) and (b). These analytical results are further compared with the FEA results shown in the following section to verify the validity of the proposed analytical model. Please cite this article in press as: Y.-C. Wu, B.-S. Jian, Magnetic field analysis of a coaxial magnetic gear mechanism by two-dimensional equivalent magnetic circuit network method and finite-element method, Appl. Math. Modell. (2014), http://dx.doi.org/10.1016/ j.apm.2014.11.058 (a) Inner air gap (b) Outer air gap Fig. 10. Finite element meshing of the coaxial magnetic gear mechanism" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000116_itoec.2018.8740499-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000116_itoec.2018.8740499-Figure2-1.png", + "caption": "Fig. 2. Design of a four-axis eight-rotor aircraft", + "texts": [ + " It can perform short-range surveillance and investigation flights in crowded and narrow environments, and can perform stable vertical take-off and landing, and relax the requirements of the application for landing and landing environment. This aircraft has a more compact shape and smaller size. It performs well in indoors and in small environments such as urban areas. It has a simple structure, is easy to maintain, reduces the cost of use, and only needs to control four rotors. The speed can achieve a variety of flight movements, and the control is more convenient.The design structure of the aircraft is shown in Figure 2. The aircraft is equipped with an optical flow module to avoid the limitation of GPS when the aircraft is flying indoors. It realizes the automatic adjustment of the flight attitude of the drone and the real-time return of the flight parameters, thus determining the flight stability and making the flight more stable [11]. The aircraft is equipped with a binocular head and a camera, and uses a two-axis stabilized head as the camera's visual axis stabilization platform. The platform can isolate the disturbance of the drone to ensure that the camera can obtain a stable image; it can also respond to the control signal quickly, and the binocular camera collects data including depth of field data, laying the foundation for hardware integration" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000204_mop.27787-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000204_mop.27787-Figure1-1.png", + "caption": "Figure 1 Excitation of the double-layer linear array by E-polarized plane wave", + "texts": [ + " Our numerical code was validated by comparing obtained results with the known analytical solution for a single circular cylinder and numerical results given in Ref. [14]. In this article, we focus on examining the reflection of a plane wave by two-dimensional (2D) finite gratings with circular profiles, although the method described in previous section is applicable to more general scattering problems. We consider the excitation of a double-layer linear array by an E-polarized plane wave. The geometry is shown in Figure 1. Each single-layer array incorporates identical cylinders, so that the total number is 2M. This multiparameter problem includes five quantities: (1) incidence angle, u0; (2) number of cylinders, 2M; (3) relative wave number, ka (a here is a radius of one of the cylinders, chosen as normalizing value); (4) horizontal element spacing, dx and (5) vertical element spacing, dy. We limit the investigation to the symmetrical positioning of the elements, setting dx 5 dy, and consider two incidence angles u0 5 0 , 90 " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002808_cp.2010.0978-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002808_cp.2010.0978-Figure7-1.png", + "caption": "Figure 7: Direct Drive Active Stator PMG", + "texts": [ + " Figure 4 below illustrates the ideal (coil) voltage and current waveforms, square waves and in phase and actual waveforms from the 150 kW demonstrator. The current (leading the voltage) is still square wave in nature, but contains harmonics generated from network bridge and electronic commutator switching. The voltage waveform illustrates (motoring) armature reaction and a slight phase (delay) shift relative to the current waveform. The phase shift is necessary to provide volts for commutation. 6 Implementation Figure 7 below illustrates a direct drive Active Stator Permanent Magnet Generator (PMG) with the electronic commutator stacks mounted on the periphery of the stator. Figure 8 below illustrates the liquid cooled electronic commutator stack. The combination of air cooled machine and liquid cooled electronic commutator is ideal for direct drive PMGs. When power density is a key requirement, the machine can also be liquid cooled. [3] 1. Active Stator is an innovative variable speed drive topology that exploits and extends the benefits of DC machine technology, enabling significant power density improvements" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001244_tasc.2016.2543267-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001244_tasc.2016.2543267-Figure2-1.png", + "caption": "Fig. 2. Comparative analysis model of IPMSM", + "texts": [ + " Skew is a common method to reduce torque ripple. However, unexpected axial force is produced when skew is applied as explained in (1). V-skew is a methodology to cancel out axial force. The upper and lower parts produce axial forces in opposite directions and those, therefore, cancel each other out to make net force zero. To validate the effect of axial force reduction, comparative analysis is conducted with a no-skew, conventional skew and V-skew model. Comparative analysis models are shown in Fig.2, and the specifications of the analysis model are detailed in Table I. Prior to axial force analysis, the skew effect on torque is conducted to verify that V-skew fulfills the basic skew effect. In Fig. 3, torque wave forms of comparative analysis models are illustrated. When skew is applied, Torque ripple is dramatically reduced compared to the no-skew model. The Vskew model has reduces torque ripple as much as the conventional skew model. As mentioned above, generated force is calculated as (5)-(7)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000223_1.4030653-Figure12-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000223_1.4030653-Figure12-1.png", + "caption": "Fig. 12 Prototype of semi-active PCCP/PEB system (a) computer-aided design model and (b) 3D printed prototype", + "texts": [], + "surrounding_texts": [ + "Passive variable-stiffness designs have been realized via contact-aided compliant mechanisms\u2014which perform a specific task or improve the mechanism performance by contact interaction. Such mechanisms can be used to bridge the gap between the capabilities of compliant mechanism and conventional rigid body mechanisms [31]. For example, Tummala et al. [32] presented a contact-aided compliant mechanism which is flexible when bending in one direction and stiff when bending in opposite direction. Similarly, Mehta et al. [33] introduced a compliant cellular structure with an internal contact mechanism which enables them to be stretched more than the corresponding structures without contact. The PCCP represents a hybrid configuration which combines advantages of rigid and flexible link for knee exoskeleton. As shown in Fig. 3 (video is available from Youtube1), the PCCP mechanism consists of base-plate and sliding-plate attached to each other with fixed and sliding pins. The fixed-pins serve to anchor the two plates together as shown in Fig. 3. The sliding pins are rigidly attached to lower base-plate while motion relative to the sliding-plate is guided (and constrained) by sliding slots. The shape, length, the location of fixed pins and pattern of sliding slot can be customized for desired sectional contour of PCCP. PRB Model. The PRB modeling and analysis framework for compliant mechanisms [24] is particularly useful when a design has been chosen and the geometry, material properties and load conditions are available. The PRB model serves to model the deflection of flexible mechanism using rigid-body components 1www.youtube.com/watch?v\u00bcwPuYBphedz8 Journal of Mechanisms and Robotics NOVEMBER 2015, Vol. 7 / 041024-3 Downloaded From: http://mechanismsrobotics.asmedigitalcollection.asme.org/ on 01/29/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use that have equivalent force\u2013deflection characteristics. The motion is modeled by rigid links connected with revolute joint. Springs are added to the joint to accurately predict the force\u2013deflection relationships of the compliant mechanism. Thus, the PRB model provides a simple method to analyze systems, and predict the deflection path and force\u2013deflection relationships. However, carefully determining the pivot location and effective nonlinear spring constant is critical to its successful deployment [24]. The PCCP mechanism is designed to operate in two distinct modes/regimes as shown in Table 1. In the beam-bending mode (mode A), the sliding pins can move freely within the guiding slot. The beam bending mode is similar with a cantilever beam loading at the free end and can be accurately modeled by two rigid links that are joined at a pivot. A torsional spring at the pivot represents the beam\u2019s resistance to deflection. After a distinct amount of travel, the sliding pins encounter the limits of the guiding slots when the PCCP mechanism transitions to the fixed-guided bending mode (mode B). Fixed-guided bending mode is a loading condition where one end of the beam is fixed while the other is guided\u2014and can be modeled by three links connected with two revolute joint and loaded by torsional springs. These modes represent distinct stiffness regimes\u2014with significantly higher stiffness possible in fixed-guided bending mode\u2014but the stiffness transition can be smoothed by appropriately shaping the geometry of the guiding slots in beam bending mode as shown in Fig. 4. 041024-4 / Vol. 7, NOVEMBER 2015 Transactions of the ASME Downloaded From: http://mechanismsrobotics.asmedigitalcollection.asme.org/ on 01/29/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use Beam Bending Mode. At the hyperflexion or hyperextension of the knee joint, PCCP is at fixed-guided bending mode and protects knee joint by changing the stiffness of structure into extreme level (22 N at the end of PCCP and 2.9 Nm at center of PCCP for 5 deg deflection) as shown in Fig. 4. On the other hand, PCCP can provide predesigned assistive load to knee joint for better mobility of users at normal range of knee joint angle (flexion\u2014144 deg, extension\u20141.6 deg, Ref. [35]) as shown in Fig. 5. The symmetric configuration of PCCP and evenly distributed pins\u2014spanning two flexible plates, constrain the shape of structure and sustain sectional contour of PCCP as a shape of arc. Thus the sectional contour of flexible plate can be approximated as an arc whose center and radius are changing with respect to the deformation of the mechanism. The base plate deflection is calculated geometrically as shown in Fig. 5. x1 \u00bc R1 sin a1; y1 \u00bc R1 R1 cos a1; \u00f0R1 \u00bc L0=2a1\u00de (1) The deflection of sliding plate \u00f0x1; y2\u00de is x2 \u00bc R2sina2; y2 \u00bc R2 R2cosa2 (2) where R2 \u00bc R1 \u00fe g and a2 \u00bc L0=2R2, R is radius of inner plate (R1) and outer plate (R2), g is normal distance between inner and outer plate, a is arc angle of inner plate (a1) and outer plate (a2). Figure 6 shows a PRB model (one link mechanism) of a beam bending mode of PCCP. A torsional spring at the pivot represents the linkage\u2019s resistance to deflection. The location of characteristic pivot is defined by L0\u00f01 c\u00de and characteristic radius is the radius of the circular deflection path which is cL0, where c is characteristic radius factor [24]. The position of end-effector is simply x y \" # \u00bc L0\u00f01 c\u00de 0 \" # \u00fe cL0cosh cL0sinh \" # (3) The reliability of PRB model can be verified by plotting the deflection of ideal bending model and PRB model as follows (Fig. 6). An optimal value for the characteristic radius factor is fitted by an optimization search, to minimize the error between measured tip-trajectory and PRB model predictions. Considering a flexible stainless beam (E\u00bc 200 GPa, STS304) that is 310 mm long and has rectangular cross section\u2014a width of 25 mm and a thickness 0.5 mm (dimension for a plate of PCCP prototype), the torsional spring constant for PRB model is KBending \u00bc cKH EI L0 (4) Journal of Mechanisms and Robotics NOVEMBER 2015, Vol. 7 / 041024-5 Downloaded From: http://mechanismsrobotics.asmedigitalcollection.asme.org/ on 01/29/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use where c is characteristic radius factor, L0 is length of beam, KH is stiffness coefficient ranging between 2.67617 and 2.63744, E is young\u2019s modulus, and I is the moment of inertia [24]. In our study, the maximum PRB angle, H, is found to be 90 deg and the range of characteristic radius is 0.8430\u20130.8521. The arc length difference between upper and lower plate during deformation allows PCCP exoskeleton to provide increased and customized force profile to the user by combining with PEB spring which is introduced in the PEB Spring section. Upper and lower plates of PCCP are fixed to sliding and fixed portion of PEB spring separately as shown in Fig. 3. Hence, the arc length difference provides the PEB spring extension which is then translated into a corresponding force profile. Fixed-Guided Bending Mode. The stiffness of PCCP increases discontinuously and drastically when the sliding pins are limited by sliding slots as shown in Fig. 7. This fixedguided mode can be modeled by a PRB model which one end of beam is fixed while the other is guided with fixed angle. To maintain the constant angle, a resultant moment has to be present at the end. The length of unit segment of PCCP prototype is 33 mm and overall assembly consists of eight unit segments. Those are distributed symmetrically from the center of PCCP as shown in Fig. 3. For each of two springs, the torsional spring constant, K, is twice as stiff as for the case of cantilever beam and there are two springs for the PRB model. Thus, the unit segment is four times as stiff as a cantilever beam with same length [24]. For our PCCP prototype, it can be considered as rigidly connected eight fixed-guided segments. The torsional spring constant for PRB model is KFixed \u00bc m 2 4 cKH EI lunit (5) where m is number of fixed-guided segments and lunit is length of each segment. As shown in Fig. 8, at fixed-guided bending mode, PCCP can support much higher load than the case at beam bending mode. It can support 22 N (normal) at 5 deg of deflection. This unique property of PCCP can provide exclusive opportunity in knee exoskeleton design which is absent in conventional mechanisms. The stiffness of PCCP knee exoskeleton can be changed drastically when user\u2019s knee joint angle reaches dangerous region which causes serious injury to knee joint. In our study, maximum PRB angle, H, is 20 deg and the characteristic radius is 0.8517. PEB Spring. The PEB spring allows for improved regulation of the stiffness of PCCP at beam bending mode. This bio-inspired (pennate muscle, Fig. 9(a)) PEB spring consists of base and slider parts, which are connected to the compliant plates of PCCP separately. Upper and lower plates of PCCP are fixed to sliding and fixed portion of PEB spring, respectively, as shown in Fig. 9(a). Hence, the arc length difference during the bending motion serves 041024-6 / Vol. 7, NOVEMBER 2015 Transactions of the ASME Downloaded From: http://mechanismsrobotics.asmedigitalcollection.asme.org/ on 01/29/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use as the linear-extension of the PEB spring, which is then translated into a corresponding force profile. Additionally, we note that in the semi-active PEB spring design variant the unloaded position of the PEB spring can be adjusted by a linear-actuator (along various profiles), which provides additional freedom in tailoring the force-profiles. Obliquely attached multiple elastic bands structure allows higher force production with smaller range of motion and size. Multiple elastic bands connected between sliding part and base part generates linear force with respect to the displacement of sliding part as shown in Fig. 9. Figure 10(a) shows schematic of basic pattern of PEB spring and force profile with respect to displacement of sliding part (d) in Fig. 10(b). The resultant force \u00f0Fn\u00de of PEB spring can be defined by Fn \u00bc 2nkd 1 pffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi d2 \u00fe p2 p ! (6) where n is number of elastic bands, b is pennate angle, d and p are relative position of sliding and base part, Fp is tension\u2014and k is spring constant\u2014of unit elastic band. As seen in Fig. 10(b), the experimental data (measured by PASCO force transducer) closely matches the theoretical model predictions\u2014the small discrepancies are primarily due to the nonhomogeneity of elastic bands and tolerances within the experiment setup. The PCCP combined with two PEB spring can be modeled as a single link mechanism (with a torsional spring located at characteristic pivot) and a nonlinear spring fixed to the ground as depicted in Fig. 11. The linear elongation of nonlinear spring, dlnl is dlnl \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi \u00f0L0 \u00fe l0\u00de x\u00f0 \u00de2\u00fey2 q l0 \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 2\u00f0L2 0c 2 \u00fe l0L0c\u00de\u00f01 cosh\u00de \u00fe l2 0 q l0 (7) while the arc length difference between base and slider plates is DArc \u00bc Arcbase Arcslide \u00bc R1\u00f0a1 a2\u00de \u00bc R1 a1 L0 2\u00f0R1 \u00fe g\u00de (8) Journal of Mechanisms and Robotics NOVEMBER 2015, Vol. 7 / 041024-7 Downloaded From: http://mechanismsrobotics.asmedigitalcollection.asme.org/ on 01/29/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use The equivalent length of nonlinear spring (approximation model for PEB spring connected to PCCP) can be decided by finding l0 which satisfies the relationship R1 a1 L0 2\u00f0R1\u00fe g\u00de \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 2 L2 0c 2\u00fe l0L0c \u00f01 cosh\u00de\u00fe l2 0 q l0 (9) Figure 11(b) shows spring force with respect to the joint angle (h)of PEB spring and the nonlinear spring of PRB model in Fig. 11(a). Semi-Active Design for Customized Load Application. As shown in Fig. 17, the preload adjustor located between base and sliding part adjusts the preload of PEB spring which allow us to modify force and torque profile generated by PCCP/PEB system in detail for better customization. Inertial measurement unit located both on shank and thigh parts measures a joint angle to provide customized force/torque to the users. The PEB preload adjustment does not directly affect knee-joint motion but fine-tunes the effective force-profile imposed by the PCCP/PEB mechanism. The number of elastic bands directly affects the magnitude of force which PEB spring can generate and the pattern (configuration of elastic bands connecting base and sliding part) of elastic bands provides an opportunity to design customized force profile. The magnitude and profile of force-profile finetuning can be modified by number/pattern/preload of elastic bands. The linear elongation of nonlinear spring with preload adjustor, dnlp, is dlnlp \u00bc dlnl dll \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 2 L2 0c 2 \u00fe l0L0c \u00f01 cosh\u00de \u00fe l2 0 q l0 dll (10) The displacement of preload adjustor, dlnlp, can be any continuous functions which is able to be generated by a linear actuator. As noted previously, the preload adjustor can be controlled by a linear-actuator along various profiles. Figure 13 depicts (a) the PRB model and (b) theoretical spring forces of semi-active PEB spring for various preload-adjuster/linear-actuator input-profiles. The three semi-active mode curves correspond to the force generated by the PEB in the semi-active mode when the preload adjuster (linear actuator) is moved as (i) cosine (dotted line), (ii) sine (dashed line), or (iii) linear proportional (dashed\u2013dotted line) function of the joint angle. (iv) The solid line denotes the force generated by the PEB in the passive mode (when the linear actuator/preload adjuster is deactivated). The ability to harness distributed compliance within the PCCP/ PEB knee brace allows us to customize force/torque profiles within a prototype that weighs less than 500 g for passive system (and 1000 g for semi-active system) comparing with the weight of commercial knee brace (0.6\u20131.5 kg). We believe that these prototypes can be further engineered to produce lighter configurations 041024-8 / Vol. 7, NOVEMBER 2015 Transactions of the ASME Downloaded From: http://mechanismsrobotics.asmedigitalcollection.asme.org/ on 01/29/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use by mostly improving material, which we are pursuing in our future work." + ] + }, + { + "image_filename": "designv6_24_0000298_1.1807415-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000298_1.1807415-Figure6-1.png", + "caption": "Fig. 6 Temperature distribution on one cross section of bearing housing", + "texts": [], + "surrounding_texts": [ + "The hydrostatic bearing is another type of widely used fluidfilm lubrication bearing. This design employs an external highpressure supply system to form the hydrostatic pressure in the bearing to supply the load capacity. In many circumstances, this kind of bearing is also called an externally pressurized bearing. Due to the rotation of the journal both hydrodynamic and hydrostatic lubrication effects can play essential roles in this kind of journal bearing. Hence, an externally pressurized journal bearing is a hybrid-operating bearing if the journal rotation speed is high." + ] + }, + { + "image_filename": "designv6_24_0003759_a:1008935628281-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003759_a:1008935628281-Figure3-1.png", + "caption": "Figure 3. Single eye Virtual Trinocular Stereo-head (VIRTUE).", + "texts": [ + " This is sketched in Fig. 2. Ignoring torsion of individual cameras, VIRTUE has eight degrees of freedom. The head is centered at EC , fixates a point EF , and has a variable baseline r . VIRTUE is also free to rotate by 8 about the vector EC \u2212 EF . The details of VIRTUE\u2019s design, and its forward and inverse kinematics are described in Lang and Jenkin (1996a, b). Although VIRTUE could be constructed as a special purpose robotic device, the single eyed VIRtual TrinocUlar stEreo-head (VIRTUE) in Fig. 3 offers a number of advantages over the use of a specially designed stereo-head. The intrinsic camera parameters for each camera in the virtual trinocular head are identical (disregarding changes due to environmental conditions like temperature, pressure, etc.) since the same camera is used for each of the three sensors. VIRTUE is less expensive than a specifically constructed stereo head as little or no special purpose hardware is required. VIRTUE also utilizes an industrial robotic manipulator with known kinematics and a standard controller" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000761_j.jmatprotec.2016.09.004-Figure10-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000761_j.jmatprotec.2016.09.004-Figure10-1.png", + "caption": "Fig. 10. Stretching stress distribution at 1.5 s, (a) radial stress", + "texts": [ + " The deformation zone generated by the squeeze pressure strain of the die wedge indicated that when the inner metal experienced maximum tangential and axial strains, the outer layer gradually decreased in size. This finding suggests that in workpiece deformation, metal flow is hindered when the penetration capability of the stretching segment is enhanced relative to the inner surface of the outer metal. Metal flows freely when the stretching strain of a circular uniform distribution decreases gradually from the surface to the inner surface of the deformed workpiece. The surface is maximized. Fig. 10 depicts CWR-stretching stress distribution at 1.5 s. The change in cross-section stress was more stable than the deformation force of the stretching of the knifing zone given the high permeability. Stress anisotropy was maximized in the workpieces that came into contact with the die parts. The workpiece deformation region of a cross section was elliptical. The radial stress ran along the radial direction period because contact with the mold surface gradually dropped to the inner layer at this time", + " The presence of an ellipse and of surrounding metal reduced the tensile stress at the left and right sides of the local metal. The tangential stress at the outer layer was compressive, and that on the metal in the inner layer was tensile. Thus, tensile stress decreased slightly in the direction of the mold. When the stress on the inner and outer metal layers was compressive, the stress on the outer metal was lower than that on the metal plate in the inner layer. Furthermore, the tangential stress distribution was small in a wedge section of a uniformly elliptic degree. As depicted in Fig. 10(c), the mold under the direction of the workpiece and die, the stress near the contact zone and metal mold wedge, and the inner surface of the workpiece were not fully compressed. The axial flow of the outer metal in the metal surface was under tensile stress. Horizontally, the tensile stress on the outer metal in the vicinity was subjected to elliptical movement, and the inner layer metal was subjected to limited compressive stress. 6 H. Ji et al. / Journal of Materials Processing Technology 240 (2017) 1\u201311 orkpie c a t h t o t M h w T m d b t s a e h u o f t d t a Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003787_222137810796063625-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003787_222137810796063625-Figure1-1.png", + "caption": "Fig. 1: Concept of Uplifting Slide Bearing", + "texts": [ + " Furthermore, seismic isolation bearings generally have a higher initial cost and life cycle cost. Under this situation, we developed a new type of bearing 2) has been developed which ensures a safety during earthquakes, while restricting the amount of seismic displacements, enabling displacements to be controlled and minimizing indeterminate forces due to thermal effects. A hybrid bearing so to call Uplifting Slide Bearing which has both horizontal and inclined sliding surfaces has been proposed (See Fig. 1). During an ordinary condition, the horizontal sliding bearing supports vertical reactions from the superstructure. During an expansion and a contraction of girders due to thermal effects, no other than friction force is applied to the piers as the bearings slide horizontally. On the other hand, during an earthquake, the superstructure displaces horizontally, and when the displacement exceeds the clearance, it touches the inclined surface and then moves upward along the inclined sliding surface" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000298_1.1807415-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000298_1.1807415-Figure1-1.png", + "caption": "Fig. 1 Hydrodynamic bearing model", + "texts": [ + " The solutions of these three standard codes are based on Reynolds equation, while CFXTASCflow solves the general Navier\u2013Stokes equations. In addition, these standard codes employ Reynolds condition to process the pressure boundary in the circumferential direction of the bearing while the Half- Sommerfeld condition is used in the following simulations by CFX-TASCflow. Both the standard codes and the CFD solutions utilized for the results presented in this paper are based on the assumption of incompressible fluid and laminar flow in the bearing. Example 1. Consider a typical cylindrical journal bearing as shown in Fig. 1. Assume the diameter of the bearing D550 mm ~1.97 in.! with the width B525 mm ~0.984 in.! and radial clearance h050.05 mm ~1.97 mil!. The included angle of inlet pocket is a530\u00b0 while its axial land width is b53.75 mm ~0.148 in.!. The supply pressure of lubricating oil is ps5103 000 Pa ~15 lbf/in.2! and its dynamic viscosity is m51.25 31022 N s/m2 (1.8131026 lbf s/in.2). The spinning speed of journal is V51000 rad/s ~'9550 rpm!. The eccentricity ratio of journal is e50.5. Figure 1 is the geometry of the bearing model with 148 K mesh density. Figure 2 shows the pressure distribution obtained from the simulation by CFX-TASCflow on this hydrodynamic journal bearing with the same mesh density. There are many factors that affect the correctness and accuracy of simulation by CFX-TASCflow. Two of these require special attention. The first is the mesh density of the geometry. Several mesh densities, including 52, 148, and 308 K, have been calculated and compared to the results by standard lubrication codes", + " Consider the hydrodynamic bearing described in example 1. Using the small perturbation method, the increment of fluid-film force components can be described as follows: DFx5kxxx1kxyy1cxxx\u03071cxyy\u0307 , (1) DFy5kyxx1kyyy1cyxx\u03071cyyy\u0307 , where ki j (i , j5x ,y), ci j (i , j5x ,y) represent the stiffness and damping coefficients of the bearing, and x, y, x\u0307 , y\u0307 represent the perturbation of displacements and velocities in the X direction and Y direction, respectively. The coordinate system is the same as that shown in Fig. 1. By giving small perturbations of displacement either in the X direction or Y direction, the stiffness coefficients can be obtained by differentiating the resultant fluid film force components on the displacements of perturbation. In standard lubrication programs, the dimensionless perturbation of displacements x/h0 , y /h0 is usually between 0.01 and 0.005. For better comparison, in this study, three dimensionless perturbations are used: 0.01, 0.005, and 0.002 for selected CFD results. In order to calculate the damping coefficients, a special capability in CFX-TASCflow, called \u2018\u2018Moving Grid,\u2019\u2019 has been em- Table 3 Results of grid quality check Checked value Required value Skew angle ~deg", + " 127, APRIL 2005 rom: http://gasturbinespower.asmedigitalcollection.asme.org/ on 08/07/20 addition, the differences in equation basis between two kinds of solutions may also be a cause. This will be a subject in future research work. In the thermal analysis, another capability in CFX-TASCflow, \u2018\u2018Heat Transfer/Natural Convection,\u2019\u2019 is activated. The fluid property functions are introduced by employing the user\u2019s FORTRAN code. Example 2. Consider a cylindrical journal bearing as described in example 1 and Fig. 1. Here the lubricating oil is ISO VG32 ~equivalent to DTE-LIGHT!. Fluid Properties. Two cases are calculated in this thermal analysis. The first case has an assumption of constant fluid properties, which is typical in normal lubrication analysis. The second case assumes the fluid properties are a function of temperature, which is a more realistic simulation. These two cases are described in Table 8. Simulation Results. In Fig. 3, the profiles of temperature and viscosity for case 2 at 0.5 of eccentricity ratio and 9550 rpm of speed are displayed" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000260_s10846-010-9508-6-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000260_s10846-010-9508-6-Figure3-1.png", + "caption": "Fig. 3 Coordinate frames and relations", + "texts": [ + " In order to design the autonomous controller, the state of the aircraft has to be described by using 6-DOF Model of the aircraft and the Equations of Motion (EOM) [7]. For this purpose, two coordinate systems are used. The first one is the body coordinates of the UAV. The noninertial body coordinate system is fixed both in origin and orientation to the moving craft. The craft is assumed to be rigid. The second one is the Earth coordinate frame. The relation between the earth and the UAV body frames indicate the basic attitudes of the UAV like in Fig. 3. One of the ways to detect the attitudes of the UAV is the use of inertial measurement equipments (IMU). In literature, many different approaches can be seen related to the autonomous control of UAVs; some of the techniques proposed include fuzzy control [1, 8], adaptive control [9, 10], neural networks [11, 12], genetic algorithms [13] and Lyapunov Theory [14]. The architecture used by the authors in [7] for their work on a Fuzzy Logic Based Navigation Control System (FLNCS) forms the basis of the architecture for the Fuzzy Logic Based Autonomous Landing System (FLANS)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003669_2014-01-2400-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003669_2014-01-2400-Figure2-1.png", + "caption": "Figure 2. Control system architecture of the \u201cGreen Wheel Loader\u201d", + "texts": [ + " The accumulator capacity installed on the demonstrator can be varied by decoupling a portion of the accumulator's nitrogen volume, thus adapting the hybrid system's performance in order to optimize the overall system efficiency. Safety-relevant subsystems such as steering and braking, as well as the chassis, axles and peripheral systems, are adopted from the standard machine. The \u201cGreen Wheel Loader's\u201d drive system consists of various complex subsystems, which are engineered by multiple partners. To ensure time and resource-efficient development of the control strategies and software, machine functionality is modularized and distributed among the drive train subsystems. Figure 2 shows the structure of the control system. The operator inputs are interpreted by a central machine controller (ECU MC) which generates the subsystems' set values according to the implemented operating strategy. Additionally, functions such as control of peripheral systems, monitoring and calibration routines are handled. Each subsystem comprises a separate controller for realizing the machine controller's demands by controlling the subsystem's integrated actors. Furthermore, subsystem-specific data acquisition and monitoring functions are implemented" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001835_j.mechmachtheory.2012.11.004-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001835_j.mechmachtheory.2012.11.004-Figure1-1.png", + "caption": "Fig 1. Constant-breadth non-circular arc cam mechanism with parallel flat-faced double translating follower.", + "texts": [ + " 62 2 x: +34 93 401 58 13. ). ll rights reserved. 004 The constant-breadth cam mechanisms belong to the family of what is known as the desmodromic or positive drive cam mechanisms; in these the closure of the higher pairs formed by the contacts between the cam and the follower is guaranteed by means of the geometry of these elements. The constant-breadth cams may be considered conjugated in themselves when using a parallel flat-faced double follower separated a distance dc equal to the cam breadth (Fig. 1). In this case the cam-follower contact forces are contained in the same plane, which means that perpendicular moments on the cam axis are prevented, the mechanism occupies less space and its dimensions are less than the mechanism of conjugate cams. The constant-breadth cams can be circular arc cams \u2013 commonly studied \u2013 or of arbitrary geometry and can drive both translating followers and oscillating ones if the adequate design constraints are established. Researchers studying constant-breadth cammechanisms frequently refer to double-dwell applications, cam profiles of circular arcs and the displacement function of the follower, synthesized via traditional methods [1,2] basically using the cycloidal and polynomial motion curves on a monomial base [3\u20135]", + " Two coordinate systems are used, one x, y in which the cam is fixed and the other 1, 2 fixed and oriented according to the follower guide. The components of the position vector OP(\u03b8) of the cam-follower contact point P are known in the coordinate system 1, 2, and are expressed in the coordinate system x, y fixed to the cam via the rotation matrix. This expression is: OP \u03b8\u00f0 \u00def g1;2 \u00bc s\u03b8 \u03b8\u00f0 \u00de s \u03b8\u00f0 \u00de 1;2 OP \u03b8\u00f0 \u00def gx;y \u00bc S\u03b8\u00bd \u00b7 OP \u03b8\u00f0 \u00def g1;2; S\u03b8\u00bd \u00bc cos \u03b8 sin \u03b8 \u2212 sin \u03b8 cos \u03b8 : \u00f01\u00de In the constant-breadth cams that drive a parallel flat-faced double translating follower (Fig. 1), the displacement function of the follower s(\u03b8) cannot be freely designed during the entire cycle of the mechanism's functioning [1,12] as it must comply with the following equality: s \u03b8\u00f0 \u00de \u00fe s \u03b8\u00fe \u03c0\u00f0 \u00de \u00bc constant \u00bc dc: \u00f02\u00de Eq. (2) is a restriction on the design process of these mechanisms, and implies that the follower's displacement function can only be defined for the interval \u03b8\u2208 [0,\u03c0]; the segment of the displacement function for \u03b8\u2208 [\u03c0+, 2\u03c0] is an image of the previous one, obtained from the cited equality [1,6,8]", + " So the final expressions are: sn \u00bc dc\u2212s0 s1\u2212s0\u00f0 \u00de \u00bc \u2212 sn\u2212sn\u22121\u00f0 \u00de \u2192 sn\u22121 \u00bc \u22122s0 \u00fe s1 \u00fe dc s2\u22122s1 \u00fe s0\u00f0 \u00de \u00bc \u2212 sn\u22122sn\u22121 \u00fe sn\u22122\u00f0 \u00de \u2192 sn\u22122 \u00bc \u22124s0 \u00fe 4s1\u2212s2 \u00fe dc : 8< : \u00f09\u00de To sum up, the proposed procedure for automatically guaranteeing the global continuity C2 of the displacement function involves defining the B\u00e9zier curve with at least six control points bi, of which the ordinates s0, s1, and s2 of the first three are chosen according to functional criteria and the ordinates sn\u22122, sn\u22121, and sn of the last three are calculated according to Eq. (9). If more control points are used for the designed segment, all except the last three have free ordinates. Fig. 5 shows a numeric example of a displacement function s(\u03b8) and its derivatives to the third order for a parallel flat-faced double translating follower driven by a constant-breadth rotating cam equal to dc=55 mm, as shown in Fig. 1. The designed segment of the displacement function is defined using a B\u00e9zier curve of degree 7, therefore having 8 control points, being the set of free ordinates {25, 28, 32, 38, 40} mm. The ordinates of the last three control points are calculated according to what was shown in the previous section, meaning the complete set of B\u00e9zier ordinates is: {si}={25, 28, 32, 38, 40, 35, 33, 30} mm. The graphs in Fig. 5 show the continuity of the displacement function and of its first two derivatives. The cam profile of Fig. 1 corresponds to the function s(\u03b8) of this example. In Cardona and Clos [13] the parametric expression for obtaining the profile of a cam that drives a flat-faced oscillating follower is shown. This expression is: OP \u03b8\u00f0 \u00def g1;2 \u00bc l1 \u00fe l2up sin s \u03b8\u00f0 \u00de\u00f0 \u00de\u2212 l1 cos 2 s \u03b8\u00f0 \u00de\u00f0 \u00de 1\u00fe s\u03b8 \u03b8\u00f0 \u00de\u00f0 \u00de l2up cos s \u03b8\u00f0 \u00de\u00f0 \u00de \u00fe l1 cos s \u03b8\u00f0 \u00de\u00f0 \u00de sin s \u03b8\u00f0 \u00de\u00f0 \u00de 1\u00fe s\u03b8 \u03b8\u00f0 \u00de\u00f0 \u00de 8>< >: 9>= >; 1;2 OP \u03b8\u00f0 \u00def gx;y \u00bc cos \u03b8\u00f0 \u00de sin \u03b8\u00f0 \u00de \u2212 sin \u03b8\u00f0 \u00de cos \u03b8\u00f0 \u00de OP \u03b8\u00f0 \u00def g1;2: \u00f010\u00de In constant-breadth cam mechanisms that drive a parallel flat-faced double oscillating follower (Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001764_tia.2010.2057398-Figure16-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001764_tia.2010.2057398-Figure16-1.png", + "caption": "Fig. 16. Dial indicator arrangement.", + "texts": [ + " Thermocouple 1 was placed at the top of the bar and thermo- couple 3 at the bottom, both at the middle of the bar length. In the same way, thermocouples 2 and 4 were positioned on a tooth at the top and the bottom, respectively. As the model considers the heat transferred from the rotor bars to the rotor lamination, the temperature rise of the tooth is critical for the model validation process. Thermocouples 5 and 6 were placed on one of the end rings. The radial end ring expansion was measured to check the calculated value. A picture of the test setup is presented in Fig. 16. With the shaft mechanically locked, as presented in Fig. 17, the motor was energized for several seconds. Two motors were tested under this condition: 2000-hp twopole copper squirrel cage and 1500-hp four-pole aluminum squirrel cage, both rated at 4.16 kV and 60 Hz. A summary of the test results is presented in Table I. The plots of the two-pole machine, tested under 2000 V, are presented in Figs. 18 and 19. All the results for the temperature rise were consistent with the calculations, indicating that the skin effect was properly considered" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000512_icaiet.2014.27-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000512_icaiet.2014.27-Figure5-1.png", + "caption": "Figure 5. Temperature distribution around first element.", + "texts": [ + "014 m thickness, heat flux of 15680 W/m2 and heat generation rate of 25365 W/m3. The input data and boundary parameters are presented in Table 1. VI. RESULTS AND DISCUSSIONS Fig. 4 demonstrates the velocity streamlines of air from inlet to outlet. There is some turbulence observed around the elements which is obvious as air passes over the elements. Around this turbulence, the velocity reaches maximum of 1.653 m/s. The air velocity becomes 0 m/s at the center of the front face of the heating elements as these are stationary. The average velocity stays 1 m/s. Fig. 5 illustrates the temperature distribution along two XZ and XY planes over the first element. There is almost no temperature increase before the air reaches the heating elements. The temperature near the elements\u2019 surface reaches 1156 K. In areas far from the heating element\u2019s surface, very little temperature rise is observed. The average temperature is around 500 K after passing the elements which is approximately equal to the temperature of the oven inlet air. The air will be mixed and homogenized in the fan later" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002010_jae-2008-999-Figure12-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002010_jae-2008-999-Figure12-1.png", + "caption": "Fig. 12. Proposed Cutting Method taking the contact mechanism into the consideration.", + "texts": [ + " This proposed method can be applied to all other USMs which are operated by the contact mechanisms similar to the RUSM. Finally, motor characteristics can be calculated as in the follow steps. We proposed in this research the CM for the analysis of the USM\u2019s velocity. We named the proposed method the CM. This was because the velocity of the USM was calculated by cutting an elliptical motion taking the contacting time into the consideration. If an USM used j number of teeth in the 1\u03bb section as shown in Fig. 12(a) then the contact duration of a peak elliptical motion was only the upper part in the cut ones into 1/j, as shown in Fig. 12(b). The rotary speed of a RUSM was determined by the velocity of elliptical displacements of which direction was to the tangential ones about the circle of a motor. Hence, to calculate the USM\u2019S velocity, the tangential velocity to the transformed horizontal x \u2032 direction taking the contact duration (x\u2032)contact had to be calculated. And, it was derived by Eq. (88) considering the upper part shown in Fig. 12. (vx\u2032)contact = \u2206x\u2032 contacting time = \u2206x\u2032( DN f\u00d7n ) (88) where, (vx\u2032)contact: the tangential velocity to the transformed horizontal x\u2032 direction taking the contact duration \u2206x\u2019: the displaced amount to the transformed x\u2032 direction as shown in Fig. 12 DN: the divided number of an elliptical motion during the contact duration f : the frequency n: the total divided number of an elliptical motion. The important point is that we propose the \u2206x\u2032 for the calculation of the USM\u2019s velocity. The \u2206x\u2032 must not be confused with \u2206x\u2032 2 as shown in Fig. 12. This is because the finally displaced total amount to x\u2032 direction is \u2206x\u2032 not \u2206x\u2032 2 during the contact duration. Hence, when a designer checks the shape of an elliptical motion, as mentioned in Section 5.5 at the second checking factor part, a designer must consider the \u2206x\u2032 not the \u2206x\u2032 2. Figure 13 shows the elliptical motions, in which the CM is applied to, at the sampled three peak points. According to the proposed CM, the speed of a USM can be increased by increasing the number of teeth. This is because the cutting number is increased in proportional to the increase in the number of teeth. The increase of cutting number induces the reduction of the contacting time more than the decrease of the \u2206x\u2032. Hence, this phenomenon causes the rise of the motor speed from Fig. 12, Fig. 13, and Eq. (88). This theory and result are verified from the experiment data as shown in Table 2. Also, this experimental data validates the reasonableness of the CM. Although, twenty-two numbers of teeth of comb-tooth structure were the fastest, we will deal with sixteen teeth for the comb-tooth structure from now. This is because over sixteen is hard to make and expensive. The relation between the velocity to the tangential direction at a circle v[m/s] and the rotating number of circle per one second f [Hz] is expressed by Eq" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000459_jae-2010-1266-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000459_jae-2010-1266-Figure1-1.png", + "caption": "Fig. 1. Transmission principle.", + "texts": [ + " The movement as a gear is confirmed and transmission efficiency is measured through the experiment by using measurement value of the transmission torque. Also we obtain guidelines for problems in structure and the future improvement. In this research, the development of magnetic gear intends to replace the tooth of spur gear with permanent magnet. The permanent magnet in order of N, S, and N. . . were attached on the surface of cylinder body, these magnets on one pair of cylinders make magnet circuits and then transmit the torque through magnetic force. Figure 1 shows the simple principle of magnetic gear\u2019s model. If two gears with position in (a), magnetic field nearby the gear is stable. So the torque on the output shaft will not occur. In case (b), a stable state has collapsed by rotation of the input side of magnetic gear. In the output magnetic gear in (c), the different pole of magnet against the input side receives an attractive force while a repulsive force works from the same pole of magnet. This leads to the rotation of the output to become stable state" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001589_iecon.2014.7049224-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001589_iecon.2014.7049224-Figure6-1.png", + "caption": "Fig. 6. Experimental setup.", + "texts": [ + " Dnet = \u2212 Mns 2Cf Cp Fco (36) Thus, it can be said that calculation of Network Disturbance means calculation of communicative operational force. In this section, experiments were conducted to evaluate the validity of the proposed calculation method. First, experimental setup is explained in IV-A. Second, in IV-B, experimental results are presented and it is shown that both total operational force and communicative operational force can be calculated. Experiments were conducted with the system shown in Fig. 6. 1-DOF linear motors were used for a master and a slave robot. Position of motor was calculated by a linear encoder, while force signal was estimated by RFOB without a force sensor. DOB was utilized in the master and slave system to realize robust control. In this paper, master and slave robots were controlled by the same computer and time-delay was constructed in the computer. Parameters in experiments are shown in TABLE I In experiments, time-delay was designed for the following two cases. \u2022 Case1: T1=20 msec T2=20 msec \u2022 Case2: T1=30 msec T2=30 msec In each case, mater robot was operated both slowly and rapidly, and the total operational force F \u2032 o calculated by proposed method was compared with the master force Fm estimated from RFOB" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002990_s0094-114x(02)00037-x-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002990_s0094-114x(02)00037-x-Figure3-1.png", + "caption": "Fig. 3. Schematic for calculating the roll-center height of the rear axle.", + "texts": [ + " According to Reimpell and Stoll [19] there is a direct correlation between the wheel track variation and the roll-center height hR. According to the same authors, this correlation is also conflicting, in that a high roll center (which is desirable for a favorable car body attitude during cornering) implies a larger track alteration. The suspension roll center can be approximately determine by finite-center analysis as the intersection between the normal to the trajectory of the path center point S projected on the vertical plane Oxz and the car\u2019s longitudinal plane Oyz (Fig. 3). The following formula: P.A. Simionescu, D. Beale / Mechanism and Machine Theory 37 (2002) 815\u2013832 825 hR\u00f0zNj\u00de \u00bc 0:5 x2S\u00f0zNj\u00de x2S\u00f0zNj\u00fe1\u00de \u00fe z2S\u00f0zNj\u00de z2Sj\u00f0zNj\u00fe1\u00de zS\u00f0zNj\u00de zSj\u00f0zNj\u00de \u00f017\u00de has been derived for calculating the roll-center height relative to the chassis reference frame. The height of the roll-center measured from the ground will be hR\u00f0zNj\u00de \u00bc h0R\u00f0zNj\u00de x2S\u00f0zNj\u00de \u00f018\u00de In the above equations zNj and zNj\u00fe1 are two successive positions of the wheel center, sufficiently close one to the other to allow a tangent-chord approximation along the trajectory of the path center point", + " For illustrative purposes, the diagrams of the magnitude of the angular velocity x and angular acceleration e of the wheel carrier have been plotted (Fig. 6) for _zN \u00bc 1:0 m/s and \u20aczN \u00bc 0 using Eqs.(A.1) and (A.2) in Appendix A. The results of the kinematic analysis have been also used in the 3D visualization of the motion of the mechanism. Fig. 7 shows superimposed positions of the suspension mechanism solution 1, corresponding to zN0 and zN0 150 mm, viewed from the front (a) and from above (b). The circle-point-surface and the center-point-surface in Fig. 3 were produced for solution 2. They have been generated as ruled surfaces of the momentary screw axis relative to the chassis 828 P.A. Simionescu, D. Beale / Mechanism and Machine Theory 37 (2002) 815\u2013832 (the circle-point-surface) and to the wheel carrier (the center-point-surface). The inclined position of the screw axis relative to car\u2019s longitudinal axis it is due to the wheel carrier rotation around its own axis, which for solution 2 corresponds to a maximum angle c of 16.2 occurring for zN0 \u00bc 150 mm" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000770_pierb13030101-Figure12-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000770_pierb13030101-Figure12-1.png", + "caption": "Figure 12. Measured radiation pattern at different frequency, (a) antenna 1 and (b) antenna 2.", + "texts": [ + " It is seen that the real part and imaginary part of impedance lies around 50-\u2126 and 0-\u2126 respectively and oscillate about it throughout the UWB range except for the WLAN frequency region where very high mismatch corresponding to real and imaginary part of impedance occur due to which during that frequency range, power will not be accepted to radiate through antenna. The surface current distribution of the proposed antenna with EBG structure is shown in Figure 11(b). From figure it is clear that at the frequency of 4 GHz and 8 GHz the maximum current is concentrated and propagated to radiating element while at frequency of 5.5 GHz maximum current is concentrated on EBG structure due to which less current is propagated to antenna to radiate. which confirms the mechanism of EBG structure and the proposed antenna. Figure 12 presents the comparison of the measured radiation patterns between antenna 1 and antenna 2 at 3 GHz, 4.5GHz, 7GHz, and 10 GHz. The radiation patterns of antenna 1 are nearly identical with antenna 2 radiation patterns. Thus, the introduction of EBG structure has little effect on the radiation patterns. It is clear from figure that the antenna have good omnidirectional radiation patterns at 3.0GHz, however the radiation pattern at higher frequencies such as at 7 GHz and 10 GHz shows some variation from omnidirectional properties" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003586_12.981149-Figure14-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003586_12.981149-Figure14-1.png", + "caption": "Figure 14. PACE Concept", + "texts": [], + "surrounding_texts": [ + "PACE will make global ocean color measurements, essential for understanding the carbon cycle and how it both affects and is affected by climate change, potentially along with polarimetry measurements to provide extended data records on clouds and aerosols. It is planned to carry an ocean ecosystem spectrometer and a potentially a multidirectional, multi-angle, and multi-wavelength imaging polarimeter. PACE is planned for launch in the 2020 timeframe. It is currently in pre-Formulation (Phase A) focusing on requirements definition and acquisition planning. In September 2011, NASA selected members for its Science Definition Team (SDT).38" + ] + }, + { + "image_filename": "designv6_24_0000559_amm.371.617-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000559_amm.371.617-Figure1-1.png", + "caption": "Fig. 1. Universal multi-axis heads", + "texts": [ + " Our researches and achievements are presented in this paper as methodology, solutions and original contributions of the authors to develop new modular structures with characteristics, functions and performance superior to the existent ones meant to optimize the construction and configuration of adjustable multi-axis heads. Within modern manufacturing processes, multi-axis heads are frequently used equipment which has to respond to new requests of technological performance increase. Universal multi-axis heads in Fig. 1, as they appear in specialized works [3,4] and catalogues of some companies, have a box of gears whose upper housing 2 and union flange 1 are placed and supported by the main spindle of the machine-tool. The lower housing 3 is made in one piece and has the role to guide and set the All rights reserved. No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of Trans Tech Publications, www.ttp.net. (ID: 136.186.1.81, Swinburne University, Hawthorn, Australia-15/05/15,19:23:23) supports 7 of the tool holder spindles 8 which have a bidirectional adjustment" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003923_ecce.2017.8096460-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003923_ecce.2017.8096460-Figure1-1.png", + "caption": "Fig. 1 The topology of 24 stator slots/10-rotor poles RPM-FS machine. (a) The RPM-FS machine topology. (b) The key geometric parameters.", + "texts": [ + "00 \u00a92017 IEEE 2374 analyzed and a comprehensive comparison of characteristics is conducted to evaluate the merits and disadvantages of RPM-FS machines with different rotor-poles. It can be concluded that the RPM-FS machine with 14-rotor-poles exhibits the highest torque density and the widest range of constant power in section IV. Finally, some conclusions are drawn in section V. II. TOPOLOGIES AND OPERATION PRINCIPLES The topology of a typical 24-stator-slots 10-rotor-poles (24s/10p) RPM-FS machine is shown in Fig. 1(a), which evolves from a 12-stator-slots 10-rotor-poles (12s/10p) SPMFS machine, and inherits the flux-switching principle [7]. It can be found that each modular cell includes a pair of rotor teeth and a piece of sandwiched magnet, and the rotor PMs are tangentially magnetized in the same direction. 12 concentrated armature windings are wound around the stator-teeth. It is worth noting that the mutual-inductance of the two adjacent armature coils belonging to different phase windings can be reduced approximately to zero due to the flux path isolation caused by the fault-tolerant teeth unwound, and then the faulttolerant capability of RPM-FS machines is improved" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001203_cencon.2015.7409516-Figure10-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001203_cencon.2015.7409516-Figure10-1.png", + "caption": "Figure 10: Finite Element Analysis", + "texts": [ + " Based on the design equations and the magnetic circuit a design approach to the proposed structure is as shown in Figure 7. Preliminary designed motor is modeled in finite element based electromagnetic software for flux density analysis in the airgap, magnetic core and magnet rotor return ring. The material properties are assigned to the stator, airgap and rotor parts in the model. The flowchart for the FEA analysis is as shown in Figure 9. The model with only permanent magnet excitation is solved in static 2D solver. Figure 10 shows boundary setting, meshing and the magnetic flux flow directions for the proposed machine. Static 2D and transient motion solver is used to solve the model [4].The permanent magnet in the rotor assembly is assigned with NdFeB material. Figure 10 shows the FEA results from the numerical tool. Meshing and simulation results projecting the Magnetic Flux Density Figure 11 proves that the simulated design is acceptable as the static analysis produced a smooth waveform with no short circuit in between phases. 0 20 40 60 80 100 120 140 160 Rotational Angle [Deg.] 0.200 0.208 0.216 0.224 0.232 0.240 Figure 12: Inductance Characteristics Figure 13 above was captured with simulations on dynamic condition to evaluate the regenerative properties. As what expected in the static analysis at 100 degree the regenerative voltage of 330V was captured, also at 60 degree of rotational angle 0V was captured" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000584_ncc48643.2020.9056012-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000584_ncc48643.2020.9056012-Figure1-1.png", + "caption": "FIGURE 1. One stack-up of the piezoelectric film based capacitive touch screen panel.", + "texts": [ + " It is an approach to implement force sensing in capacitance TSPs by employing piezoelectric materials [4], [5] which generate charges on the surface when a force load is applied. The amount of charge produced is proportional to the strength of the force and the piezoelectric film thickness. Thus when the piezoelectric material is used as substrate of the touch sensors, the induced charge can be read by the touch sensors for force sensing. Four stack ups widely used in industry are investigated, and one of them is depicted in Fig. 1. A thin layer of the piezoelectric film (\u223c10\u00b5m) is underneath the touchscreen glass (\u223c0.5mm). The electrodes are much thinner compared to the piezoelectric film and cover glass so they are not shown in the figure. When a force is applied to the glass surface, the stress transmitted to the piezoelectric film layer will result in the induction of charges, which will be measured to interpret the force level. Among the four stack-ups, the highest responsivity and signal-to-noise ratio (SNR) are 0.42 V/N and 59" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001976_ip-g-2.1991.0011-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001976_ip-g-2.1991.0011-Figure4-1.png", + "caption": "Fig. 4 Variation ofparameters A , , A,, A , , B, , B , and 8, with ABIB", + "texts": [ + " 1 cannot be used to quantify the nonlinearity in the differential amplifiers, since this requires an expression for the differential current AI as a function of the differential input voltage AV. The model of eqn. 2 was therefore developed along empirical lines by comparing the truncated Fourier series with the differentialcurrent/ differential-voltage characteristic of Fig. 3 for each value of AB/B: b y = A, + 1 A, COS \" = I where y = AI/I The parameters T, A, , A, and B. were obtained by using the 12-point method [9]. This procedure yields a family of parameters A,,, A, and Bm which are dependent on AB/B. These parameters, shown in Fig. 4, were fitted to IEE PROCEEDINGS-G, Vol. 138, No. I , FEBRUARY 1991 Relative inputjoutput characteristic ~Josymmetrical diflerential 57 simple closed-form analytical expressions as follows: A, = 0.3041667(AB/B) A, = 0.2541667(AB/B) A4 = -O.O541665(AB/B) A, = -O.o004165(AB/B) Bl = 0.992 B5 = -0.0015066862 - 0.0049835 - l:14 B , = -0.0059399 + 0.017183 - I\",\"> I 0.2 < - < 0.4 1 ~ 1 1 - J T = 2.2, A, = A, = A, = B2 = B4 = 0.0 Now, using eqns. 2-10 calculations were made and are shown in Fig. 3, from which it is obvious that the proposed model accurately represents the differentialcurrent/differential-voltage characteristic of the asymmetrical JFET or MOSFET differential amplifier" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003953_tia.1986.4504787-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003953_tia.1986.4504787-Figure4-1.png", + "caption": "Fig. 4. Disk motor simplified structure used for torque calculation.", + "texts": [ + " Then the motor torque consists of an average component and harmonics of order multiple of six: * The damping effects of the magnets are negligible. * The stator currents are symmetrical and contain no even harmonics. * The stator winding is symmetrical. * The rotor field distribution is symmetrical about the pole axis. * The fundamental components of the feed currents and the corresponding EMF's are maintained in phase. The expression of electromagnetic torque produced by one phase can be formulated by referring to the simplified disk motor structure shown in Fig. 4. At instant t, the instantaneous torque produced by phase a is the interaction of the magnetic field B(a, t) and the current 00 Tem= To+ E T6, cos (n6c t+\u00a2>6J) n= 1 (7) where To is the average torque and T6n are harmonics torque. The torque ripple factor (TRF) can be defined as the ratio of peak-to-peak torque ripple to average torque: TTRF-=PP T TR (8) where Tpp is peak-to-peak torque ripple. In many cases, it may be more convenient to define the torque ripple factor as the ratio of harmonics torque to average torque: TH(9 TRFH = 2 (9) To 751 IEEE TRANSACTIONS ON INDUSTRY APPLICATIONS, VOL" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002998_tns.1984.4333520-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002998_tns.1984.4333520-Figure2-1.png", + "caption": "Fig 2 View of 3 Axis Positioner and Optical Bench", + "texts": [ + " Comparison shows that a semiconductor laser does stimulate similar effects to that of a LINAC and it has become a valuable method alongside topographical analysis as a laboratory-based assessment technique. Technical Description The Laser Simulator Facility consists primarily of a laser diode, a pulsing circuit, lens system and a 3 axis positioner on which is mounted the device under test. A block diagram of the complete laser system is shown in Fig 1. The laser and its pulsing circuit are mounted on an optical bench with the lens system and neutral density filters, and is targeted along the Z axis of the 3 axis positioner, Fig 2. Since the system uses a Class IIIb laser, making it unsafe to view the beam directly, it is housed in a light tight cabinet incorporating various safety interlocks which will power down the laser and prevent triggering if access is gained during operation. A photograph of the complete system is shown in Plate 1. The laser diode used in the simulator is the LD168 a single heterojection gallium arsenide stacked diode array, with a wavelength of 904 nm, an emitting area of 0.02 inches square and a radiant output power of 150watts" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001227_s12008-007-0028-y-Figure9-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001227_s12008-007-0028-y-Figure9-1.png", + "caption": "Fig. 9 \u2018Mesh 4\u2019 significant nodes and finite element, at the end of the campaign", + "texts": [ + " Hyperreduction was introduced to reduce strongly the number of DOF used to calculate KT and R. For the three meshes, Table 4 resumes the maximal number of DOF needed (from Fig. 8) compared with the FE DOF number (from Table 2). It clearly appears the gain is important. The more the FE model is refined, the more the gain is high. For the three meshes, the numbers of significant DOF have the same order of magnitude. Gain for meshes 1 and 2 are less interesting since theses meshes are less refined. Figure 9 shows solid mesh, for mesh 4. Only significant FE and nodes, associated to the significant DOF, are drawn. Only DOF localized at the extremities (boundary conditions) and alongside the solid corners are significant: all DOF inside the solid have less importance for shape functions. 3.6 The four methods comparison Goal of this study is to find out the best method to realize rapidly a campaign of load cases useful for a virtual prototype. Methods comparison is based on CPU time spent to calculate the campaign" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000083_ir-03-2018-0042-Figure11-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000083_ir-03-2018-0042-Figure11-1.png", + "caption": "Figure 11 Forces acting on the wearable vehicle leg", + "texts": [ + " At the beginning, a reference frame is assigned to the hip, knee and ankle joint as shown in Figure 10. All-terrains wearable vehicle B.M. Sayed et al. Industrial Robot: the international journal of robotics research and application Volume 46 \u00b7 Number 6 \u00b7 2019 \u00b7 740\u2013762 To simplify the model and calculate the joint moment at the sagittal plane, the D-H parameters are assigned as given in Table II. By using the frame assigned, the Denavit\u2013Hartenberg (DH) table that represents the translational and rotational relationship between links are constructed as shown in Figure 11. Ti 1 i \u00bc cosu i sinu icosai sinu isinai aicosu i sinu i cosu icosai 0 cosu i sinai aisinu i 0 sinai cosai di 0 0 0 1 0 BB@ 1 CCA (1) The parameters of the left leg-exoskeleton can be substituted in coordinate transformationmatrix. For left hip joint, transformationmatrix with i = 1 is given: T0 1 \u00bc cosu 1 sinu 1 0 L1cosu 1 sinu 1 cosu 1 0 L1sinu 1 0 0 1 0 0 0 0 1 0 BB@ 1 CCA (2) Transformationmatrix of the knee joint with i =2 is given: T1 2 \u00bc cosu 2 sinu 2 0 L2cosu 2 sinu 2 cosu 2 0 L2sinu 2 0 0 1 0 0 0 0 1 0 BB@ 1 CCA (3) Transformationmatrix of the ankle joint with i =3 is given: T2 3 \u00bc cosu 3 sinu 3 0 L3cosu 3 sinu 3 cosu 3 0 L3sinu 3 0 0 1 0 0 0 0 1 0 BB@ 1 CCA (4) The shorthand notation Cu i = cosu i, Su i = sinu i is used to simplify the form of equations", + " The second test: during stance motion, the wearable vehicle\u2019s contact force is added to obtain the same trajectories planning. This analysis was conducted to perform two complete cycles of walking motion to get the torque values. Then, the torque pattern for each joint to perform the leg stance motion is recorded to verify the ability of the actuator selected to produce enough torque to support the user\u2019s leg. During swing motion, the torque required at each joint can be determined by using free body diagrams of different joints as shown in Figure 11. The free body diagram shows the forces acting on hip, knee and ankle joints respectively. All data are calculated based on biomechanical data of the human at the worst case. The study of human body segment parameters is used to get masses and lengths of different links. The segment lengths were expressed as a percentage of the body height (Winter, 2009). Table III summarizes the lower body segment lengths expressed as a fraction of body height (H). The data was gathered and adapted from Karimi and Jahanian (2012)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003027_imece2014-38788-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003027_imece2014-38788-Figure4-1.png", + "caption": "Figure 4: Vector schematic of a four-bar function generation mechanism in both its and position", + "texts": [ + "org/about-asme/terms-of-use 5 Copyright \u00a9 2014 by ASME This section provides a review of a rigid-body synthesis technique that may be used for function, path and motion generation, and path generation with prescribed timing. The dyadic approach reviewed here is for illustrative purposes only, and the user should not be limited towards it. A vector schematic of a planar linkage, that is, using the complex number technique is proved to be the simplest, yet the most versatile method for synthesis of rigid-body mechanisms [30]. Most of the planar linkages may be thought of as a combination of vector pairs known as dyads [30]. In function generation, the vector loop closure \u0305 \u0305 \u0305 \u0305 \u0305 \u0305 (Figure 4) produces the following equation: \u0305 ( ) \u0305 ( ) \u0305 ( ) (4) where j is the precision-position. For path generation, motion generation (rigid-body guidance), and path generation with prescribed timing, loops \u0305 \u0305 \u0305 \u0305 \u0305 and \u0305 \u0305 \u0305 \u0305 \u0305 (Figure 5), formed by dyads \u0305 \u0305 and \u0305 \u0305 , respectively, produce the following equations: \u0305 ( ) \u0305 ( ) \u0305 (5) \u0305 ( ) \u0305 ( ) \u0305 (6) where j is the precision-position. Equations (4) through (6) can be expanded for each precisionposition to synthesize a rigid-body equivalent mechanism for function, path and motion generation, and path generation with prescribed timing" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001567_aps.2009.5171557-FigureI-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001567_aps.2009.5171557-FigureI-1.png", + "caption": "Fig. I.Geometry of2-mode 4-port antenna.", + "texts": [], + "surrounding_texts": [ + "The authors wish to acknowledge the Hong Kong RGC grant HKUST 618005." + ] + }, + { + "image_filename": "designv6_24_0000740_igarss.2013.6721289-Figure2.5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000740_igarss.2013.6721289-Figure2.5-1.png", + "caption": "Fig 2.5 Structure architecture and component layout (Access panels opened)", + "texts": [ + " The SDS-4 system architecture is expected to be simple, with single, non-redundant bus components. The adoption of commercial off-the-shelf (COTS) parts is also permissible, if required, subject to adequate radiation testing. The satellite structure is an aluminum honeycomb with face sheet black-anodized for thermal characteristic purposes. In structure, the satellite is divided into two areas. This design is to provide maximum access to components during harnessing and testing. The front side in Fig. 2.5 contains the mission component area. The satellite dimensions are 500 \u00d7 500 \u00d7 500 mm and its weight is around 50 kg, which is the constraint of the H-IIA LV piggyback ride[2]. We conducted the system-level tests after completing the assembly and integration of the SDS-4 Satellite. Especially, in this chapter, we focus on system real sky tests and electromagnetic compatibility (EMC) tests because they were peculiar to micro satellites and important for mission success of SDS-4. The real sky tests using the SDS-4 PFM are very unique taking advantage of micro satellites" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000715_062035-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000715_062035-Figure3-1.png", + "caption": "Figure 3(b). Heat flux for cast iron Figure 3(c). Temperature for cast iron", + "texts": [ + " As It is noticeable, for the given input voltage the Joule heat value is the maximum for Cast Iron > EN8C > Structural Steel. This means that maximum heat is produced in Cast Iron for the same voltage. However, the electric field intensity is the maximum for Structural Steel > EN8C > Cast Iron. For Heat Flux Structural Steel > EN8C > Cast Iron. This implies that the rate of heat energy transfer is maximum in Structural Steel and minimum in Cast Iron. 9 IOP Conf. Series: Materials Science and Engineering 263 (2017) 062035 doi:10.1088/1757-899X/263/6/062035 Figure 3(a) shown above, is a close up view of the heat flux variation in the teeth section in the rack. As seen, it is not even around the teeth gaps and varies individually. The figures given above depict the contours obtained for the parameters Heat Flux and Temperature for all of the simulated materials. These are based on the iterative values whose boundary limits are mentioned in the tables 5-7. As seen from figures 3(b), 4(a), 5(a) we notice that the Heat Flux is maximum in the teeth region of the rack gear and starts decreasing as we move across the length of the rod" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003682_eumc.2018.8541699-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003682_eumc.2018.8541699-Figure3-1.png", + "caption": "Fig. 3. Pyramidal absorber model", + "texts": [ + " A truncated pyramidal absorber optimized for the 2.6-4 GHz band and a FSS surface optimized for the 6-8.5 GHz band (FCC/ETSI/KCC UWB radar band). The first design is a truncated pyramidal absorber commonly used in anechoic chambers and general purpose absorbers. The material parameters of the G/PLA were defined in the simulator and the pyramidal shape was optimized for absorption in the 2.6-4 GHz band given a maximum height of 30 mm limited by the waveguide sample fixture. Unit element simulations is shown in Fig. 4. In Fig. 3 the 3D model of the pyramidal horn unit element and a WR284 waveguide section filled with absorber is presented. Due to the practical implications of a free space setup, which require a large sample size of preferably 6\u03bb \u00d7 6\u03bb, we measured a section of the pyramidal absorber structure in the waveguide test fixture. The sample is measured at fundamental mode of a rectangular waveguide, which is TE10. In this mode the magnetic wave field impinging the sample is not transverse and differs significantly from a free space plane wave propagation" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001939_2015-01-0088-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001939_2015-01-0088-Figure3-1.png", + "caption": "Figure 3. Aluminum Gear Shift Fork - FE Model", + "texts": [ + " Alloy 383 (ADC12) is alternatives to A380 for intricate components requiring improved die filling characteristics. Alloy 383 offers improved resistance to hot cracking (strength at elevated temperatures). Optimized Gear Shift Fork geometry is modeled in Pro-E, where the complete wireframe generated. The data is then translated to Initial Graphics Exchange Specification (IGES/STP) format and read into HyperMesh\u2122, where the Gear Shift Fork is FE modeled by 3D Tet mesh generation solid Element type. Fig.3 shows a typical CAE Tet meshed model of the Gear Shift Fork. According to the applied loadings originating from different categories of mechanics, this linear elastic analytical procedure could further be divided into three load steps, 1. Stiffness Evaluation (Displacement & SPC forces) 2. Stress Analysis 3. Contact Pattern Simulation Final design from Stiffness approach has been evaluated under operating loading condition - to get the equal Displacement and Reaction force at both legs [3]. Section Modulus I2 is fine tuned to obtain equal deflection within acceptable variation of 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003650_physrevb.78.165318-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003650_physrevb.78.165318-Figure1-1.png", + "caption": "FIG. 1. Schematic of a scattering region in a nanowire connected to electronic reservoirs.", + "texts": [ + " Coherent switching effects by either tuning the electronic energy or detuning one of the side-coupled dots are discussed. Finally, in Sec. IV we shall summarize and draw conclusions. The physical system under investigation is a twodimensional nanowire with side-coupled quantum dots. The considered infinitely long wire allows electrons to propagate along the wire direction while being confined transversely. The considered mesoscopic system can be viewed as three divided regions, namely, the left source lead, the scattering region including the side-coupled dots, and the right drain lead, as is depicted in Fig. 1. The Hamiltonian describing the system can be written as H = H0 + Vsc x,y , 1 where Vsc is the scattering potential describing the sidecoupled quantum dots localized inside the scattering region. The unperturbed Hamiltonian is given by 1098-0121/2008/78 16 /165318 9 \u00a92008 The American Physical Society165318-1 H0 = \u2212 2 2m 2 + Vc y , 2 where m stands for the effective mass of the electron. The confining potential Vc y is assumed to have the parabolic form Vc y = 1 2 m 2y2, 3 leading to the discrete subband energy levels n = n + 1 2 4 with n=0,1 ,2 , " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001796_lapc.2012.6402949-Figure9-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001796_lapc.2012.6402949-Figure9-1.png", + "caption": "Figure 9. The simulated surface current distribution at 2.6 GHz (a) conventional circular patch antenna (b) proposed miniaturized antenna", + "texts": [ + " The plots show that, the reflection coefficient improves as the outer ring becomes thicker, and that the resonant frequency also increases. It is also noted that the bandwidth increases with the thickness of the outer ring. The resulting sensitivity of S11 as a function of frequency to both the central circle radius and the thickness of the outer ring are shown in Figure 8. Over the range 2.6 GHz to 2.8 GHz, it can be seen that the change in the thickness of the outer ring has a bigger influence on the resonant frequency of the antenna structure compared to the central circle radius. Figure 9 compares the simulated surface current distribution of the proposed antenna with its conventional circular patch counterpart. It is noted that the circular ring slot forces a longer current path along the outer radius of the slot. Furthermore, at 2.6 GHz, a strong surface current is observed in this region. However, only a weak current is observed in the inner circle. The simulated radiation pattern of the proposed antenna at 2.6 GHz in the x-z plane (\u00d8 = 0\u00b0) and y-z plane (\u00d8 = 90\u00b0) is plotted in Figure 10" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000718_bmei.2010.5639976-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000718_bmei.2010.5639976-Figure8-1.png", + "caption": "Figure 8. Steering force anlysis.", + "texts": [ + " 7(b) will work better when rounding an obstacle near the target. III. BASIC FORCE ANALYSIS When a bevel-tip needle is used, the steering force is generated by its interacting with tissues of human body. Therefore, the angle \u03b8 of the tip is the first important factor impacting the steering force. Assume the interacting force between a bevel-tip and tissues is F , which can be divided into bF and rF . bF is the force to bend the needle, and rF is the force to resist the insertion of the needle. As shown in Fig. 8, the relationship among F , bF and rF are cos sin b r F F F F \u03b8 \u03b8 =\u23a7 \u23a8 =\u23a9 . (1) It is shown that smaller tip angle \u03b8 generates larger bending force. The value of \u03b8 is determined by both needle bending requirement and needle\u2019s using purpose. While inserting, bending in the first articulation happens firstly as it is more flexible than the head. After the head is bent an angle I\u03b2 , as shown in Fig. 9, the thrusting force IF can be divided a force IbF and a force ItF on the first articulation: sin cos Ib I I It I I F F F F \u03b2 \u03b2 =\u23a7 \u23a8 =\u23a9 " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002426_tvcg.2007.1033-Figure12-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002426_tvcg.2007.1033-Figure12-1.png", + "caption": "Fig. 12. There is also a well-defined mapping from contact forces to joint torques. For a static grasp, these joint torques must match exactly those generated by muscle activation (Fig. 11).", + "texts": [ + " First, the joint torques generated by the tendons due to muscle activation can be computed as follows (Fig. 11): J \u00bcMPa; \u00f010\u00de where P is diag\u00f0p1;max; p2;max; . . . ; pn;max\u00de, where pi;max is the maximum force that can be generated along tendon i. Matrix M contains joint moment arm information and converts tendon forces to joint torques. Parameter a is an n 1 vector of activation levels, ranging from 0 (inactive) to 1 (at maximum force), for n tendons. Then, when the hand grasps an object, we can also map from the contact forces to the joint torques as follows (Fig. 12): 0J \u00bc JTf; \u00f011\u00de where f is a 3m 1 vector of contact forces with m as the number of contacts, and JT is the contact Jacobian, which maps contact forces to joint torques. Finally, we can map the contact forces to wrenches on the object using the grasp matrix G (Fig. 13): W \u00bc Gf; \u00f012\u00de G \u00bc I3 I3 . . . I3 Ro1 Ro2 . . . Rom ; \u00f013\u00de Rok \u00bc 0 rokz roky rokz 0 rokx roky rokx 0 0 @ 1 A; \u00f014\u00de where m is the number of contacts and rok \u00bc \u00bdrokx roky rokz T is the grasp map moment arm for the kth contact. To take both the hand grasp ability and the task requirement into account, we consider the following grasp quality metric: Q \u00bc min i kwi;maxk ktik ; i \u00bc 1; " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003985_s10846-013-9874-y-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003985_s10846-013-9874-y-Figure4-1.png", + "caption": "Fig. 4 Model of a micro manipulator with IPMC micro gripper in ADAMS software", + "texts": [ + " For obtaining the end effector velocity from joint variables in each direction, the combined effector velocity matrix is written as \u23a1 \u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a3 Vx Vy Vz \u03c9x \u03c9y \u03c9z \u23a4 \u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a6 = \u23a1 \u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a3 \u2212d3S\u03b81S\u03b82+a1S\u03b81 d3C\u03b81C\u03b82 0 d3C\u03b81S\u03b82+a1C\u03b81 d3S\u03b81C\u03b82 0 0 d3C2\u03b81S\u03b81+d3S2\u03b81S\u03b82 0 0 S\u03b81 C\u03b81S\u03b82 0 \u2212C\u03b81 S\u03b81S\u03b82 1 0 \u2212C\u03b82 \u23a4 \u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a6 \u00d7 \u23a1 \u23a2\u23a3 \u03b8\u03071 \u03b8\u03072 V3 \u23a4 \u23a5\u23a6 (4) Where \u03b81 is joint angle from x0 and x1 about z0 axis; \u03b82 is joint angle from x1 and x2 about z1 axis; \u03b8\u03071 is the first joint rotational speed; \u03b8\u03072 is the second joint rotational speed; V3 is velocity; d3 is the distances along the z axis from joint 2 to joint 3; a1 is short linear distance along x axis; Vx, Vy and Vz are velocities of manipulator along their respective axis; \u03c9x, \u03c9y and \u03c9z are rotational/angular speed of manipulator along their respective axis; S\u03b81 is Sin\u03b81 and C\u03b81 is Cos\u03b81; S\u03b82 is Sin\u03b82 and C\u03b82 is Cos\u03b82. For solving these mechanisms, the basic model of micro manipulator is developed in ADAMS software as shown in Fig. 4. During developing the model, each shaft has diameter 20 mm and length 30 mm. They are integrated with fixed bracket. The rotation of first shaft mechanism is given by mounting the motor with a revolute joint and second shaft mechanism is linked with first shaft mechanism in a perpendicular direction. The size and shape of both the mechanisms are identical. The rotation of second shaft mechanism is also provided by mounting another motor with revolute joint. The lead screw mechanism is constructed by a slider for linear motion which is restricted upto 10 mm because the distance between one hole to next hole is 10 mm" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000800_941748-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000800_941748-Figure6-1.png", + "caption": "FIGURE 6. ATypical Optimization Problem.", + "texts": [ + " Next, the students are introduced to 3D analysis with problems such as that shown in Figure 5 [4]. These types of problems can be easily entered into the FEA software without the need for advanced 3D modeling. In order to introduce the students to non-structural problems, they are given examples in thermal stress analysis and heat transfer. These are followed by an assignment to determine the thermal stresses in a bi-metallic joint. The students are introduced to shape optimization for weight reduction in a simple structural problem such as that shown in Figure 6 [5]. For their final project, the students are given a more open-ended design and optimization project. In the most recent semester, students were asked to design and optimize an A-arm from an automobile suspension system. The objective was to minimize the weight of the arm given the overall dimensions and loading conditions. An picture of one of the submitted solutions is shown in Figure 7. OBSTACLES One of the major problems associated with using advanced software in course work is the additional time involved with software training", + " Although separate books can be found that cover concepts in Computer Graphics [6] and Finite Element Method [7], no single book covers the concepts that are required to comprehend the concepts and practice them at the undergraduate level. Perhaps the most critical obstacle in teaching with design software is the large amount of computer equipment necessary to support its use. Not only are powerful workstations required to run the software, but the amount of disk storage required to allow 40 or so students to build, analyze and possibly optimize a part or assembly is quite daunting. For example, to run a design optimization using the CAD/FEA interface module for the example problem shown in Figure 6 required well over twenty megabytes of disk storage. While this problem can be partially addressed by allowing students to use temporary disk space on the individual machines, the storage requirements for saving the part models alone becomes extensive over the course of a semester. CONCLUDING REMARKS carefully used in the classroom to empower students to explore more alternatives and design possibilities and to tackle problems once thought out of reach in a teaching environment. This can help in closing the gap between the sorts of problems that students encounter during their formal education and those that they are faced with in their careers" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001929_20.120032-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001929_20.120032-Figure2-1.png", + "caption": "Fig. 2. Rectilinear model of the basic rotor design.", + "texts": [ + " Because of this saliency, the magnetic field in the air gap is somewhat more complex than for a uniform air gap machine and detailed analysis requires complete solutions of the field. However, reasonably accurate results can be obtained also with analytical techniques through various simplifying assumptions. CLd. Manuscript received June 24, 1991. Fig. 1. Cross section of a initial design permanent magnet rotor. 0018-9464/92$03.00 Q 1992 IEEE 111. ANALYSIS A. Analyrical Approach We start our analysis from the simplified two-dimensional and rectilinear model of the rotor cross section as shown in Fig. 2. Because of periodicity, the model covers only one half of the pole pitch i.e. 90 electrical degrees. This model does not take into account end-effects nor nonhomogeneous air gap surfaces. Furthermore, the materials are assumed to be linear, allowing us to determine the magnetic fields produced by different sources separately. The permeability of the iron is infinite. The quantities 1, and 1- denote the effective air gap lengths over the iron pole and over the magnet, respectively. By introducing these effective air gap lengths, the impact of stator slotting can be taken into account 151. The mmf wave produced by the permanent magnet, S,, rotates in synchronism with the stator field at angular speed os and acts along the d-axis. However, the fundamental component of the armature mmf, Sa,, acts along both d- and q-axes depending on the torque angle 6 between the stator and rotor magnetic axes: s, = Pmcos(o,t) (1) s,, - P,COS (e-o,t-i) (2) where 0 is the angular coordinate, cf. Fig. 2. The maximum value of the mmf wave in (1) is - HJ. P* - - 2 (3) where H, is the coercivity and 1, is the length of the permanent magnet. The peak value fia is given by (4) where k, denotes $e winding factor, Nph series turns of the winding per phase, 1. the peak armature phase current and p the pole number. Dividing the armature mmf wave into two components acting along the two symmetry axes, we are able to analyze two separate fields, i.e. the direct-axis and quadrature-axis field. The general expression for the reluctance is where \\ is the length and 4 the area of the flux path" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001283_iemdc.2013.6556163-Figure11-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001283_iemdc.2013.6556163-Figure11-1.png", + "caption": "Figure 11. Rotor lamination of the PMto house a copper cage. Version 1 has o version 2 has also the red bars", + "texts": [ + " EXPERIMENTAL RESULT To validate the conclusion at section III, transient FEA are experimentally validated on prototype, whose rotor drawing is reported motor, originally equipped with permanent m intended for traction, is the one used in ratings are reported in Table II. ors have slightly LS1 motor is at the efficiency factor is lower, us 5.0 A). efficiency grows expected due to Ampere ratio at little decrease in longer designed aintains a good ity significantly. MPLE MACHINES AT LS3 Inductio n motor 398 398 4,95 5,0 14.2 15,1 1500 1381 2231 2183 0,745 0,794 50 50 318 330 48 168 31 31 0,834 0,790 S the results of the a LSSyR motor in Fig.11. This agnets (PM) and [9], whose main E MOTOR PRIOR TO SSYR copper bars have been inserted inste then welded at the hands to copper the cross section of the rotor lamina the copper has been inserted. The te two stages. At first, the cage is only of each pole, as represented by the g Later on, the red copper bars have b layer, the smallest one. Fig. 12 repo prototype, in versions 1 and 2, befor the small red bars. Fig. 13 exhibits the equipment us of the LSSyR prototype has been me rotor by means of a known force ap The inertia of Version 1 is Jmot = 0. been assumed to have the same iner little impact when more rotating bo are involved. Figure 13. LSSyR motor prototype load ad into the saliencies, and end rings. Fig. 11 shows tions and the areas where sts have been divided into in the three bigger layers rey copper bars in Fig. 11. een added into the fourth rts the picture of the rotor e and after the insertion of assisted SyR of [9], modified nly the copper bars in grey, in the outer layers. ental tests. a) Verion 1; b) ed for the tests. The inertia asured by accelerating the plied with a known lever. 0057 kgm2. Version 2 has tia, which is wrong but of dies other than the motor ed by the DC machine. The load torque is produced by a DC se AXEM series [10]. Its inertia is 0.0067 kgm2 of LSSyR prototype, DC machine and coupli kgm2 (2,2 times the one of motor alone)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001238_ias.1998.732255-Figure17-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001238_ias.1998.732255-Figure17-1.png", + "caption": "Fig. 17. Flux linkages vs. current components in the simplzed model of saturated IPM motor", + "texts": [ + " Non Linear Case A model with constant values for h, and ld (as model illustrated in Section VII, point C) is particularly advantageous for FW analysis and, in addition, for design purpose. In fact, it allows simple equations to be used to describe the motor performance. In the circle diagram, the centre of the voltage limit ellipses is constant with the load, then the maximum speed is easily determined. Moreover, if the mutual coupling is neglected, variations of both curves, ellipses and hyperbolae, are only on the iq-axis. flux-linkages shown in Fig. 17 and tiescribed by Aq(iq) =L,iq iq I, can be adopted. With this simple model only five: FE field solutions are needed to characterise the motor. The first three solutions, at very low current, allow Am, Ld, Lq to be estimated. They are Am=O.O7Vs and, referring to Figs. 14- 15, L~0.29mH, Lq=0.59mH. Two more field :;elutions, with high q-axis current components, allow I\\ps and L, to be determined. If &=O. 12Vs at ic250A and &=o. 14Vs at iq=350A (see Fig.15) are imposed itwice in the second of (17), one can achieve A,=O" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002362_amm.364.285-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002362_amm.364.285-Figure1-1.png", + "caption": "Fig. 1 The composition of the stamping processing sets of on-line system", + "texts": [ + "110, University of British Columbia, Kelowna, Canada-14/07/15,02:06:26) performance of the mechanical structure, at the system level, so as to realize the optimal design of the system level [6-7] . The analysis of the stamping processing complete set of control line system The composition of the stamping processing sets of line system. The stamping processing complete sets of line system is composed of three presses, one feeding manipulator, three transfer manipulators, one double open buttress unit [8] . All devices are arranged in the form of parallel tandem as shown in figure 1, the original is a single manual production line, this line system can be used for different types, automatic industrial production line of different numbers of presses, improve the efficiency of the sheet metal forming and the safety and reliability of the production, at the same time, which can greatly reduce the number of operating personnel and labor costs. The workpieces is followed as the sequence of from left to right, as shown in figure 1, which the feeding mechanism is double open buttress unit can ensure the automatic line non-stop continuous operation when change another stock, thus greatly improve the production efficiency. The stamping processing sets of control line system has three processing station and two auxiliary stations (the reclaimer station and the discharge station). One of the core is that the total control system is the SSCNET field bus technology; the second core is that the feeding system is controlled by the double open buttress unit; the core of three is that the materials are controlled by the high-speed dynamic characteristics of the manipulator to transmit between machines, respectively, and next give a detailed analysis of each core" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002224_3527608117-Figure5.12-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002224_3527608117-Figure5.12-1.png", + "caption": "Fig. 5.12 Principle of DC plasma spraying (Linde AG).", + "texts": [ + " As for wire flame spraying, the application of filler wires extends the spectrum of use of the coating materials substantially. Arc spraying in controlled, inert or reactive environments has gained no industrial importance so far and single-wire arc spray pistols are also still in the laboratory stage. In the latter the arc burns between the wire and the nozzle wall, which functions as a permanent electrode. Aligned anodically poled wires can be processed with an extremely small widening angle of the spray stream. 5.2.6.6 Plasma Spraying 5.2.6.6.1 DC Plasma Spraying In DC plasma spraying (Fig. 5.12) an arc is ignited between a water-cooled copper anode, designed as a ring nozzle, and a similarly water-cooled, pin-shaped tungsten cathode. The gases flowing between both electrodes (Ar, He, N2, H2) are thereby dissociated, ionized and form a plasma jet, into which the spraying additive materials can be brought. The outstanding advantage of plasma spraying is the high plasma temperature of approximately 6000 to 15 000 K, which makes it possible to process materials with very high melting points" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001472_9781118354179.auto089-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001472_9781118354179.auto089-Figure4-1.png", + "caption": "Figure 4. Schematic diagram of the clutch assembly and control system. (Janic\u0301ijevic\u0301, Jankovic\u0301, and Todorovic\u0301, 1987. Reproduced by permission of Faculty of Mechanical Engineering, University of Belgrade.)", + "texts": [ + " Wet clutches and hydrodynamic couplings are different designs studied separately; \u2022 \u201cDisc type\u201d: Friction surfaces are \u201cflat rings.\u201d Cylindrical (drum type) and conical friction surfaces are nowadays very rare. \u2022 \u201cNormally\u201d engaged type: When there is no outside intervention, the clutch will be fully engaged, transmitting the torque. \u2022 Consists of two main sub-assemblies: \u2013 Clutch \u2013 Control/(De-) Actuating system A schematic diagram of the clutch assembly and control (de-actuating) system is shown in Figure 4 (Janic\u0301ijevic\u0301, Jankovic\u0301, and Todorovic\u0301, 1987), with the clutch assembly being highlighted within a dotted box. The diagram shows disengaged clutch [note pedal force Fp on the clutch pedal (12)], with the friction disc (1, also called friction plate Encyclopedia of Automotive Engineering, Online \u00a9 2014 John Wiley & Sons, Ltd. This article is \u00a9 2014 John Wiley & Sons, Ltd. DOI: 10.1002/9781118354179.auto089 Also published in the Encyclopedia of Automotive Engineering (print edition) ISBN: 978-0-470-97402-5 or ring) being axially free, as indicated by the f gap between the friction disc and pressure plate (2)", + " Depending on the gears engaged, the rotational speed will be related to the vehicle speed. If the gearbox is in the neutral position, and the clutch disengaged, the speed of the friction disc (1) and input shaft (9) will depend on their previous speeds, inertia of the components, friction and aerodynamic losses, and ultimately, any contacts between the friction disc (1) with the pressure plate (2) and/or the engine flywheel (8). It should be noted here that when clutch is fully disengaged, as shown in Figure 4, the gaps between the friction disc and the flywheel (8) and pressure plate (2) cannot be controlled. On disengagement, the pressure plate (2) will move away from the friction disc (1), creating a small gap typically, of the order of a millimeter. A drop in the force normal to the friction surface will result in drop of friction force. The overall design, the speed and distance of disengagement, axial stiffness of the friction disc, inertia forces and quality of the spline connection between the friction disc (1) and shaft (9), as well as thermal expansion and axial run-out of the friction surfaces of the friction disc, pressure plate, and flywheel will determine any contacts that may still exist between these components, despite the clutch being nominally disengaged", + " Such contacts will generate torque and heat, but for a well-designed clutch assembly in good condition, these friction effects will be negligible. No doubt, for such a complex mechanism rotating at high speeds, all components must be well designed, in terms of both stiffness and elasticity, adequately machined and connected. Component tolerances, both dimensional and geometrical, play vital role in successful design. Friction, axial, and centrifugal loading, thermal aspects and wear of friction surfaces represent additional requirements that must be taken into consideration when designing clutch assemblies. To engage the clutch (Figure 4), the force Fp on the clutch pedal (12) should be gradually reduced, which will reduce and eventually eliminate gaps f between the friction disc (1) and pressure plate (2), and between the friction disc (1) and flywheel (8). On full clutch engagement, a gap will be created between the thrust (axial) bearing (7) and disengagement lever (5), enabling full force of the (preloaded) spring/s (3) to be exerted to the pressure plate (2), which will press the friction disc (1) against flywheel (8). Friction forces will be developed on the contact (friction) surfaces between the friction disc (1) and flywheel (8) and between the friction disc (1) and pressure plate (2)", + " The friction forces will be proportional to the coefficient of friction and normal forces, hence maximum friction forces will be developed when the clutch is fully engaged. The returning spring (10) has a role in ensuring the clutch pedal (12), link (11), and lever (6) are returned in a normal operating position, and bearing (7) is no longer in contact with disengagement lever/s (5), and consequently not rotating. This is vital in order to ensure maximum friction force and minimize bearing wear. Figure 4 shows schematically the main components of a clutch assembly and control mechanism. It is indicated that the coil springs (3) positioned circumferentially between the cover and pressure plate provide normal force. In addition to such a design, normal force can be provided in a number of ways: using diaphragm spring, large central coil spring, centrifugal force, magnetic force, hydraulic pressure, and others. Some of the most commonly used and interesting solutions will be presented. Encyclopedia of Automotive Engineering, Online \u00a9 2014 John Wiley & Sons, Ltd. This article is \u00a9 2014 John Wiley & Sons, Ltd. DOI: 10.1002/9781118354179.auto089 Also published in the Encyclopedia of Automotive Engineering (print edition) ISBN: 978-0-470-97402-5 Figure 5 (Nunney, 1998) shows the most common design used today on road vehicles, a diaphragm spring clutch assembly. Nominally, all clutch assembly components shown in Figure 4 are present; however, the spring is of diaphragm type, not requiring separate disengagement levers [marked (5) in Figure 4]. For this role, the diaphragm spring has slots toward the inner diameter, which take on the role of the levers, enabling axial movement (disengagement) of the pressure plate. Fulcrum rings (Figure 5) provide \u201cpivots\u201d for relative movement (flexing) of the spring in relation to the clutch cover, minimizing spring contact stresses, and ensuring high mechanical efficiency. The rivets connect the spring, via the wire rings to the cover. For torque transfer, separate flexible straps are positioned in the circumferential direction and attached to the pressure plate and clutch cover", + " In such designs, both friction discs and pressure plates are manufactured from special carbon fiber materials, able to withstand high temperatures, providing high coefficient of friction values. These assemblies are extremely light and have low rotary moment of inertia but high costs. Typically, lives of such components are limited and measured in a number of races, with the race start being the most severe operating duty. Fundamental layout, components, and operation of a mechanical clutch control system are shown in Figure 4. The schematic diagram shown in Figure 22 further illustrates operation, giving lever ratios of each component: clutch pedal (ip), transfer mechanism (im), bearing fork (ib), and clutch lever (ic). As indicated, in order for the clutch to be disengaged, the product of the pedal force Fp, all lever ratios and overall mechanical efficiency \u03b7 Encyclopedia of Automotive Engineering, Online \u00a9 2014 John Wiley & Sons, Ltd. This article is \u00a9 2014 John Wiley & Sons, Ltd. DOI: 10.1002/9781118354179.auto089 Also published in the Encyclopedia of Automotive Engineering (print edition) ISBN: 978-0-470-97402-5 of the control system and clutch itself must be higher than maximum normal force FN" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001826_870305-Figure17-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001826_870305-Figure17-1.png", + "caption": "Fig. 17", + "texts": [], + "surrounding_texts": [ + "870305\n9\nTrends: See chart Also higher degree of integration: Formation of large pre-assembled units, consisting of instrument panel. heater or air conditioner, steering, pedal~, fuse-and-relay board.\nHaterial: Trends:\nPP, ashtray: FS 31 (phenoplast, saw-dust-filled) - Multi-color injection-molded - Textile decor - Softer surface", + "10\no Heater and air conditiorer\n870305\nHaterials:\nTrends:\n- Fousing parts: PP TF - Blower wheel: PON - Thin-layer technique for - Flaps hard/soft\n- Flilps: - Controls:\nhousing parts\nNetal/PU foam ABS", + "870305\n11\nHaterials:\nTrends:\n- Housing parts: PP TF - Blower wheel: POM - Thin-layer technique for - Flaps hard/soft\n- Flaps: - Controls:\nhousing parts\nHetal/PU foam ABS" + ] + }, + { + "image_filename": "designv6_24_0002279_s1003-6326(15)63931-0-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002279_s1003-6326(15)63931-0-Figure2-1.png", + "caption": "Fig. 2 2D axisymmetric geometrical model (a) and mesh design (b) of FE model", + "texts": [ + " In the extrusion process, the extrusion pad, which is usually made from low alloy steel with lower deformation resistance, deforms when the pipe is completely extruded. Thus, the whole extruded pipe can be pushed out by the deformed extrusion pad. This extrusion process can improve not only the production efficiency, but also the material utilization ratio. In current study, considering the symmetric geometry of the billet and extruded pipe, a 2D axisymmetric FE model for the process was established. Because of the same moving speed between extrusion needle and extrusion pad in the extrusion process, the ram shown in Fig. 2(a) was used to describe the geometries of the extrusion pad and the extrusion needle. On the other hand, the extrusion die and the container were fixed and contacted with each other in a practical extrusion process. Thus, the die shown in Fig. 2(a) was used to represent the geometries of the container and bottom die. In addition, in order to balance the simulation accuracy and efficiency, the mesh refinement is necessary in the main deformed region (see Fig. 2(b)), and the mesh design method was described and verified in Ref. [15], which is reasonable and acceptable. As mentioned above, the DRX is a key microstructure evolution mechanism of Inconel 625 alloy in the hot deformation process. In Ref. [6], the DRX evolution behavior of this alloy in the hot deformation process was studied through the isothermal compression experiments, and the dimensions of the experimental samples were d8 mm \u00d7 12 mm. At the same time, the DRX evolution models were also developed, which are expressed by Eqs" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003701_s13239-016-0278-6-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003701_s13239-016-0278-6-Figure3-1.png", + "caption": "FIGURE 3. Experimental (a) Setup (b) Schematic.", + "texts": [ + " The experiment is conducted primarily to determine constants K1, K2 and K3 of Eqs. (3) and (5) for cylindrical device and also to get feel for movement of device under different conditions. The experiment is conducted with sponge in glass tube, depicting sponge as device and glass tube as vessel. The aim is to find out and model minimum value of differential air pressure across sponge initiating its movement in glass tube by fitting constants K1, K2 and K3 of Eq. (3) and (5). The setup shown in Fig. 3 is used. The air from air compressor is allowed to pass through glass tube which has tightly fitted sponge in it. The diameter of glass tube is 1.1 cm with length as 36 cm. The sponges of diameter 1.1 cm and lengths of 2.5 , 5 , 7.5 , 10 , 12.5 and 15 cm are considered. The air is passed through sponge and the differential air pressure at which movement of sponge starts is noted down. This experiment is used to fit constants K1 and K2 of Eq. (3). The differential air pressure required initiating movement of sponge with diameter 1" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002295_b978-0-08-011216-9.50010-9-Figure7.8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002295_b978-0-08-011216-9.50010-9-Figure7.8-1.png", + "caption": "FIG. 7.8", + "texts": [ + " Find the maximum torque which the shaft can transmit in addition to carrying the bending moment if the allowable direct stress is not to exceed 5 ton/in2 and the maximum shear stress 3J ton/in2. Which of these two limiting stresses is reached? Give the maximum value of the other stress. 2. A mild steel shaft of 2 in. diameter is subjected to a bending moment of 17,280 in. lb. If the yield point of the steel in simple tension is 30,000 lb/in2, find the maximum torque that can also be applied according to (a) the maximum principal stress, (b) the maximum shear stress, (c) the shear strain energy theories of yielding. 3. A pulley shaft is supported in bearings as shown in Fig. 7.8. Pulley A receives 60 h.p. at 250 rev/min through a vertical belt and power is trans mitted from pulley B by a horizontal belt. The ratio of the belt tensions is Fi/F2 = 3. Determine a minimum diameter for the shaft using the shear strain energy criterion of yielding. Yield stress in simple tension for the shaft material is 18 ton/in2. 4. A pipe is formed into a U-shape, the straight portions being 3 ft long and the semi-circular end being of 1 ft radius. One end of the pipe is clamped so that the U lies in a horizontal plane" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001310_6.1995-1532-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001310_6.1995-1532-Figure8-1.png", + "caption": "Fig. 8 Towing sequence", + "texts": [ + " Two systems to rotate the EAP in horizontal position are currently studled: either a plug is installed by dlvers into the nozzle, and air is injected to expel1 the water from the 20 Amaican m u t e of Aermaldics rod Astrcaautics motor,or a special buoy is injected and inflated shows the towing configuration with the two EAP in the motor for the same purpose. behind the vessel, which bring the booster to Once the first booster in an Kourou harbour in nearly 80 hours. horizontal position, the vessel moves to the second EAP and conducts the same operations. Figure 8 Program history The Booster Recovery System development has been dwided into two phases: - phave 1 (1989 to 1992): system studes have been conducted to preQct EAP reentry conditions using wind-tunnel test results, and the system preliminary design has been acquired. At the end of phase 1, element specification ( parachutes, structures, control box) were issued. Two parachutes companies were associated to the program during phase 1, - phase 2 ( 1993 to 1995): at the beginning of phase 2, call for tenders have been organised to award hardware development contracts" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003289_s00170-016-9067-5-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003289_s00170-016-9067-5-Figure1-1.png", + "caption": "Fig. 1 a Schematic illustration of FTPW process. b Schematic configuration of FTPW joint", + "texts": [ + " Thermal process . Cooling time\u0394t8/5 . Microstructure . Mechanical behavior Friction welding processes are widely used to join and repair structures in many sectors [1\u20135]. Friction taper plug welding (FTPW) is one of the best innovative friction welding processes, and it is capable of joining dissimilar materials applied in the transportation, energy, and offshore industries [6]. FTPW involves forcing a highly rotating taper plug co-axially into a drilled blind hole, and its schematic is illustrated in Fig. 1a. Schematic configuration of FTPW joint is shown in Fig. 1b. This welding method has great benefits due to solid-state welding: no defects related to melting, automatic process offering remote control, and weld quality independent of operators [7]. It is a fast process alternative to traditional arc welding processes. When it comes to weld thick-walled material, the advantage of FTPW is more obvious compared to conventional thick-walled repair techniques which involve expensive consumables and lengthy processing [8]. Moreover, it is not necessary to use flux or shielding gas to make this process more environmentally friendly [9]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000232_18.335900-Figure9-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000232_18.335900-Figure9-1.png", + "caption": "Fig. 9. The volume of the four-dimensional truncated polydisc O,(t; R ) is proportional to the shaded areas.", + "texts": [ + " By proper selection of the size of the constituent 2D constellation, virtually any point on the curve governing the tradeoff between shaping gain and CER2, or PAR2 may be obtained. APPENDIX VOLUME AND AVERAGE ENERGY OF THE TRUNCATED POLYDISK 1) Volume Calculation: Set N = 2n. The volume of the truncated polydisc can be written as r: -k ... -kri I tR2 where the elementary volume element has been taken as a product of area elements dA, = d(vr,?). This transforms the problem of integrating over a complicated 2n-dimensional region into integrating over a portion of an n-cube bounded by hyperplanes. For example, Fig. 9 shows the region of integration when n = 2. Integrating over this region is itself quite complicated, especially for large n. However, the problem may be simplified by borrowing a tool from probability theory. Observe that the region of integration is proportional to the probability that the sum of n independent random variables, each uniformly distributed on [0, R2], is less than tR2. Let Z = X , be a sum of n independent random variables, each uniformly distributed on [0,1], and let c#Jn(t) P[Z I t] be the cumulative distribution function (cdf) for Z" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000483_1.c035581-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000483_1.c035581-Figure4-1.png", + "caption": "Fig. 4 Configuration of the DEP STOL demonstrator.", + "texts": [ + " In addition, the twin fuselage configuration has the advantages of increasing the payload capacity and the location options for equipment installation [25], which is conductive to a rational arrangement of the numerous items of equipment and batteries in DEP aircraft. However, some issues of the twin fuselage configuration remain, such as interference between fuselages, increases drag; and the structural mass increases due to the addition of a fuselage. In light of these design requirements and the considerations stated earlier in this paper, the twin fuselage tandemwing configurationwas used for the demonstrator after making the necessary tradeoffs, as shown in Fig. 4. D ow nl oa de d by B R U N E L U N IV E R SI T Y o n Ju ne 2 , 2 02 0 | h ttp :// ar c. ai aa .o rg | D O I: 1 0. 25 14 /1 .C 03 55 81 Twenty-four EDFs were used as thrusters for the demonstrator. To control the thrusters conveniently, six EDFs were grouped together, and therewere four groups. The trailing-edge portion of thewingwith the EDFs mounted was connected to the wing through a shaft, and it tilted as the shaft rotated, driven by the servo system mounted in the wing. The EDFs, together with the trailing-edge portion of the wing with the EDFs mounted, were tilted between horizontal and vertical configurations to accomplish horizontal flight and STOL" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000087_msf.685.278-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000087_msf.685.278-Figure1-1.png", + "caption": "Fig. 1 Model of flaking by RCF.[3]", + "texts": [ + "[7] Then, the crack initiation site associated with pre-existing defects in specimen interior is produced from not only the particles but also Stage I cracks. The whole size of the initiation site involving Stage I cracks is larger than that of the defect itself. A comparatively larger pre-existing defect in the specimen interior may introduce higher stress concentration than the surface flaws. Especially at lower cyclic stress, the difference is considered to become more distinct, since the extrusion-intrusion mechanism becomes less active. Developed model of flaking by RCF was proposed as shown in Fig. 1. [3] Firstly, local damage is produced in the stress concentration region of matrix near inclusion. The initial crack is generated there, grows in the horizontal direction and finally results in flaking. Based on the linear mechanics relationship between stress intensity factor and critical defect size, stress intensity factor range, \u2206KII\u2019 was defined as follows: \u2206KII ' = 2\u03c4 0 \u03c0a . (1) where 2a is diameter of inclusion or defect and t0 is shear stress amplitude in the direction parallel to track at the depth of flaking" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001247_j.ijleo.2013.09.021-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001247_j.ijleo.2013.09.021-Figure1-1.png", + "caption": "Fig. 1. The structure of Cassegrain system. (1) Laser, (2) pre-collimation lenses system, (3) primary mirror and (4) secondary mirror.", + "texts": [ + " Obviously, the optical antenna is the most important compoent of the optical system. In this paper, a Double pyramidal system in Cassegrain system as been designed, meanwhile an optical system software has been et up to analyze the performances of the traditional Cassegrain ntenna and the performances of this optical antenna. . Structure for optical antenna The Cassegrain system consists of two reflecting surfaces, a conave parabolic main dish and a convex hyperbolic secondary dish hich is shown in Fig. 1. The wavelength of laser is 1550 m, by the optical design softare Ze-MAX, we have designed a pre-collimation optical lenses ystem, which is shown in Fig. 2. It includes two aspheric cylinder enses, which are perpendicular with each other for the generator f the cylinder lens. \u2217 Corresponding author. E-mail address: yanghj@uestc.edu.cn (H. Yang). 030-4026/$ \u2013 see front matter. Crown Copyright \u00a9 2013 Published by Elsevier GmbH. A ttp://dx.doi.org/10.1016/j.ijleo.2013.09.021 3. Analysis of Cassegrain optical antenna 3" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001122_29.45557-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001122_29.45557-Figure7-1.png", + "caption": "Fig. 7 . Experimental system.", + "texts": [ + " Therefore, a plot of the evaluation index in relation to the threshold level is characterized by a single peak at which the threshold level is optimum. Hence, at first we set the threshold level at which the evaluation index is highest. At every subsequent chip position recognition step, the evaluation index on either side of the previous threshold level is examined and the threshold level with the higher index value is selected. Thus, a VP with almost the same binary representation as that of the RP can be obtained. IV. EXPERIMENTS A. Experimental System A system used for the experimental evaluation of the proposed algorithm is shown in Fig. 7. A semiconductor chip is placed on a stage with X-, Y-, and &direction servomechanisms. The chip is illuminated from an oblique direction, magnified using a microscope, and projected on the imaging surface of a TV camera and suitably quantized for the required positional resolution. The image signal thus obtained is then sent to a small-scale image processor [ 101. The position recognition algorithm described in the previous sections, including the matching procedure used for image thresholding, is executed mostly as hardware operations within the image processor" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003205_ecce.2014.6954109-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003205_ecce.2014.6954109-Figure1-1.png", + "caption": "Fig. 1: Machine thermal model with water cooling of stator jacket and forced-air convection cooling.", + "texts": [ + " The hysteresis core loss is evaluated using the apply-loop method, while the eddy current core loss is calculated by the Fast Fourier Transfer (FFT) approach [12]. Both of these core loss calculation techniques are built-in function blocks in JMAG. Each magnet is circumferentially segmented into 6 pieces, resulting in approx. 90% magnet loss reduction [13]. B. 3D Thermal FE Model A 3D FE model is adopted for the thermal analysis. The candidate machines are assumed to be mounted in water jackets with a coolant temperature of 65\u2070C. Both machine ends are air-cooled by an external blower with an average air velocity of 2 m/sec, as illustrated in Fig. 1. Conduction heat transfer is calculated by the FE thermal solver after the thermal conductivities for each of the machine components are specified. The thermal conductivity [Wm-1K-1] of the stator winding is estimated [14,15] by taking a weighted average of the insulation thermal conductivity kcu,ir and the copper thermal conductivity , according to (1), as follows: = \u2217 + 1 \u2212 \u2217 , (1) where is the fractional slot fill factor. The windings, magnets, stator iron, and rotor iron are assigned as the heat sources for the copper loss, magnet loss, stator core loss, and rotor core loss, respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001672_chicc.2018.8484049-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001672_chicc.2018.8484049-Figure1-1.png", + "caption": "Fig. 1: The figure caption", + "texts": [ + " In Section 3, the principle of the proposed PDTC algorithm is introduced and applied to the SRM. Section 4 presents the numerical simulation results. Section 5 contains conclusions. 2 Predictive Model SRM has a unique double-salient pole structure, and only the stator core is wound with excitation windings. SRM is operating according to the \u201dprinciple of minimum reluctance\u201d. The rotors continue to work by applying the excitation to power converter in a circular way. The structure and power converter of three-phase 12/8 SRM are shown in Fig. 1. Due to the severe nonlinearity of the SRM, the torque is a complex non-linear function of the stator winding inductance and rotor position, and it is difficult to describe with exact analytical expressions. Neglecting the influence of interphase coupling, the state equation of SRM can be simply described as follows:\u23a7\u23aa\u23aa\u23aa\u23a8 \u23aa\u23aa\u23aa\u23a9 d\u03c8p dt = up \u2212Rpip, p = 1, 2, 3 d\u03c9 dt = 1 J (Te \u2212 k\u03c9\u03c9 \u2212 TL), Te = 3\u2211 p=1 Tp d\u03b8 dt = \u03c9 (1) where \u03c8p, up, Rp, ip and \u03b8 are stator phase flux linkage, stator phase voltage, stator phase resistance, stator phase current and rotor position, respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001939_2015-01-0088-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001939_2015-01-0088-Figure8-1.png", + "caption": "Figure 8. Gear Shift Fork - Contact Pattern - Loads and Boundary Conditions", + "texts": [ + " Surface to surface contact is provided between mating surfaces of forks legs & third pad with synchronizer ring. Contact starts at fork legs that will have a hinge effect before it touches third pad meeting the required deflection. A minimum nominal gap is provided for 3rd pad to share the load when max/abuse load appears on the load shifting Jaw. Gap could be adjusted if the FOS in the legs goes below 1.0 for the Max/Abuse Load as shown in the Fig. 9 and Fig. 10. Loads and Boundary conditions as shown in the Fig. 8. Once the third pad come into contact stress pattern shifts from fork legs to the middle of the fork as shown in the Fig. 11. Stress induced should be below the Yield strength of the material (FOS > 1.0). If not, the web thickness will be increased to meet the requirement. Experimental Verification As per VE Commercial Vehicles Ltd, standard durability duty cycle, rig has been setup as shown in the Fig. 12 and tested the transmission assembly in which Aluminum Gear Shift Fork's (1st & Reverse Fork and 4th & 5th Fork) are the test components" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002366_012003-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002366_012003-Figure5-1.png", + "caption": "Figure 5. Model of 12Slot-14Pole SegSta PMFSM", + "texts": [ + "1088/1757-899X/917/1/012003 (15) Since the area of armature slot SA can be estimated by considering the general equation of a trapezoidal area as follow: ( ) (16) where b1 is the length of inner slot area, b2 is the length of outer slot area and h is the height of the slot area as follows: (17) (18) (19) (20) (21) Substituting Equations (17), (20) and (21) into (16) yields ( ) ( ) (22) With WA are considered as the armature, areas of SA can be simplified as follows: [ ] (23) Therefore, the complete initial design of three phase SegSta 12S-14P PMFSM with segmental stator model is illustrated in figure 5, while the design parameters are listed in Table 1. IOP Conf. Series: Materials Science and Engineering 917 (2020) 012003 IOP Publishing doi:10.1088/1757-899X/917/1/012003 In order to validate the operating principle of the proposed design based on U-phase, V-phase, and Wphase, 2-D finite element analysis (FEA) has been performed to analyse the electromagnetic behavior. Moreover, it is significant to analyse the magnetic flux linkage and flux characteristic of the propose design since it determine the output performance of the electric motor" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002448_2011-01-0862-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002448_2011-01-0862-Figure3-1.png", + "caption": "Figure 3. Package layouts for different engine configuration: (a) In-line - cylinder; (b) In-line 3- cylinder; (c) Flat 2-cylinder; (d) Vee 2-cylinder.", + "texts": [ + " NVH takes on particular significance within an E-REV due to the fact that the engine is not operating for large periods of time, and should not be noticeable over the electric-only operation. This requirement is pertinent for the gas exchange system layout of a RE engine. Table 1 shows the list of reciprocating engine concepts that were considered as part of the assessment. Parametric CAD models were created that allowed a package volume comparison to be made and a cost / weight model was also generated from benchmark engine and component data. The results of this analysis, compared to a baseline in-line 3- cylinder engine, are also shown in Table 1. Figure 3 shows package layout models for the different engine configurations. The methods used to assess these alternative layouts have a significant effect on the selection, where possible objective processes were used. The concept level CAD models allowed package space to be determined for each engine concept. These models were compared by overall dimensions and also the box volume occupied by the RE unit. This way all solutions could be fairly assessed even if one dimension of the box was unusually large and limited by basic engine geometry" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002388_2015-01-0090-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002388_2015-01-0090-Figure4-1.png", + "caption": "Figure 4. Locations of effective lumped masses representing various main structural components.", + "texts": [ + " In this paper, the simulations were conducted to predict strains generated on the frame under the circumstance of the truck being driven up the ramps. Steady state responses of generated strains when the truck rested on the top of the ramp were considered. The followings were the assumptions used in simulation environment: 1. Main structural components including a passenger cab, a rear cargo, an engine and a gear box were simplified as lumped masses [10] located at each respective Centre of Gravity (CG) (Figure 4). Those mass were 250 kg, 400 kg, 116 kg, and 56 kg respectively. A gravitational acceleration of 9.8066 m/s2 was applied to the model. Since the effects of carrying loads were not studied in this paper, the representation of the payload was not included in the model. 2. Axles and leaf springs of Hotchkiss suspension were considered as rigid beams (Figure 5). Beam assumption could be used to simplify suspension components because the main focus was on a vehicle structure [11]. 3. Semi elliptical leaf springs were approximated as front and rear cantilevers connected to each other" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001895_iembs.1998.745920-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001895_iembs.1998.745920-Figure1-1.png", + "caption": "Figure 1: Laser Doppler Velocimetry Measurement Grid", + "texts": [ + " After collection, data from each measurement location were discretized into 20 ms time bins, forty three of which made up the pulsatile cycle. The ensemble averaging of the measurements within the bins enabled the calculation of mean and fluctuating components of velocity, principle turbulent normal stresses, ad maximum turbulent shear stress (MTSS). The coopdinate rotation method used to attain these stresses is descn\u2019bed in detail by Fontaine et al[4]. ent Gr& Velocities were measured over a grid encompassing one q~aaer of the flow chamber cross section. This grid was located 1 mm downstream of the fully open valve leaflets (Figure 1). Measurement sites were spaced 1 to 1.5 mm apart within the grid to allow detailed spatial characterization of the pulsatile flow fields. Condltlons: . . The fluid used in this loop was a solution of (by volume) 79% SaMated sodium iodide, 20% glycerin, ad 1% water. This solution simulated the viscosity of blood at high shear rates (3.5 cSt) and the rehtive index of the acrylic flow chambers. Matching the kinematic viscosity of blood enabled the imposition of dynamic similarity between the flow within the model and the flow within a human aorta" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001181_aps.2007.4396027-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001181_aps.2007.4396027-Figure2-1.png", + "caption": "Fig. 2. Antenna Placement with Phantom Hand Fig. 3. Antenna Gains in dB", + "texts": [ + " Actual antennas were measured using an SMA connector and a ground plane of copper clad FR4 measuring 450 mm by 535 mm. The hand is modeled as homogenous layers of dielectric material. Concentric cylinders of bone, muscle and skin are used to simulate the fingers of the hand. The palm of the hand is similarly constructed of rectangular boxes. The front plane of the hand extends vertically from the ground and is placed 12 mm from the closest edge of the antenna. Hand models and their orientation with respect to the antennas are shown in figure 2. Results Without the presence of a hand, the antennas share a number of properties. The far field electric fields are primarily in the Theta direction and provide omni directional polar patterns in the horizontal plane. Gain patterns with and without the hand model are shown in figure 3. At low elevation angles, \u03b8 <15 degrees, the helical antenna exhibits the effects of a pattern null. However, the TBH antenna demonstrates improved gain at this angle as the pattern becomes dominated by the antenna\u2019s vertical loops" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000587_0029-5493(65)90101-9-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000587_0029-5493(65)90101-9-Figure4-1.png", + "caption": "Fig. 4. Geometrical relations.", + "texts": [ + " APPLICATION OF THE THEORIES TO INTERSECTING SHELLS In o r d e r to be able to make use of the y ie ld loci for i n t e r s ec t i ng she l l s as r e p r e s e n t e d in sec t ion 2, some approx ima t ions to the shel l conf igura t ion mus t be in t roduced. Although the fo l - lowing a rgumen t holds t rue for any she l l of r e v o - lution, we r e s t r i c t o u r s e l v e s to the spec ia l c a se of a r a d i a l out let f rom a s p h e r i c a l she l l subjec t to in te rna l p r e s s u r e as shown in fig. 4. F o r the c y l i n d r i c a l p a r t of the she l l , the equat ions of equ i l ib r ium, with the notat ion of f ig. 5, a r e given by: dQ NO dMx + - - r = P ' dx - Q ' N x = \u00bdpr \" (10) F o r N o = const the in teg ra t ion of eq. (10) y i e ld s to (11) M x = \u00bd( , - ~ - ) x ' 2 + A x , B . The constants A and B can be evaluated using any suitable stress distribution, provided this assumed stress field nowhere exceeds any of the chosen yield surfaces of section 2, e.g. / a t x = l , Q = 0 , M x = M c , N o = N c , (12) a t x : 0 , Q : Q ' , M x : M ' c , NO : N c " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003027_imece2014-38788-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003027_imece2014-38788-Figure7-1.png", + "caption": "Figure 7: Solid model of the compliant mechanism designed in example 1b", + "texts": [], + "surrounding_texts": [ + "Specifying appropriate energy/torque for a mechanism at various precision positions is cumbersome. For simplicity, a heuristic judgment may be made between the energy/torque specifications and the rotation of the pseudo-rigid-body links of the compliant mechanism, to ensure the specifications are appropriate. On various occasions, however, a designer may still need assistance with providing appropriate specifications. The above presented optimization formulation guides the designer with this. The function value at the end of the optimization process is an excellent indicator of the energy or force/torque specifications. If the function value is not close to zero, then some iteration must be conducted by changing the initial estimates drastically. This will ensure a search for the global minimum. If the function value at these various starting positions is still not close to zero, then an unrealistic problem definition may exist. In this instance, the following steps should be conducted to better understand the change in direction: 1. Determine whether or not the energy/torque at various positions is in agreement with the rotation of the pseudo-rigid-body links. 2. If the result from Step 1 is deemed satisfactory, then the user should either increase or decrease the energy/torque specifications. 3. Examine the function value at the end of Step 2. If the function value is approaching zero, then continue in the same direction until the desired function value is achieved. In case the function value is diverging further, change the direction and repeat Step 3. The above process is illustrated in examples presented in the following section." + ] + }, + { + "image_filename": "designv6_24_0002148_kem.620.318-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002148_kem.620.318-Figure5-1.png", + "caption": "Fig. 5 Pressure distribution on the brake shoe", + "texts": [ + " 6, we can obtain: the brake shoe upper pressure q2=0.09112KN/mm, the lower pressure q1=0.0353KN/mm. Similarly, the pressure of the brake shoe ends in wheel backward revolving can be gotten: the brake shoe upper pressure q1=0.0406KN/mm, the lower pressure q2=0.08621KN/mm. From the above results, we can see that when the wheel forward, the upper pressure is 2.58 times of the lower and the lower pressure is 2.58 times of the upper when the wheel backward. The actual proportion of the pressure distribution acts on the brake shoe shown in Fig. 5. (1) If the design is applied that the wheel and the brake shoe are concentric, during the train braking, the pressure on the brake shoe from the wheel is not averagely distributed because of the friction generated by the friction from the wheel tread. (2) When the wheel is forward revolving ,the pressure on the upper end of the brake shoe is larger than that on the lower end, otherwise, the upper end will be smaller than the lower end. (3)The ratio of the pressures between the upper end and the lower end, in wheel forward revolving is larger than that in wheel backward revolving" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000029_j.iheduc.2009.12.004-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000029_j.iheduc.2009.12.004-Figure3-1.png", + "caption": "Fig. 3. Users perceive the forces (direction and magnitude) while pushing the cube.", + "texts": [ + " Students are required to fall back to the textbook and 2D illustrations. \u2022 Control over a continuous (large) range of physical parameters is adjustable by the user. Such a fine resolution cannot be achieved in a real experiment because of limited human mechanical ability. Motivated by these limitations we designed and implemented an environment that simulates the force of friction and the associated paradigms. Students use the haptic device to manipulate a cube on an inclined plane and receive force feedback from the device (Fig. 3). Students may apply varying amounts of force and directly receive varying resultant forces from the cube. They can also change the values that affect frictional force, such as the mass of the cube, the coefficients of static and kinetic friction, and the slope of the plane along which the cube moves. In addition to overcoming the limitations of the traditional approach, the visuo-haptic simulation provides several other benefits, such as: \u2022 Affordability. Low-cost haptic devices that are connected to the existing computers in the school laboratories" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000735_tmech.2004.828657-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000735_tmech.2004.828657-Figure4-1.png", + "caption": "Fig. 4. Schemes of surface tension for the cubic gripper/sphere interaction (left) and for the needle gripper/sphere interaction (right).", + "texts": [ + " The analytical form of the electrostatic force is difficult to be derived and integrated for a real finger tip, both with a needle and a cubic shape. In order to overcome this limitation, three-dimensional (3-D) electrostatic simulations have been performed by using the MEMS developing tool CoventorWare\u2122 2001. In the adhesion phenomena, the effect of humidity (i.e., the surface tension [34]) is expressed by (5) where and are the characteristic radii of the external cylindrical surface of the meniscus bridge (Fig. 4), and is the surface tension of water. The first term of (5) is the capillary contribute to the total adhesion force, where the amount in brackets is the inverse of the so called Kelvin radius, . The Kelvin radius is an intrinsic propriety of meniscus and it depends only on physical parameters (6) where is the Boltzmann constant, and HR are, respectively, the temperature, the molecular volume and the relative humidity of air. In our case, with a relative humidity of 60%, the Kelvin radius is about 2 nm. Fig. 4 shows that is larger in the case of the needle grasping tool, thus increases for the cubic grasping tool, in order to keep the Kelvin radius constant. Then, by considering (5) we observe that must be larger for the cubic tool rather than for the needle tool. The diagram of the surface tension reported in Fig. 5 confirms this conclusion. The substrate adhesion forces have been calculated with the assumption of a sphere, halfspace interaction, on the basis of the contact theory described in Section III-A" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002244_isape.2010.5696431-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002244_isape.2010.5696431-Figure3-1.png", + "caption": "Fig. 3. A planar SW antenna and its feed network.", + "texts": [ + " THE SLOW-WAVE (SW) ANTENNA TO CIRCUMVENT ANTENNA GAIN BANDWIDTH LIMITATION The broadband miniaturized slow-wave (SW) antenna is another application of the present theory of broadbanding or size reduction technique for the TW antenna [9]. Although this technology has been conceived and patented for a decade, its significance and application are more recent, driven largely by the growing broadband requirements in wireless systems and the increasing availability of low-cost high-quality and workable ceramic substrates. Fig. 3 shows a photograph for a planar SW antenna and its feed network with an SMA connector for input/output. The SW antenna consists of a two-arm spiral, I inch in diameter, above a square ceramic substrate with high dielectric constant. Underneath the substrate is the feed network in a thin metallic casing, which is rather bulky but can be reduced in size. The SW structure is the region shown in Fig. 2 which is loaded with one or more layers of dielectric substrates. Fig. 4 shows measured VSWR of the planar SW antenna in Fig. 3. As can be seen, over a wide frequency range of 1-5 GHz, the VSWR is generally < 3 and often < 2. Note here that, without the SW technique, the planar spiral needs a diameter of > 4 in. in order to provide the radiation zone shown in Fig. 2 for mode-I operation at I GHz. Therefore, its I-in diameter is only v.. of that required for ordinary mode-I TW antenna operation; this is a significant reduction in antenna size. Let us examine the theory of this problem by noting first that the phase velocity, V, of a wave is given by V=fA (5) Since the presence of a dielectric substrate reduces the wavelength A, and the frequency would remain constant regardless of the SW structure, the phase velocity of the TW is reduced or slowed and transformed into a slow wave. The slowing of the phase velocity is gauged by a \"slow-wave factor\" or \"SW factor\" given by SW Factor= VJVs=AoIAs (6) where Vo and v., are the phase velocity in free space and the SW structure, respectively; and Ao and As are the wavelengths in free space and the SW structure, respectively. With the ceramic substrate in Fig. 3, phase velocity Vs of the magnetic current M in Eqs. 2 and 3 is reduced by a SW factor about 4. As a result, the TW is brought into coherence between adjacent arms at a radiation zone with a diameter v.. of the case without using the SW technique. Several other SW antenna models were also designed and studied. In addition to impedance matching, their radiation performance is studied. Generally gain enhancement of 10 to 30 dB was observed at low frequencies for which the antenna is electrically small, as exemplified in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003944_20.619602-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003944_20.619602-Figure2-1.png", + "caption": "Fig. 2. Structure of the measured LDMs (units in mm).", + "texts": [ + "00 0 1997 IEEE I I Where, I is the coil current in A, @ is the magnetic flux by coil in Wb, Em(= Zm/pA) is the magnetic resistance of the magnetic circuit in H - l , 1, is the length of the magnetic path in m, p is the magnetic permeability in H/m, and A is the sectional area of the magnetic path in m2 Therefore, the relation between the inductance and the magnetic circuit is as follows; L = N ~ / R ~ The inductance is proportional to the square of the number of the coil windings and in inversely proportional to the magnetic resistance. B. Measurement of Coil Impedance It is difficult to calculate the magnetic resistance of LDM, because of the flux leakage. The factor which shapes the moving coil of inductance is clarified by the measurement of various coils. Fig. 2 shows the structure of measured LDMs. Table I shows those specifications. Variables are the length of coils in stroke direction I,, the height of coils h, and the thickness of the center yoke t,. The length of the yokes in stroke direction and the number of the coil windings are constant. The coil was k e d , the impedance is measured with an impedance analyzer. Fig. 3 shows the relation of variables to the resistance component and the inductance component. They are values of the resistance component and the inductance component when the size 50 40 30 20 10 0 0 2 4 6 8 e, CL: Thickness o f c e n t e r yoke t,(mm) (a) Influence of thickness of center yoke tc e m 0 10 20 30 40 a l c E H Height of c o i l hc (mm) (b) Influence of height of moving coil h," + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001397_icitacee.2016.7892413-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001397_icitacee.2016.7892413-Figure1-1.png", + "caption": "Fig. 1 The force working on the teeth", + "texts": [ + " The resolution of the disc rotation can be measured by counting the number of pulses. Direction of rotation can be determined by knowing which the channel leading to others because both the channel will always be different quarter-turn phase (quadrature signal). Often there is a third channel output, called INDEX, which generates one pulse per revolution is useful for calculating the amount of rotation that occurs. The gear is a gear-shaped machine element that serves as transition rotary motion and power of the engine components to one another as shown in Figure 1. Efficiency approaching 978-1-5090-0890-2/16/$31.00 c\u00a92016 IEEE 68 98% so that the gear is widely used to make the drive transmission to the shaft. There are several factors that determine measures on gears can be show in Table 1 as follows: Linear ball screw actuator is a mechanical equipment which uses the rotational motion of the moving objects with minimal friction. Screw on the shaft serves as groove ball bearing so that the removal can be done with precision / high-precision position that can be shown in Figure 2" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002010_jae-2008-999-Figure20-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002010_jae-2008-999-Figure20-1.png", + "caption": "Fig. 20. Prototyped flexible PCB for an electric source transmission.", + "texts": [], + "surrounding_texts": [ + "Many researchers have been proposed the various kinds of analysis method for the piezoelectric motor considering the contact mechanism. However, the exact analysis has been impossible up to now. Also, there was no arranged design method for the piezoelectric motor. The analysis and design of the piezoelectric motor was using the trial-error-method or using the analysis of vibrator without the consideration of the contact condition. The former is inefficient in the aspect of time and cost and the latter cannot estimate motor characteristic at all. To address these problems, we suggested the analysis and design method of the USM using analytic method combined with numerical method, which was named as the CM in this research. By the proposed analysis and design method, the USM can be designed systematically considering the speed of motor. To verify the suggested analysis and design method and for the demand of small actuators in industry, the 8.5[mm] outer diameter RUSM was developed in this research. Using the prototyped RUSM, it was verified that the proposed analysis and design method was correct by comparing its outcomes with the experimental data. But, the nonlinearity resulting from the heat generation and from the high mechanical force which is applied to the motor was not considered in the proposed analysis method. Consequently, further investigation is required to identify and solve the problems related to the nonlinearity of the USM. For the vibrator of the USM, various kinds of comb-tooth structures were prototyped. Hence, we found out the relation between the speed of the motor and the number of teeth for the first time. It has the significant meaning in the sense that the tooth design method has not been suggested yet. Also, the experiment data showing the relation between the speed and the number of teeth validated the suggested CM. To sum up, the exact analysis of the USM has been impossible and the arranged design method for the USM has not been suggested till now. Hence, it is remarkable that the USM can be analyzed and designed systematically by the proposed method, while considering the contact condition in case of no mechanical force is applied to the motor ideally. It is also important to note that the analysis and design of many other kinds of machines, which use similar mechanism, is possible with the suggested method. Appendix The piezoelectric material is a z-axis poling. The material coefficients are: Mechanical stiffness matrix for constant electric field E: cE = 13.25 6.94 6.46 0.0 0.0 0.0 6.94 13.25 6.46 0.0 0.0 0.0 6.46 6.46 10.52 0.0 0.0 0.0 0.0 0.0 0.0 2.68 0.0 0.0 0.0 0.0 0.0 0.0 2.68 0.0 0.0 0.0 0.0 0.0 0.0 3.16 \u00d7 1010[N/m2] Piezoelectric matrix: e = 0.0 0.0 0.0 0.0 12.82 0.0 0.0 0.0 0.0 12.82 0.0 0.0 \u22126.61 \u22126.61 13.5 0.0 0.0 0.0 [C/m2] Permittivity matrix for constant mechanical strain S: \u03b5S = 7.32 0.0 0.0 0.0 7.32 0.0 0.0 0.0 6.28 \u00d7 10\u22129[F/m2] Density: \u03c1 = 500[kg/m3] Mechanical quality factor: Q = 900. The elastic body of the vibrator is made of the phosphor-bronze. The material coefficients are: Mechanical stiffness matrix: cm = 179.75 96.79 96.79 0.0 0.0 0.0 96.79 179.75 96.79 0.0 0.0 0.0 96.79 96.79 179.75 0.0 0.0 0.0 0.0 0.0 0.0 41.481 0.0 0.0 0.0 0.0 0.0 0.0 41.481 0.0 0.0 0.0 0.0 0.0 0.0 41.481 \u00d7 109[N/m2] Density: \u03c1m = 780[kg/m3] Mechanical quality factor: Qm = 3000." + ] + }, + { + "image_filename": "designv6_24_0003515_tmtt.2017.2650912-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003515_tmtt.2017.2650912-Figure2-1.png", + "caption": "Fig. 2. Fully inkjet-printed multilayer transformer. (a) Top view of sample. (b) 3-D stack-up [16].", + "texts": [ + " As such, the ability to additively manufacture such elements with high performance and quality is an essential requirement toward empowering the emergence of fully printed IoT motes. To a great extent, the SOTA in this area shows a great deal of maturity. Indeed, inkjet printing, for instance, allowing the deposition of both conductors and dielectrics, has been used for the fabrication of high quality RF lumped elements, such as metal\u2013insulator\u2013metal (MIM) capacitors with self-resonancefrequencies (SRFs) above 1 GHz [15], multilayer inductors with SRF above 1 GHz, as well as multilayer printed transformers [16] (shown in Fig. 2). In addition, the high (and increasingly so) quality demonstrated by printed conductors has been applied to the fabrication of high performance antennas with operation frequencies up into the mm-wave range [17]. AMTs, along with advances in materials science, are likewise opening the door for the fabrication of flexible devices, through the integration of strechable conductors, as exemplified by the stretchable encoding module shown in Fig. 3. Furthermore, the unique properties of AMTs have also been used to empower the birth of unique elements, such as drill-less vias [18] and physically reconfigurable components (also known as \u201cOrigami\u201d) [19], which could provide optimal solutions for dynamic and moving IoT platforms of the future, such as unmanned aerial vehicles or intelligent cars" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001939_2015-01-0088-Figure10-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001939_2015-01-0088-Figure10-1.png", + "caption": "Figure 10. Gear Shift Fork - Contact Patch Displacement Results", + "texts": [ + " Contact analysis has been carried out to check the contact status using FE Ansys Solver. Surface to surface contact is provided between mating surfaces of forks legs & third pad with synchronizer ring. Contact starts at fork legs that will have a hinge effect before it touches third pad meeting the required deflection. A minimum nominal gap is provided for 3rd pad to share the load when max/abuse load appears on the load shifting Jaw. Gap could be adjusted if the FOS in the legs goes below 1.0 for the Max/Abuse Load as shown in the Fig. 9 and Fig. 10. Loads and Boundary conditions as shown in the Fig. 8. Once the third pad come into contact stress pattern shifts from fork legs to the middle of the fork as shown in the Fig. 11. Stress induced should be below the Yield strength of the material (FOS > 1.0). If not, the web thickness will be increased to meet the requirement. Experimental Verification As per VE Commercial Vehicles Ltd, standard durability duty cycle, rig has been setup as shown in the Fig. 12 and tested the transmission assembly in which Aluminum Gear Shift Fork's (1st & Reverse Fork and 4th & 5th Fork) are the test components" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001472_9781118354179.auto089-Figure14-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001472_9781118354179.auto089-Figure14-1.png", + "caption": "Figure 14. Power flow in the diaphragm spring clutch. (Reproduced with permission from ZF Friedrichshafen AG.)", + "texts": [ + " Such solution can work well even with manual transmission, with the clutch and motor functions complementing each other and reducing vehicle fuel consumption and emissions. This is the most commonly used design and the presentation here will be concentrating on the typical assembly and component designs and materials. Figure 5 shows a cross section of the typical diaphragm clutch design, with all components of the assembly. Figure 13a show a photograph (from ZF Sachs) from the \u201cengine side,\u201d including the friction disc. Figure 13b shows \u201cgearbox side\u201d view. Figure 14 (from ZF Sachs) shows the power flow\u2014when the clutch is engaged and disruption to power flow when the clutch is disengaged. For the disengaged clutch, it can be clearly seen that all components of the clutch assembly rotate with the engine flywheel, with the exception of the friction disc. The main clutch components shown in Figure 5 are for the \u201cpush-type\u201d clutch. This is the most common type, but it Encyclopedia of Automotive Engineering, Online \u00a9 2014 John Wiley & Sons, Ltd. This article is \u00a9 2014 John Wiley & Sons, Ltd" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002605_973233-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002605_973233-Figure1-1.png", + "caption": "Fig. 1. Schematic Diagram of a Rear Suspension", + "texts": [ + " The second enhancement allows MARS to produce a more robust design, which minimizes performance variation that results from both design parameter and environmental variation. Applications of MARS can be found in [6,7,10-121. Since IDESIGN is a gradient-based optimization program, the derivatives of the objective function and the constraints with respect to the optimization parameters are needed. These derivatives are essentially the derivatives of the performance indices, which can be calculated by the finite difference method. A SUSPENSION PROBLEM In this example the rear suspension system shown in Fig. 1 was synthesized using optimization techniques. Two cases were studied : one to design the geometry, and the other to develop the stiffnesses of the rubber bushings. The bushings are employed to reduce the vibration generated from the road surface. PERFORMANCE INDEX CALCULATIONS - The performance indices for this suspension are listed in Table 1. These 13 indices are called SDF (Static Design Factors). A description of these factors can be found in [13]. SNAC\"~', an in-house computer program of General Motors', was used to calculate these indices based on given suspension geometry and bushing rates" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003274_ijvd.2012.047403-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003274_ijvd.2012.047403-Figure3-1.png", + "caption": "Figure 3 Detail of PMSM MA-55, modelled for simulation and used for in-wheel traction systems", + "texts": [ + " This expression is defined as follows: m e L m rt t dF J dt \u03c9\u03c9\u2212 = + (8) and finally the electromagnetic torque: 3 ) 2 (e sd sq sq sdt i ip \u03c6 \u03c6\u2212= (9) The inverter frequency is related to the rotor speed as: e m p \u03c9\u03c9 = (10) Finally, the behaviour of a PMSM could be modelled, by equations (4)\u2013(8) and equation (10), and using the variables and parameters of Table 1, with the next five expressions (Weizheng et al., 2009): )( sq s sq e sd sq sqd L i u R i dt \u03c9 \u03c6+= + (11) ( ) sd s sd e sq sd sd fd L i u R i dt \u03c9 \u03c6 \u03c6\u2212 + = + (12) 3 ) 2 (( )e sd sq sd sq f sqt i ip L L i\u03c6\u2212 += (13) e m p \u03c9\u03c9 = (14) m e L m rt t dF J dt \u03c9\u03c9\u2212 = + (15) These equations could be easily modelled in a simulation environment as, for example, MATLAB/SIMULINK. In this paper, a MA-55 INFRANOR PMSM, Figure 3, has been used for the simulation tests with the parameters1 shown in Table 1. Open loop control of a synchronous motor with variable frequency can be satisfactory at variable speed when the motor works with stable values of torque and without many requirements on speed. When the drive specifications require fast dynamic response and high accuracy in speed or torque control, the open loop control does not offer this possibility. This is the reason to operate the motor with closed loop control, where the operation dynamic drive system plays a fundamental role as an indicator of the system, which takes part (D\u00edaz and Pardo, 2004)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001856_eeng.2016.7845985-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001856_eeng.2016.7845985-Figure1-1.png", + "caption": "Fig. 1: Profile of 6/4 & 6/10 & 6/8 SRM", + "texts": [ + " In [6], a comparison between 8/6, 8/10 and 8/14 SRM which is designed using the above PD formula is made to prove the torque characteristic improvement. This paper analyzes an in-wheel SRM with different number of rotor poles for the application to electric vehicles. The analysis selects 6/4, 6/8 and 6/10 three phase SRM for comparison which is mainly focused on torque, torque density and torque per loss. Also, the analysis can be extended to four phase in-wheel motor to examine the torque performance. The size of these motors is based on the in-wheel motor in [7]. The common motor parameter and profile of the three motors is shown in Figure 1 and Table I respectively: 6/4 For 6/4 SRM, because of the large conducting angle, the stator and rotor pole need to be thicker to prevent high saturation and set as 35 degree and 37 degree respectively. For 6/8 SRM, it is better to have larger stator pole angle [4] and the stator and rotor pole arc angle are set as 22 degree and 21 degree. Similarly, for 6/10 SRM, the above two angles are 18 degree and 16 degree respectively. Due to higher conducting angle for 6/4 SRM, the thickness of rotor back iron has to be larger and set as 40mm, while the other two is set as 32mm" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003388_euma.2000.338648-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003388_euma.2000.338648-Figure1-1.png", + "caption": "Fig. 1. Cylindrical cavity filter design", + "texts": [ + " The derivatives of the matrix A(k) are obtained using the automatic differentiation technique [5-6] which provides derivatives that are sufficiently accurate in the predefined precision conditions. The derivation the eigen system (2) leads to a iterative linear system which allow to calculate the nth derivatives using the previous ones. IV. A. Cylindrical cavity filter The following test case has been chosen to demonstrate the great flexibility and capacity of the pole expansion method to model different microwave structures. The considered structure is a cylindrical cavity filter whose dimensions are shown in Fig. 1. The E-plan, H-plan and port symmetries have been used for studying the circuit. thirty modes are needed to be calculated to evaluate the scattering parameters of the structure. Fig. 2 shows the pole expansion results compared to a commercial EM solver calculating the performances directly using the EM simulation for each frequency point. A very good agreement between the two results is noted. Also, Fig. 3 shows the modes positions and their weights (ie: Cin defined in (4)). It can be seen from this figure that the presence of the two modes positioned at 15" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002190_iemdc.2011.5994637-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002190_iemdc.2011.5994637-Figure1-1.png", + "caption": "Fig. 1. Generator base design", + "texts": [ + " Finite Element Method Magnetics (FEMM) is used to investigate in this paper. FEMM is a simulation program that can be used to perform low frequency electromagnetic problems on two-dimensional planar and axisymmetric domains. It address the issue of linear/nonlinear magneto-static problems, linear/nonlinear time harmonic magnetic problems, linear electrostatic problems, and steady-state heat flow problems [5]. A. Generator Design The generator design used here is based on a generic design shown in Fig. 1. The generator basic layout is drawn using LUA scripting language which is supported by the FEMM software to provide an easier access towards the changing and control of the number of stator slots In this design only the stator slots and stator slot tooth does not have a fixed width and spacing. All other fixed components are shown in the Table 1 and is used throughout the simulation to ensure that there are no extra variables that would affect the outcome. 978-1-4577-0061-3/11/$26.00 \u00a92011 IEEE 453 The materials used in this generator are also kept constant throughout the simulation to provide a consistent output" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002797_dscc2011-6167-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002797_dscc2011-6167-Figure4-1.png", + "caption": "Figure 4. Flexible Tube Actuation System", + "texts": [ + " Both rigid inner tube and flexible outer one are driven inside left ventricle by the thrust force produced by DC motor 2. DC motor 3 enables the catheter to rotate around its long axis, which is important in steering prosthetic valve toward aorta. DC motors 1, 2 and 3 are attached to the catheter and rotate with it around the catheter\u2019s long axis. This allows the rollers to stay tight on inner and outer tubes and prevent them to slide over the catheter\u2019s surface. Finally, DC motor 4 is the one that distinguishes new design from the previous one. Figure 4 shows inside the system actuation box. As depicted in Figure 4, motor torque is directly transmitted to the roller via motor shaft that slides the flexible outer tube over the inner one. Also, to tighten the roller to the catheter to avoid slippage, there is another lower roller coupled to the upper one by means of two gears installed on the wall of actuation box. The whole transmission system is installed inside a 50\u00d750\u00d725 mm box. Although introducing the flexibility in the mechanism enables us to benefit from a number of advantages highlighted in previous section, it leads to undesired consequences as well" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002591_wac.2002.1049455-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002591_wac.2002.1049455-Figure6-1.png", + "caption": "Figure 6. Fiber-optic ultra ininature (125 pm outer diameter) pressure sensor.", + "texts": [ + " It can bend to one-direction using one SMA actuator which was fabricated using photolithography and electroche~nical etching from a TiNi SMA sheet. The active guide wire is veiy flexible because it consists oftlie meandering SMA actuator and the stainless coil spring. The guide wire bends to over 60 degree (length of bending mechanism is 51iiin) by applying current of 50mA in the air at 25OC. RELATED SENSOR SYSTEMS Fiber-optic Pressure Sensor Small diameter (125pin) fiber-optic pressure sensor shown in Figure 6 has been developed for catheter and guide wire [7]. A micromachined thin diaphragm is formed at the end of an optical fiber and the deformation of the diaphiagm by the pressure is detected optically. This fiber-optic sensor is free to electrical hazard and its signal is not affected by an electromagnetic interference. A forward-looking ultrasonic imager has been developed for catheter use (Figure 7). Piezoelectric PZT ceramic transducer array was made at the end of the catheter. Improved 1-3 composite piezoelectric transducer, built-in integrated circuit and micro relay for multiplexing the drive pulse are being studied as integrated ultrasonic imager at the end of catheter [8,9] " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000700_apcap.2014.6992573-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000700_apcap.2014.6992573-Figure2-1.png", + "caption": "Fig. 2. The predict radiation pattern", + "texts": [], + "surrounding_texts": [ + "Keywords\u2014planar magnetic-electric dipole; complementary antenna; unidirectional patterns\nI. INTRODUCTION\nIn recent years, with the development of wireless communication technology, there has greatly increased the demand for broadband antennas, which also requires antennas with a stable unidirectional radiation performance to increase the anti-interference ability. Therefore, there has been more attention to these antennas with broadband and unidirectional radiation pattern. As a kind of widely used antennas, dipole antenna has a good characteristic with wide impedance bandwidth, but a bad performance in directivity. It is well known that an electric dipole has figure-8 shape in the E-plane and figure-O shape in the H-plane, while the magnetic dipole has figure- O shape in the E-plane and figure-8 shape in the Hplane. When the electric dipole and the magnetic dipole are orthogonal placed with each other and excited simultaneously with appropriate amplitude and phase, the complementary antenna could achieve a unidirectional radiation performance and identical radiation patterns in the E-plane and the H-plane [1]. On the basis of this theory, a variety of electric dipole and a magnetic dipole complementary implementation structure have been proposed in the past decades, such as a novel slot and a monopole antenna, a slot and a invert L antenna, etc [2-4]. Although these design could achieve excellent unidirectional radiation pattern over the operating bandwidth, their structures are complex, and their working band are very narrow.\nBased on these research, professor K.M.Luk designs a new design of wideband magnetic-electric dipole [5]. A wideband short-circuited patch antenna is selected as the magnetic dipole, a planar dipole antenna is chosen as the electric dipole as well, which are both excited by an \u0413-shaped probe feed. By\nadjusting the parameters of the \u0413-shaped probe feed, the input impedance of the antenna can get matched well. Under the reflection of the metal ground plane, the antenna can achieve more than 40% of the impedance bandwidth, low crosspolarization, low back radiation, symmetric radiation pattern and stable gain over the operating band can also be obtained. The drawback of this design is that the dimensions of the ground plane is as large as the operation wavelength, and the profile of the antenna is a quarter of wavelength high.\nElectrically small antenna composed of a planar electrically small electric and magnetic dipoles which are based on electric and magnetic near-field resonant parasitic (NFRP) elements is introduced by Peng Jin[6]. A dipole augmented with symmetric top and bottom loadings acts as a electric dipole, while the two half-moon-shaped capacitively loaded loops operates like a magnetic dipole. The design is a three-metal, two-dielectriclayer structure, the electric element lies on the outside of one dielectric layer, and the magnetic element lies on the outside of the other layer. The directly driven printed electric dipole element and its CPS feedline constitute the middle metal layer. This design has a high radiation efficiency, which directivities patterns approach the Huygens source. Its dimensions is less than a quarter of wavelength, however, the structure is very narrow in bandwidth, and is unable to meet the real needs.\nThis paper proposes a complementary planar magnetic dipole antenna. A folded dipole is selected as the electric dipole [7], periodical capacitive loaded loop antenna can be equivalent to the magnetic dipole[8]. The two element are placed in the front and back of the Teflon substrate (\u03b5r=2.65, tan\u03b4=0.002, h=0.8mm) respectively. This structure is very simple and planar, and does not need large metallic reflectors. The proposed antenna could realize a good performance in directivity with symmetrical radiation patterns in the two polarization plane, and could achieve good Characteristics of gain and front-to-back ratio.\nII. ANTENNA DESCRIPTION AND DESIGN GEOMETRY\nThe geometry of the proposed antenna is shown in Fig. 1. The periodical capacitive loaded loop antenna is printed on one side of the Teflon substrate, which is feed by the parallel stripline connected to it. The folded dipole is printed on the opposite side of the substrate, which is excited through coupled feeding. The geometry parameters optimized by HFSS is shown in Table. 1. The operation frequency is 1.4GHz.\n978-1-4799-4354-8/14/$31.00 \u00a92014 IEEE Harbin, CHINA", + "According to the concept of the mu-zero resonance antenna [8], periodical series capacitive loading is connected to a loop antenna, which can achieve a omnidirectional radiation pattern in the far field. The loop is periodic symmetric, and the periodical series capacitive loading is realized by adding interlaced coupling lines at the end of each section. These capacitive loading can provide a very small phase correction between the adjacent sections, so that the current flowing along the loop maintain identical phase and uniform, therefore, although the loop antenna is no longer electrically small, it can still achieve a omnidirectional radiation pattern, which is identical to that of a magnetic dipole. In addition, the periodical series capacitive loading is also helpful to achieve a wide impedance bandwidth. The input impedance of the antenna can\nbe tuned by adjusting the length of the parallel stripline, which can make the antenna easy to be matched.\nThe folded dipole consists of two parallel half-wave dipoles with the tail ends of them be connected. As the two conductors are very close to each other, there is a very strong mutual coupling between them, thus the currents along the two conductors are identical. The folded dipole has the characteristic that is equivalent to the tuning stub, so it can effectively compensate the input impedance of the antenna with the variation of frequency. By selecting appropriate dimension of the antenna, the reactance of the input impedance can be adjusted close to zero, therefore, the operation bandwidth of the folded antenna more wider than that of the novel half-wave dipole.\nAccording to Huygen\u2019s source theory, the electric dipole has figure-8 shape in the E-plane and figure-O shape in the Hplane, while the magnetic dipole has figure- O shape in the Eplane and figure-8 shape in the H-plane. When the two dipole are orthogonal placed with each other and excited simultaneously with appropriate amplitude and phase, this design can form a complementary antenna. The omnidirectional radiation patterns of the two dipole would be superposed in the far field, so the complementary antenna could form a symmetrical cardioid radiation pattern in the Eplane and the H-plane, a desirable front-to-back ratio and unidirectional pattern can be achieved as well. In this paper, the periodical capacitive loaded loop antenna works as the magnetic dipole, and the folded dipole works as the electric dipole. The periodical capacitive loaded loop antenna is feeding directly by feed cable, and the folded dipole is feeding indirectly by coupling parallel stripline. This method can make their exciting phase difference close to 90\u3002, so the radiation pattern can achieve a good complementary result.", + "III. ANALYSIS\nThe proposed antenna is simulated by Ansoft\u2019s HFSS, the simulation results of the gain, impedance bandwidth, radiation pattern are obtained and shown below.\nThe simulated S11 of the magnetic-electric dipole antenna is shown in Fig. 3. It is clear that the VSWR\u22642 (S11<-10dB) impedance bandwidth of proposed antenna is measured as large as 500MHz from 1.3GHz to 1.8GHz(35.7%), the proposed antenna has a broadband performance. The simulated gain of the antenna shown in Fig. 4 is about 3.3dBi to 5.3dBi, a desirable gain performance is achieved across the operating bandwidth as well.\nMore attention is paid to that whether the radiation patterns are stable over the whole impedance bandwidth with identical patterns in the E-plane and H-plane. The simulated radiation patterns at 1.3, 1.4, 1.5, 1.6, 1.7, 1.8GHz, are shown respectively in Fig. 5, and the front-to-back ratio is given in Fig. 6. It can be seen that, the antenna can obtain unidirectional radiation patterns in most of the impedance frequencies, the radiation patterns in the E-plane and H-plane are stable and nearly symmetric. The direction of maximum radiation in the H-plane is slightly offset of the x-axis when the frequency is higher than 1.6GHz. The front-to-back ratio is over 10dB in both E-plane and H-plane in the low frequencies(1.3GHz~1.6GHz), but less than 10dB in the high frequencies(higher than 1.6GHz) , as a result of the radiation direction offset, which gives an identical variation tendency to that of the radiation pattern. The reason of the characteristics deterioration is that the current along the loop antenna can\u2019t maintain an identical phase and amplitude as the frequency increase." + ] + }, + { + "image_filename": "designv6_24_0000552_6.1992-3673-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000552_6.1992-3673-Figure1-1.png", + "caption": "FIGURE 1. The 2D Scramiet Vehicle", + "texts": [ + " Three alternative modeling techniques for the problem, of varying simplicity and accuracy, ax presented and discussed. Quality of approximation with the different analysis techniques is studied and compared to firstorder Taylor series approximations. Design optimization results with the different approximation techniques are used to demonstrate the viability of local-global approximations in CFD, and, finally, areas of difficulty and areas requiring further research are identified. .d W The Design Problem A complete 2D scramjet vehicle configuration is to be designed and is shown in Fig. 1. The net thrust is selected as the objective function. Geometric design variables define the front end and nozzle of the vehicle. The design variables for the nozzle are the initial nozzle turning angle, a, and a measure of wall curvature (Fig. 2). Alternative geometric parameters can serve as design variables defining the curvature. Either the nozzle exit height or il direct v curvature parameter can be used. Design variables for the front end are lengths of the nose and ramp. the vertical locations of their cnrners and the fore-aft position nf the cowl lip (Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001505_1.2738131-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001505_1.2738131-Figure4-1.png", + "caption": "Fig. 4 AISI 1018 cold-drawn steel cup after lapping and polishing: \u201ea\u2026 reports the large contact area cup \u201enominally 75.40 mm2 \u2026 and \u201eb\u2026 the small contact area cup \u201enominally 2", + "texts": [ + " Each friction pair is composed of a uZn30 brass disk and a cup-shaped 1018 steel piece. The inter- ace of contact is a small annular region corresponding to the rim f the steel cup. All the experiment use brass disks with the same eometry see Fig. 2 ; to ensure that the contact is established well way from the edge of the disk, the diameter of the brass disks is 1.75 mm. In order to investigate the specimen size effect on the friction henomena, steel cups with two different contact areas were used see Fig. 3 for drawing and Fig. 4 for pictures of physical amples : one case has a nominal contact area of 75.40 mm2 Fig. a and the other of 13.35 mm2 Fig. 4 b . 2.2 Sample Preparation. In order to investigate the grain ize effect, experiments were performed with brass samples havng two different grain sizes: the set of specimen with small grain ize has an average grain size of 32 m obtained by heat treat- ent at 550\u00b0C for 1 h and then cooled in firm air at 25\u00b0C , ournal of Manufacturing Science and Engineering ded 01 May 2009 to 129.105.215.213. Redistribution subject to ASM AUGUST 2007, Vol. 129 / 679 E license or copyright; see http://www.asme.org/terms/Terms_Use" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003223_003754977803100504-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003223_003754977803100504-Figure1-1.png", + "caption": "Figure 1 - Simplified valve schematic. The numbers 1, 2, 3 refer to the pipe carrying the fluid, the valve disc, and the relief outlet to atmospheric pressure, respectively.", + "texts": [ + " Numerical results for the transient response of selected system variables were obtained from the nonlinear model using CSMP-III7 on an IBM 370/158 computer. Eigenvalues and frequency response of linearized system models were generated from a FORTRAN version of the CSMPIII model used as a subroutine in general-purpose analytical programs.4,11 Using a load module, the typical computer execution time for simulating the transients for a period of 1 second was about 5 CPU seconds on an IBM 370/158 (loading time excluded). NOMENCLATURE DESCRIPTION AND OPERATING PRINCIPLE The relief valve considered in this analysis (see Figure 1) is typical of those used in process and power industries for liquid service. It is a continuous-action device which functions as a springloaded pressure regulator. As the driving force (fluid pressure) overcomes the spring force, the valve starts to allow the liquid to escape. Further increases in system pressure lift the valve disc from its seat, and the flow rate increases. The flow is diverted downwards by the cup-shaped disc, and its direction is reversed. This creates reaction forces that lift the disc further", + " (4) Changes in fluid density are negligible. (5) The pressure difference between valve opening and closing (hysteresis) is not significant. The following parameters were evaluated and found to be negligible: (1) Pressure-drop caused by gravity inside the valve (2) Frictional pressure drop in the vertical portion of the valve orifice (3) Fluid mass inside the valve compared to the mass of the moving parts. DEVELOPMENT OF MODEL EQUATIONS A valve schematic showing the fluid control volume and moving parts is given in Figure 1. Conservation of linear momentum6\u2019l0 in Cartesian tensor notation yields On a lumped average basis, the z-components in (1) for a homogeneous incompressible fluid can be approximated as where mass flow rate W = pAlul = pA2u2. The surface forces Sz in the z-direction can be split into two parts: the force exerted by the valve disc on the fluid, and the z-component Fz of the forces acting on the entire control surface except the valve disc. Substituting (3) in (2), and neglecting the term Aozop (because Aozop p \u00ab M) yields at RICE UNIV on June 11, 2014sim", + "3 Using these data in Equation 13, ~ is obtained as a single-valued function of a, which is substituted in (10) to obtain effective valve area A as a function of normalized valve lift. A (~) is approximated by a third-order polynomial in the range of interest. The coefficients a0\u2019 a1\u2019 a , and a should be evaluated from the manufacturer\u2019s data. Dynamic equations The dynamics of the valve is given by Equation 15 is combined with Equation 7 to give The pressure drop from location \u20191\u2019 to location \u20193\u2019 in Figure 1 is primarily caused by flow resistance and inertia. The stagnation pressures at locations \u20191\u2019 and \u20192\u2019 are almost equal in the steady state but may differ significantly under transient conditions because of fluid inertia in the orifice. Then, the pressure drop from location \u20192\u2019 to location \u20193\u2019 can be treated as the effect of flow resistance only and, in a form similar to (11), can be expressed as The distributed fluid-flow process has been approximated by a lumped model in which the pressure drop from \u20191\u2019 to \u20193\u2019 has been split into two parts: the transient component between \u20191\u2019 and \u20192\u2019 and the steady-state component between \u20192\u2019 and \u20193\u2019" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002258_tap.2017.2730250-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002258_tap.2017.2730250-Figure8-1.png", + "caption": "Fig. 8. Photographs of proposed water patch antenna. (a) 3D view, (b) Top view, (c) Side view.", + "texts": [ + " For realizing omnidirectional radiation pattern, the circular water patch and water ground are chosen with diameters of D1 and D2 respectively. The reason for choosing air is mainly due to its low dielectric constant of 1. The lower of the dielectric constant, the wider the impedance bandwidth can be achieved. Another reason is that the air is a kind of low-cost optical transparent material as well, which matches well with the water and the plexiglass container. Some valves are placed on top of the container for injecting liquid and removing the air, which can be shown in Fig. 8. In order to obtain wide impedance bandwidth, a metallic disk shape load and a thick air substrate about 0.15 \u03bb0 (\u03bb0 is the wavelength in free space at center frequency 2.4 GHz) between the two water layers are adopted. The disk load with 1 mm thickness is on top of the vertical feeding probe introducing some capacitance to cancel some inductance due to the vertical probe. Thus, wide impedance bandwidth is realized. The detailed parametric studies of both the disk-loaded probe and thickness of air substrate H1 are given in part V" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003819_9783527818815.ch8-Figure8.8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003819_9783527818815.ch8-Figure8.8-1.png", + "caption": "Figure 8.8 Schematic illustrations showing (a) desirable fillet weld profiles, (b) acceptable fillet weld profiles (that are less-than-attractive), and (c) unacceptable fillet weld profiles. Source: American Welding Society (2002). Used with permission of The American Welding Society, Miami, FL, USA.", + "texts": [ + "5 shows a tee and a lap joint. As for any fusion weld, there are good and there are bad fillet welds. For the most part, what makes a fillet weld good or bad is whether it has what is considered a proper profile (or cross section). In worst cases, bad welds are unacceptable because they would fail to provide required structural integrity. In other cases, although the resulting fillet weld might be acceptable from a structural standpoint, it is, nevertheless, unattractive or unsightly. The upper portion of Figure 8.8 schematically illustrates some examples of \u201cdesirable fillet weld profiles\u201d in (a) that would not only provide structural integrity but are also attractive. It also illustrates \u201cacceptable fillet weld profiles\u201d in (b) that are deemed would provide structural integrity but are less attractive. Most importantly, the figure illustrates several examples of \u201cunacceptable fillet weld profiles\u201d in (c) that would likely not provide suitable structural integrity, beyond being unattractive. The keys to structural integrity for fillet weld profiles are (i) balanced leg lengths, known as \u201cthroats,\u201d in the two generally orthogonal directions, (ii) sufficient penetration into each joint element, (iii) a face that extends beyond the hypotenuse of an imaginary triangle drawn from each toe of the fillet,2 i.e. has convexity, (iv) good transition from the fillet weld to the joint elements at each toe, and (v) not pronounced shape irregularity. Figure 8.8 shows representative examples of the most common problems with fillet weld profile. Like butt welds, fillet welds can exhibit other features that detract from visual appearance (e.g. weld spatter), as in the following sections. Just as the wake left by a boat on the otherwise smooth surface of a lake hints at how fast the boat was traveling by how sharply V-shaped the wake is, so too do the ripple marks on the solidified crown surfaces of a fusion weld given an indication of the speed at which a welding source was moved (i" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002362_amm.364.285-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002362_amm.364.285-Figure3-1.png", + "caption": "Fig. 3 The assembly model of unstacking unit", + "texts": [ + " First need to determine the design of some main parameters: the total length of the feed system, the feeding car\u2019s displacement in X-axis direction, the deliver car\u2019s displacement in Y-axis direction, the stroke of the hydraulic cylinder and the suction cylinder, the thickness of the workpiece and positioning accuracy. Choose the belt drive as the main transmission model of the sheet metal, the key parts of the mechanism include skip, hydraulic station, the thickness measurement device, suction device, lifting device, and finally the model of the double open buttress unit are created by using the software of the SolidWorks, as shown in figure 3. The unstacking unit\u2019s operating process including automatic sheets, sheet metal transfer, thickness measurement, and the operating procedures is shown in the following figure 4. 1-skip 2-hydraulic station 3-feed system Fig. 4 The operating processes of unstacking unit 4-thickness measurement device 5-suction device 6-lifting device 7-body The analysis of manipulator. At present, the sheet metal forming production line is operated at the high only about 2s left to the robot manipulator. In such a short period of time, the manipulator has to operate at a high speed of 200-250 m/min, so that the operations such as loading, moving and unloading can be fulfilled" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002778_pawr.2013.6490173-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002778_pawr.2013.6490173-Figure2-1.png", + "caption": "Fig. 2: Relationship between efficiency of ET and Class AB amplifiers", + "texts": [ + " The power efficiency of Class AB amplifiers degrades considerably, however, with increasing PAPR. Accordingly, attention is being focused on amplifier architectures such as Doherty, outphasing and Envelope Tracking (ET) that can preserve high efficiency even with high PAPR signals. ET is a major contender for emerging LTE systems. The structure of ET amplifiers is depicted in Fig. 1, which shows, in addition to the RF PA, an envelope amplifier (EA) which provides a dynamically varying power supply voltage that tracks the output envelope voltage. Fig. 2 shows how the efficiency of typical RF stages varies with output power, and illustrates how substantial improvements can be provided by ET amplifiers for high dynamic range signals. ET amplifiers have been shown to achieve efficiencies up to 61% for basestation WCDMA signals with 6.6 dB PAPR [1] and are well adapted to multiband applications with large power control ranges. For widespread deployment of ET systems, however, envelope amplifiers with low cost, high accuracy and high bandwidth are needed" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003421_1.2807062-Figure12-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003421_1.2807062-Figure12-1.png", + "caption": "Fig. 12 Plan-view diagram of a two-axle vehicle", + "texts": [ + " \u2022 Examine the wheel-rail contact forces and track following error with the different gain settings. \u2022 Check the robustness requirements against uncertainties with the different gain settings \u2022 Determine the optimal set of the control gains to be used according to design priorities or preferences e.g., for the minimum actuation requirements for the best robustness, etc. . Control Design Applied to a Two-Axle Vehicle A two-axle vehicle is used in the paper to demonstrate the efectiveness of the proposed control design approach, and a plan iew of the vehicle is given in Fig. 12. The vehicle configuration as been used in the past for the study of a number of wheelset ontrol strategies, two of which will be used in this study for omparison\u2014a full state-feedback optimal controller tuned to chieve a desirable stability without interfering with the natural urving 4 and an active yaw damping control tuned with the use f genetic algorithms for the best damping ratios 25 . 11002-6 / Vol. 130, JANUARY 2008 om: http://dynamicsystems.asmedigitalcollection.asme.org/ on 01/27/2016 Table 2 Comparison of different controllers Ka GA tuned 84,600 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003286_piee.1968.0211-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003286_piee.1968.0211-Figure3-1.png", + "caption": "Fig. 3 Effect of similarity transformations a Real eigenvalues b Complex eigenvalues", + "texts": [ + " It follows from this that the ^circles for all matrices related by a similarity transformation form a coaxial set of circles associated with the common eigenvalues. If the eigenvalues are real, the coaxial-circle set has the form of the force field between equal and opposite charges located at the real eigenvalues. If the eigenvalues are complex, the coaxial-circle set has the form of the equipotentials for equal and opposite charges located at the complex conjugate eigenvalues. This is illustrated by Fig. 3. 3 Higher-order flows In higher-order spaces, a similar resolution of the velocity vector x may be made in the plane containing x and JC; the properties of such higher-order mappings are still being studied. A most useful extension of the previous results can, however, be obtained in a simpler way. Suppose that a flow is specified by i=Ax) and let a pair of orthogonal vectors x and y span a specified plane in the state space. Resolve JC along x and y as (y, x} ) a n d s = - r y and let (28) where xB is the orthogonal projection of x onto the plane spanned by JC and y and " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001751_piers.2016.7735552-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001751_piers.2016.7735552-Figure7-1.png", + "caption": "Figure 7. Measured radiation patterns at (a) 5.8 GHz and (b) 6.15 GHz.", + "texts": [ + " This structure is suspended in air at a height of 1 mm from 70 mm\u00d7 70 mm square ground plane. The structure is fabricated and tested. The photograph of prototype structure is shown in Figure 6(a). Measured return loss using Agilent 9916A vector network analyzer and simulated return loss are shown in Figure 6(b). Measured results agree with the simulated results. The variation in measured and simulated results may be due to fabrication error, error in feed position and alignment error in the substrate layer. Measured radiation patterns are shown in Figure 7. The proposed antenna features stable radiation pattern over the desired band. The optimum antenna offers impedance bandwidth (S11 < \u22129.5 dB) and gain of 11.3 dB with gain variation of less than 2 dB over the 5.7\u20136.5 GHz band. Bandwidth of an antenna depends on electromagnetic coupling of two resonating frequencies of RIS and 2 \u00d7 2 MSA array. Optimum coupling provides wide bandwidth. Therefore the effect of dimensions, inter-element spacing and shape of RIS element on resonating frequency of RIS, fabricated on a suspended substrate below a 2\u00d7 2 MSA array have been analyzed" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000567_tmag.2017.2703844-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000567_tmag.2017.2703844-Figure4-1.png", + "caption": "Fig. 4. Formation of magnetic ring. (a) Toroid current sheet. (b) Structure of magnetic ring.", + "texts": [ + " The lead angle is given by arctan 2 2 arctan 1a ap R GR (5) where Ra is the radius of airgap. Considering that the value of varies with the radius, the radius of the airgap center is chosen on average. Based on energy conservation, the gear ratio, G, is derived by t n F G v T p (6) where and v are the rational and linear speed, respectively. It indicates that the ratio of thrust force and torque is constant, which is equal to the value of gear ratio. It\u2019s obvious that the thrust force of PMLS is the axial projection of the normal force. The toroid current sheet shown in Fig. 4(a) can be regarded as the projection of a helicalshape current sheet on xy plane. And the same current density, Jc, is assigned. According to the left-hand rule, axial force will be obtained with the contribution of radial component of magnetic field. The axial force Fa is given by '2 ( , ) R h a c rR F J B r z rdr (7) where B\u2019 r is the corresponding radial component of magnetic field. Compared with (3), the axial force Fa differs with thrust force Ft on the radial component of magnetic field. So Fa equals to Ft on condition that the same distribution of radial component of magnetic field is guaranteed. With the toroid current sheets shown in Fig. 4(a), the topology of magnetic ring, similar to the structure of PMLS, is developed in Fig. 4(b). The magnetic ring has an array of PM rings which is magnetized radially. The same axial width of PMs is guaranteed and the number of rings corresponds to the turn number of helical PMs on PMLS. The axial force is produced when the inner and outer parts travel axially, relatively. Since the structure of magnet ring is circular symmetric, the magnetic fields distribution of magnetic ring could be solved 0018-9464 (c) 2016 IEEE. Personal use is permitted, but republication/redistribution requires IEEE permission", + " Effect of the Number of Pole Pairs The effect of the number of pole pairs on thrust force was analyzed using 3-D FE model in [7], and the conclusion that the thrust force remains essentially constant as the number of pole pairs increases was drawn. According to the conclusion in [7], 2-D axis-symmetric FE model is applicable no matter how many the pole pairs are. Fig. 6 presents the 3-D FE models with the numbers of pole pairs being the only difference, the axial width of the four models is equal to 5 mm so that the same 2-D FE model shown in Fig. 4 is shared. The nut travels axially by a distance of 5 mm (the magnet axial width) while the screw is stationary. The simulation results of models in Fig. 5 and Fig. 6 are shown in Fig. 7. It can be seen that the characteristic of thrust force varies with the number of pole pairs, not in agreement with the conclusion in [7]. It\u2019s worth noticing that the max lead angle of PMLS in [7] is 14.15 deg, however, the lead angle greater than 14.15 deg is considered in this part. So the conclusion in [7] is inappropriate when lead angle increases up to a certain value" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001212_6.2008-5955-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001212_6.2008-5955-Figure4-1.png", + "caption": "Figure 4. Sample Dynamic Geometric Update", + "texts": [ + " If modifications of the general baseline are required, the original model can be updated and that modification can simultaneously be reflected in all the other disciplinary analyses due to the component inheritance and association present in the implemented object-oriented model. Additionally, the unified geometric surface generation can be used to modelled the effect of dynamic changes in the aircraft configuration such as those present in morphing aircraft, where the geometry can be updated dynamically as shown in Figure 4. 8 of 23 As previouly mentioned, consistent surface generation is essential to match the level of detail required by different disciplinary analyses. In the current implementation a watertight surface which is used as the unified outer-mold-line for the different analyses is generated using the pySurface class. Different approaches can be used to intersected and loft together the different surface components, to maintain consistency and extensibility in the surface generation they are integrated through the use of a generic Surface abstract class as shown in Figure 5" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002460_5138-ms-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002460_5138-ms-Figure3-1.png", + "caption": "Fig. 3 Example of Mesh Layout", + "texts": [ + "350\" is determined from welding workability. Diaphragms between stubs of overlapped cast steel node are removed from casting technology point of view. 2. Description.Qi. FE Analysis The structural characteristics of overlapped cast steel nodes are determined on the basis of a lot of numerical experiments conducted with finite element analysis. The computer program for structural analysis is the latest version of MSC!NASTRAN. 2.1 Analysis ~del A typical mesh layout of a finite element model is shown in Fig. 3. As shown in this figure, the fine mesh regions along the intersection are well controlled. Dimensions of the node models, K60, YT45, KT45 and KTS45 are listed in Table I. All of them are analyzed In the frame panels shown in Fig. 4. Dimensions of the frame panels are also given in this figure. References and illustrations at end of paper. 479 2 Static and Fatigue Strength Design of Overlapped Cast Steel Node .. OTC 5138 2.2 Loading and Supporting COndition From the structural characteristics point of view, the main role of a jacket node is to support the axial forces of incoming braces resulting from shear deformation of the frame structure" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001596_jsen.2014.2356598-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001596_jsen.2014.2356598-Figure4-1.png", + "caption": "Fig. 4. Diagram of rotating spindle and example error regions.", + "texts": [ + " In this paper, we assume the availability of a third radio, referred to as a base station, to synchronize the transmitter and receiver using the reference broadcast synchronization (RBS) algorithm [33]. The choice of radio synchronization method, though, is independent of the error avoidance algorithm. The synchronization process can be initiated periodically to account for the sensor nodes\u2019 clock drift. The functional structure of the measuring system including transmitter, receiver and base station is illustrated in Fig. 3. A personal computer records data from the receiver through UART (universal synchronous receiver/transmitter). Figure 4 shows the diagram of the Micaz radio sensor mounted on the rotating spindle and example error regions along the spindle circumference. 2) Experimental Procedure: Before each experiment, speed is measured and saved in the base station\u2019s configuration packet. To begin an experiment, the base station sends the configuration packet to both the transmitter and the receiver. Upon receiving configuration packet, the transmitter enters the training phase and starts sending probe packets, the receiver then calculates the PER profile based on the received packets" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003647_j.msea.2014.02.078-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003647_j.msea.2014.02.078-Figure5-1.png", + "caption": "Fig. 5. Surface plot showing the measured \u03a33 boundary length following a one hour anneal at 1025 1C of the IN600 samples deformed to various strains at different temperatures.", + "texts": [ + " Interestingly, these two samples also contained significantly high fractions of internal twin boundaries. In the other samples, the majority of twin boundaries were observed to either bisect the grain or originate from grain boundaries. Results from the assessment of the character of the grain boundaries in the various samples are summarized in Table 3 and Figs. 5\u20138. Over a constant scan area of 350 \u03bcm 350 \u03bcm, the total \u03a33 boundary length in each of the samples was averaged and plotted in the form of a surface plot as a function of both strain and deformation temperature, Fig. 5. Characterization of the as-deformed samples revealed that in all instances, the overall length of the \u03a33 boundaries decreased modestly when compared to the total length of \u03a33 boundaries in the as-received IN600 material prior to deformation (13,211 \u03bcm). Following annealing, however, \u03a33 boundaries were created and the measured length of \u03a33 boundaries in samples deformed at 25 1C/11% strain, 538 1C/11% strain, 760 1C/11% strain, 25 1C/28% strain, 760 1C/ 28% strain increased modestly to length values greater than 13,211 \u03bcm", + " Grain growth occurred in all of the deformed IN600 samples following annealing and this led to an overall reduction in the length of the random grain boundaries. Since samples deformed to strains of 28% and 51% at deformation temperatures of 25 1C, 538 1C and 760 1C were observed to recrystallize either partially or fully, the resulting grain refinement served to create random grain boundaries and counteract the reduction attributed to grain growth. This is reflected in by the bulge in the center of the surface plot in Fig. 5. Overall, the average length of the random grain boundaries was observed to be inversely proportional to the average grain sizes measured for the specimens, Fig. 6. Large grain sizes correspond to relatively low values of the random grain boundary length while the fine grains were associated with the higher values. The average grain size was observed to increase as a function of temperature, but decrease as a function of increasing strain. With quantification of the average lengths associated with random and \u03a33 boundaries in the various samples, the length fraction ratio (total \u03a33 boundary length/total length of all boundaries) was plotted, Fig", + " A bi-modal grain size distribution characteristic of partial recrystallization was observed to occur during annealing of samples deformed to 28% and 51% strain at 25 1C and 538 1C, Fig. 4b and e. Due to the relatively small size and lack of \u03a33 boundaries in the newly recrystallized grains, the fraction of random grain boundaries increased, Fig. 6, while the grain size decreased, Fig. 7. Since the recrystallized grains are also replacing the original grain structure that contained moderate fractions of \u03a33 boundaries, the overall length of the \u03a33 boundaries in the partially recrystallized specimens also decreased, Fig. 5. Increasing the deformation strain to 80% produced fully recrystallized microstructures after annealing, Fig. 4c and f. Once fully recrystallized, the new strain free grains do grow during annealing and some \u03a33 annealing twins form. Compared to the samples deformed to 11% at 25 1C and 538 1C, grain refinement does occur following annealing of the 80% deformed samples, Fig. 7. As there was no residual strain or significant dislocation density within the grains to stimulate the formation of \u03a33 boundaries during annealing the average overall \u03a33 length and \u03a33 length fraction are relatively low in these fully recrystallized samples, Figs" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003573_tcsi.2010.2043994-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003573_tcsi.2010.2043994-Figure4-1.png", + "caption": "Fig. 4. Two arbitrarily oriented uniform linear arrays in free space.", + "texts": [ + " Regarding the receive impedance matrix , we have (51) Finding the mutual coupling between the antennas of the receiver and the transmitter is complicated by the fact that the mutual coupling depends on the medium that connects the receiver and the transmitter. In order to keep things simple, we consider only the case where the receiver and the transmitter are located in free space. Suppose the receiver is located at elevation from the transmitter\u2019s point of view. Similarly, the transmitter is located at elevation from the receiver\u2019s point of view (see Fig. 4). Let us call the distance between the th receive and th transmit antennas. Then, , where (52) while is the distance between the first antenna of the transmitter and the first antenna of the receiver. The electric field vector at the th antenna of the receiver excited by the th antenna of the transmitter becomes (53) Recall from the discussion in Section II-D that (53) requires such that the antennas at the receiver do not disturb the field. Herein, is the vector of the complex current envelopes of the receiver-side ports of the multi-antenna multiport (see Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000718_bmei.2010.5639976-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000718_bmei.2010.5639976-Figure4-1.png", + "caption": "Figure 4. An articulation.", + "texts": [ + " The articulations are structured to be bendable in the direction along the bevel-tip. The parts of a needle between two articulations are called sections, which are more flexible than the other parts of the needle. The lengths of the head and sections are named as L0, L1, and L2 etc., respectively. The articulations are designed to be bendable only along the direction of the bevel-tip. The maximum angles of articulations are sequentially noted as 1m\u03b1 , 2m\u03b1 , and 3m\u03b1 etc. An enlarged articulation and its bent status are shown in Fig. 4. Combinations of different L and different m\u03b1 generate different needle bent shapes. Take the needle with two articulations, as shown in Fig. 3, as an example, and suppose the sections are rigid. 1) Different section lengths with equal articulation angles When L0, L1, and L2 are equal, 1m\u03b1 and 2m\u03b1 are both equal to 10 degree, the needle bent shape is given in Fig. 5(a). When L0, L1, and L2 are different, for example, L0 is half L1, and L1 is half L2, generated needle bent shape is given in Fig. 5(b)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000164_a:1011136822422-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000164_a:1011136822422-Figure2-1.png", + "caption": "Figure 2. Command derivation of the proposed dynamic interaction controller for surface tracking with force regulation.", + "texts": [ + " The proposed compliance control strategy is basically the type of hybrid control. Yet, to determine a proper CR when the stiffness of the environment and the shape of the environmental surface are not known in advance, a learning algorithm is developed for force regulation. Compared with previous approaches for on-line estimation and adaptation of uncertain environments [8, 32], the proposed learning algorithm is simple and with fewer parameters for adjustment, as demonstrated in the derivation of CR for the contact case below. Referring to Figure 2, CR consists of Ds for surface tracking, Df for achieving a desired contact force, and PR, as described in Equation (4): CR(k + 1) = Ds(k) + Df (k) + PR(k), (4) where CR = [XCR , YCR , ZCR ]t , Ds = [XDs , YDs , ZDs ]t , Df = [XDf , YDf , ZDf ]t , PR = [XPR , YPR , ZPR ]t , and k is the time step. Ds and Df are derived from Dj and the measured contact force Fc = [ Fcx, Fcy, Fcz ]t , as follows. In Figure 2, when the robot manipulator moves to contact with the environment from the free space, Dj is usually not along the direction of the environmental surface due to the imprecise manipulation of the operator. Dj can generally be divided into two portions: one is along and the other perpendicular to the environmental surface. As shown in Figure 2, Ds is taken just as the projection of Dj on the environmental surface. However, since the shape of the environmental surface is not known exactly, Dj cannot be projected onto the surface directly. Instead, the measured contact force Fc, providing the directional information for the environmental surface, is used for deriving this projection, as described in Equation (5): Ds(k) = 1 \u2016Fc(k)\u20162 \u00b7 F 2 cy + F 2 cz \u2212Fcx \u00b7 Fcy \u2212Fcx \u00b7 Fcz \u2212Fcx \u00b7 Fcy F 2 cx + F 2 cz \u2212Fcy \u00b7 Fcz \u2212Fcx \u00b7 Fcz \u2212Fcy \u00b7 Fcz F 2 cx + F 2 cy \u00b7 Dj(k)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000735_tmech.2004.828657-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000735_tmech.2004.828657-Figure7-1.png", + "caption": "Fig. 7. Schemes and pictures of the different experimental configurations. (a) Grasping of a sphere by the needle tip; (b) by one finger tip; (c) by two finger tips; and (d) by three finger tips.", + "texts": [ + " These simple tools have been fabricated in steel by traditional machining technologies (except for the needle which is commercialized by NOVOFINE\u00ae) and have been actuated by motorized manipulators with 1- m accuracy (MS-314, Marzhauser-Wetzlar). The tools with 2 and 3 cubic fingers have just 1 DOF because all the fingers are actuated by one precision positioner. Table I illustrates the results of the tests performed with two spheres of different diameters, one micro screw, one cylinder, one pipe and one silicon chip. The tests have been performed in an environment with a relative humidity of 60%. Fig. 7 shows some pictures of the different experimental scenarios. A direct comparison between model and experiments can be performed just for the two spheres grasped by the needle and the single finger. In both cases the adhesion forces between the tools and the spheres are not sufficient to overcome the gravity force and the substrate interaction force (compare Fig. 6). The instantaneous adhesion of the smaller sphere to the one-finger gripper is consistent with the larger surface tension force exhibited by the cubic finger, as predicted by the model" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002224_3527608117-Figure11.10-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002224_3527608117-Figure11.10-1.png", + "caption": "Fig. 11.10 Schematic structure of a twin atomizer.", + "texts": [], + "surrounding_texts": [ + "11.3.3\nFor large volume production it is recommended to use specially designed spray forming units for a specific alloy. The disadvantage of this is that if there was an unexpected alloy change at the plant this would lead to an extensive service processing time. The manufacturing plant in St. Avold, France consists of a 1.2 t inductive melting furnace, a 2.5 t holding furnace with a flanged fore-hearth and the actual spraying chamber.\nThe melting furnace is filled from a container by a tilting apparatus. From there the molten metal is transferred via a launder into the holding furnace. The complete unit is situated on load cells to record and control continuously the data related to the metal. By controlled pressure increase in the holding furnace the molten metal is forced into the fore-hearth. The PEAK Werkstoff GmbH runs two serial spray forming units and one F & E spray forming unit.\nFor spray forming the holding furnace is moved in the direction of the spray chamber and the fore-hearth is docked onto the spray chamber. The pressure of the fore-hearth on the spray chamber can be adjusted by load cells in the hydraulic cylinders, therefore absolute impermeability can be guaranteed and the sealing elements are not mechanically destroyed.\nThe most important feature of the holding furnace is the pressure control. Using that, it is possible to keep the bath level constant during spraying, independent of the furnace filling level. A constant bath level is a mandatory condition to control the rate of metal flow accurately. This happens by a continuous increase in the overpressure within the holding furnace at a speed which corresponds exactly with the rate of flow. The measured bath level, taken by a float lever in the fore-hearth, is used as the actual value for control. At the end of the spraying process the pres-", + "sure is reduced only as far as to ensure that the applied foam filter in the forehearth is just still wetted by the molten metal. This is important for a long service life of the filter. Furthermore, the cleaning effort of the fore-hearth is thereby considerably reduced.\n11.3.4\nnozzles are identical for each process.\n284 11 Spray Forming \u2013 An Alternative Manufacturing Technique for MMC Aluminum Alloys\n11.3.5\nThe atomization unit applied in spray forming is a two-stage arrangement, consisting of two concentric gas ring nozzles, the primary gas nozzle and the secondary gas nozzle. New ceramic funnels are inserted in the primary gas nozzle for each spraying run. The task of these nozzles is to avoid direct melt contact with the steel ring nozzles and their retainers.\nPrimary and secondary gas nozzles are fixed tightly at the nozzle retainer, at their feeding connection the feed-in of the primary gas, or rather secondary gas, takes place.", + "11.3.6\nThe primary gas, which surrounds the discharging concentric melt stream, leads the jet through the secondary gas nozzle up to the atomization point. Furthermore a widening of the melt stream is avoided, so that the formation of enclosed recirculation zones near the nozzle area is prevented. In the case of such \u201cflow dead areas\u201d, a pre-diffusion of the melt would take place. Enclosed recirculation zones are avoided by sufficiently high primary gas pressures. The open jet effect developing during the emission from the atomizer, with the formation of a low pressure area between primary and secondary gas nozzle, is mainly compensated for by the primary gas stream. A constant gas pressure of the primary gas stream is an important contribution to good flow behavior.\n11.3.7\nThe secondary gas nozzle, also called the atomizer nozzle, is responsible for the atomization (diffusion) of the melt stream. The secondary gas, which flow through holes arranged concentrically at a certain angle, destroys the guided melt stream at the atomizing point. This point is a few centimeters below the secondary gas nozzle.\nDuring the impact of the secondary gas stream on the melt stream, an impulse transmission occurs. A spray cone builds up, whose particle spectrum consists of particles varying in size from 1 to 400 \u00b5m. The shape of the spray cone formation is dependent on different spraying parameters, for example the nozzle geometry, the pressure and speed relation and the gas / metal ratio.\nBesides diffusing the melt stream, the secondary gas has a further function to accelerate the diffused droplets onto a substrate, which is situated in the center of the spraying chamber, and to cool down the droplets before impact. Figure 11.11\n28511.3 Techniques" + ] + }, + { + "image_filename": "designv6_24_0001626_tia.2017.2766585-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001626_tia.2017.2766585-Figure1-1.png", + "caption": "Fig. 1. Coil placement in the machine.", + "texts": [ + " The FEM results confirmed that the torque ripples in the optimized IM have been reduced significantly, and hence, validate design suggestions obtained from optimized IM model. Mutual and self-inductance calculation is considered the basis for most of the model formulations of the IMs. Self-inductances and the mutual inductances can directly be connected with the winding formations. Different types of winding corresponds to different types of pitch and belt factors. To calculate the inductance first we need to determine the spatial distribution of magneto motive force of single coil in the air-gap periphery as shown in Fig. 1. The spatial distribution for the magnetomotive force (MMF) is calculated for any common phase \u201cq\u201d in the winding shown in Fig. 2. Using this establishment, the harmonic magnetic flux linkages can be determined between any two generalized phases \u201cq\u201d and \u201cr\u201d. These flux linkages are responsible for the harmonic inductances. To calculate the self-inductances we will use \u201cq = r\u201d and for mutual inductances \u201cq = r\u201d. Starting with this assumption, the relation between the currents determined from differential and the torque output of the motor with the coenergy concept" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000194_s12239-015-0047-9-Figure14-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000194_s12239-015-0047-9-Figure14-1.png", + "caption": "Figure 14. Traction limitation of each tire.", + "texts": [], + "surrounding_texts": [ + "Traction force distribution is based on the desired yaw moment which is determined in Section 4. Transfer case system can control the torque distribution between the front and rear axles, so the total longitudinal traction force of the vehicle is not changed with an assumption with no torque loss due to friction and so on. Torque vectoring system can create a torque bias between the left and right wheel torque of rear axle; so while it makes yaw moment of the vehicle, the total longitudinal traction force of rear is not changed Izz\u03b3\u00b7 lf Fyf lrFyr wr 2 ---- Fxrr Fxrl\u2013( ) lf Fyf lrFyr\u2013 Mz+=+\u2013= may Fyf Fyr+= F\u0302yf lrmay Mz Izz\u03b3\u00b7+ + L ----------------------------------- F\u0302yr lf may Mz\u2013 Izz\u03b3\u00b7\u2013 L -----------------------------------=,=" + ] + }, + { + "image_filename": "designv6_24_0003345_amm.813-814.964-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003345_amm.813-814.964-Figure2-1.png", + "caption": "Figure 2. 3-D Cad Model of Brake Pedal.", + "texts": [ + " No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of Trans Tech Publications, www.ttp.net. (#69200032, Swinburne University, Hawthorn, Australia-24/08/16,18:07:48) will travel additional distance before it stops. The design of automatic braking control systems is clearly dependent on the braking system characteristics - Material property. Earlier Brake Pedal Figure 1. Originals Brake Pedal. The pedal shown in figure is used by the driver of a vehicle to operate the brakes. This pedal is hinged about base plate by a hinged pin our foot apply force on this pedal. As shown figure 2, the 3D cad model of brake pedal is drawn in creo2.0 software as per dimension specification of old brake pedal.As existing pedal has more mass,best cad models are developed with consideration of different aspect as shown in figure 3. The material used for the brake pedal is is cold rolled steel sheet(FePo3 En10130 series) series which has following chemical and mechanical properties as shown in Table 1. As shown in Figure 4, with the help of ANSYS software, the geometric model was divided into tetrahedral finite elements of higher order (solid187)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000957_1.1559581-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000957_1.1559581-Figure4-1.png", + "caption": "Fig. 4 Cardinal spline and its curve fitting", + "texts": [ + " Hence, when a curve is specified, it is better not to use the correction function for a point. This is because we can represent the curve by a spline curve itself when points are given as a point sequence. In order to treat the points which specify a curve similarly to the point specification and simultaneously with it, we use the Cardinal spline as a correction function. The Cardinal spline is a spline that interpolates values, which are 1 for the corresponding segment parameter and 0 for the other segment parameters. In Fig. 4, the Cardinal splines for six points are shown in the top left, the Cardinal splines multiplied by correction values and their summation are in the top right, and the modified curve is below, which is obtained using these splines as correction functions. The correction function for a curve does not satisfy the first requirement, non-negative, but it generates a smooth curve, because the requirement is essential especially for the case that one point is specified. ring DECEMBER 2002, Vol. 2 \u00d5 267 ess" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000234_icelmach.2016.7732762-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000234_icelmach.2016.7732762-Figure6-1.png", + "caption": "Fig. 6. Lamination of mcC: for 2 and 4 poles according to the PM magnetization direction", + "texts": [ + " The iron losses account for a slightly high percentage in the 2-pole motor due to the higher stator flux density. Attention should be paid to the synchronization in the case of 4-pole, since the high PM flux linkage causes a high breaking torque during start up. This is the cause of the relatively high synchronization voltage of the 4-pole motor, see Table IV. However it remains below the rated one. C. mcC: for 2 and 4 poles with induced pole The geometrical data of mcC are shown in Table V. Its lamination, illustrated in Fig. 6, is substantially different with respect the one of mcB, showed in SubsectionII-B: there are only two rotor flux barriers in the rotor. In the case of 4-pole, the PMs are arranged so as to obtain an induced pole rotor structure [13]. Thus, two poles located in alternate positions require PMs and the other two do not. These induced poles are achieved by the flux re-distribution. On one hand, good performance of the 4- pole machine are achieved when the flux barrier angle is at about 45 degrees with respect the PM magnetization axis", + " However, it should be mentioned that the volume of mcC is 32% higher than the mcA one and 20% higher than the mcB one. It is worth noticing that the 4-pole machine exhibits a high percentage of iron losses, even if the no-load flux density in the stator parts are rather low. This is due to the stator flux density that increases considerably when the machine is loaded. Experimental measurements are carried out on a LSSM prototype, whose geometrical data are in Table VII. Its lamination is of the type shown in Fig. 6, that is, the mcC. It has been optimized to run as 2-pole motor and hence bad performance are expected when it works as 4-pole induced pole motor. However, the measurements carried out in this Section allows the analysis presented in the paper to be validated. The prototype is mounted on a test bench in which a torquecontrolled master IM is used to reproduce the effect of the load. The motor under test has been connected to the master motor through a torque transducer and supplied by the grid. A sketch of the experimental set-up is shown in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001388_s1644-9665(12)60163-0-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001388_s1644-9665(12)60163-0-Figure7-1.png", + "caption": "Fig. 7. Scheme of deep drawing by hydroforming (a) and example of product (b) [26]", + "texts": [ + " By this process, the limit drawing ratio value of sheet metal can be increased. The liquid can be used as a punch, a draw die or an assisting way to improve sheet formability. Actually, almost all of the materials used in conventional stamping can be used in sheet hydroforming. Depending on the different means, the liquid pressure in the die cavity is from around 30 to 150 MPa, but the usage of 200 MPa has also been reported [13, 25]. Roof for luxury class car (Figure 5), deep and partially conical cup (Figure 7) or cups with complicated geometries of bottoms (Figure 6) are examples of parts made by sheet hydroforming. It is well known that formability of the lightweight materials usually increases at elevated temperature levels [27\u201328].Warm forming technology with selective heating enables manufacturing of lightweight parts by utilizing the increased formability at elevated temperature [28]. However, it is quite difficult to determine optimal temperature distributions in tooling [29]. Warm hydroforming technology for lightweight materials is currently being developed to achieve reduced number of manufacturing steps and part consolidation" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001826_870305-Figure16-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001826_870305-Figure16-1.png", + "caption": "Fig. 16", + "texts": [], + "surrounding_texts": [ + "8 Component\nliP + defroster air duct ......... Insfr. cluster bezel\n--- Glove box + door\nPPO/PS PP/EPDM PPO/PS, PC/ABS PPO/PS\n870305\nGolf C (lEurope)-lnjection-Molded Instrument Panel\nFig. 1J\nVW Golf GL - Foam-Plastic Instrument Panel, Individual Components\nFig. 14", + "870305\n9\nTrends: See chart Also higher degree of integration: Formation of large pre-assembled units, consisting of instrument panel. heater or air conditioner, steering, pedal~, fuse-and-relay board.\nHaterial: Trends:\nPP, ashtray: FS 31 (phenoplast, saw-dust-filled) - Multi-color injection-molded - Textile decor - Softer surface", + "10\no Heater and air conditiorer\n870305\nHaterials:\nTrends:\n- Fousing parts: PP TF - Blower wheel: PON - Thin-layer technique for - Flaps hard/soft\n- Flilps: - Controls:\nhousing parts\nNetal/PU foam ABS" + ] + }, + { + "image_filename": "designv6_24_0001983_el:19790293-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001983_el:19790293-Figure1-1.png", + "caption": "Fig. 1 Geometry of lens-type compact antenna range", + "texts": [ + " The novel feature of the lens type of compact antenna range reported in this letter is the introduction of a controlled amount of loss into the lens, which compensates for the amplitude taper and produces an electric field across the lens which is nearly constant in both amplitude and phase. The lens transforms the spherical phase front radiated by the source into a plane phase front at the secondary surface of the lens. Refraction may take place at one or both surfaces, but we have shown that the thinnest lens is produced when all the refraction occurs at the surface facing the source and the secondary surface is flat (Fig. 1). The use of geometric optics and the application of the principle of equality of path lengths leads to the equation of the surface of the lens: R = Hn - rj c o s 0 \u2014 1 where r\\ is the refractive index of the lens material. This is the equation of a hyperbola with eccentricity equal to the refractive index and the focus at the feed point. The thickness of the lens along the axis is a function of the focal length, lens diameter and refractive index of the lens. For small F/D or low t], the lens is very thick, e", + "5 In the near-field pattern method, a Gaussian-shape pattern in the single-mode region changes at Xc owing to appearance of the second-order modes. The method has the disadvantage that the criterion of pattern deviation from a Gaussian shape has an ambiguity. In this letter, a new method for measuring the cutoff wavelength of the second-order modes is described. The method is found to offer a measuring technique that is highly sensitive and accurate compared with the previous methods mentioned above. Fig. 1 shows the experimental setup. A light beam from a halogen lamp through a monochromator is polarised in the Y-direction by a polariser and is launched into a test fibre in an off-axis condition. The output power is detected after passage through an analyser set in the X-direction. The second-order modes are excited below the cutoff wavelength Xc. The feature of the mode excitation changes with the offset direction except for the HE2 i-mode.6 When the offset is along the X-direction, which is perpendicular to the polarisation plane of the input beam, only the TM01- and HE2i-modes are excited" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002572_978-3-319-76276-0_5-Figure10-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002572_978-3-319-76276-0_5-Figure10-1.png", + "caption": "Fig. 10 Fatigue zone for dynamic middle axle inner suspension spring deflection of 7, 5.27, 4.42 and 3.97 mm", + "texts": [ + "3 mm, the displacement amplitude response is determined for varying damping coefficient as shown in Fig. 11. From Fig. 11, it is observed that when damping coefficient is increased from 100,000 N s/m to 250,000 N s/m, displacement amplitude reduces from 3.58 mm to 2.21 mm and this reduces dynamic amplitude of suspension spring. The fatigue analysis is carried out corresponding to the dynamic spring deflection of 7, 5.27, 4.42 and 3.97 mm and given in Table 3 and its fatigue zone with crack propagation obtained by FE analysis is shown in Fig. 10. Hence it is suggested to increase damping coefficient of damper to increase life of inner suspension spring i.e. from 1.62 104 cycles to 6.67 106 cycles. Thus the damping coefficient of 250,000 N s/m will reduce the dynamic amplitude to 1.99 mm and the expected life of inner suspension spring will be approximately 6.67 106 cycles which is considered to be infinite life. Hence damping coefficient of 250,000 N s/m is recommended for existing suspension (Fig. 11). iii. Provision of Shim Along with Change in Damping Coefficient of Inclined Damper Third modification discusses, the shim provision on end axle box housing reduces the shear stress on middle axle primary inner suspension spring, also increase in damping coefficient of inclined damper reduces the amplitude of suspension spring and both modification increases its fatigue life as discussed earlier" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000514_1.c034448-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000514_1.c034448-Figure1-1.png", + "caption": "Fig. 1 Distributed propulsion aircraft with wing shaping control.", + "texts": [ + " The concept of distributed propulsion can be leveraged to alter the generalized force, mass, and stiffness of a wing structure advantageously in connection with aeroelastic wing shaping control for the benefit of improved aerodynamic performance of a flexible wing aircraft. This study examines a fuel-optimal performance during climb, cruise, and continuous descent that accounts for wing aeroelasticity and optimal thrust distribution along the wing span. The propulsive forces and moments produced from the distributed propulsors mounted along the wing span can be used to optimize L\u2215D by modifying the wing twist and bending. Figure 1 illustrates one concept of a conventional transport aircraft having an inboard turbofan engine core that serves as a generator to provide power to the distributed propulsors, which could be electric fans along the wing span. It should be noted that a variety of configurations of differing numbers and locations of generators and propulsors, extent of thrust, and nacelle sizes, and placement along the wing span either above or below the wing as D ow nl oa de d by U N IV E R SI T Y O F M A N C H E ST E R o n M ar ch 1 2, 2 01 8 | h ttp :// ar c" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002605_973233-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002605_973233-Figure4-1.png", + "caption": "Fig. 4. A Straight Segment", + "texts": [ + " The clip geometry is defined by the centers and radii of the four arcs, from which the length and angle of the four connecting straight segments are calculated. PERFORMANCE INDEX CALCULATIONS - The derivation of the maximum insertion force is as follows. Castigliano's Theorem gives the deflection d along the direction of the contact force p as and where U is the total strain energy in the clip, ej is the compliance for ith segment and n is the total number of segments in the clip. A straight segment shown in Fig. 4 is used to demonstrate how the compliance is derived. The segment is subject to two loads: a force p and a moment pa during insertion, where a is the length of moment arm. The compliance ej from Eq. (8) can be calculated as where Mj is the moment at any cross-section, s is the distance along the segment, El is the flexural stiffness of the clip, and L and or are respectively the length and the angle of the segment. For the circular segment under the same loads (see Fig. 5), the compliance ej can be calculated as 3 c: + 3c; g2 = r [2c1c2 + (--- )8 - 2ci sine - 2clc2 cose 2 2 1 2 -CIC~ sin 0 + -(c2 - c;) sin281 4 (13) where p is the initial angle, and 8 is the total angle of the segment" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001630_j.biosystemseng.2013.10.015-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001630_j.biosystemseng.2013.10.015-Figure2-1.png", + "caption": "Fig. 2 e Blade designs evaluated (dimensions in mm unless oth width and (c) straight, e side view: (d) conventional.", + "texts": [ + " The rotor assembly was fitted to the test rig which travelled over a soil bin at a forward speed of 0.67 m s 1 (typical speed for walking behind a two-wheeled tractor such as the Chinese-made Dong Feng DF 12 using gear 2). Previous work on strip-tillage using a modified rotary till drill (e.g. Esdaile, 2009; Hossain et al., 2009; Justice et al., 2004) used commercially available left- and right-hand bent C-type blades with sidelong sections. These 43 mm wide blades are referred to as \u2018conventional blades\u2019 (control treatment) in this experiment (Fig. 2a). For the tests some conventional blades were modified to reduce the width of the sidelong sections to 22mm (Fig. 2b), and designated \u2018half-width blades\u2019. In addition some blades were manufactured with no sidelong section to provide the narrowest version of the C-type blades, and were designated as \u2018straight blades\u2019 (Fig. 2c). The straight blades had a thickness of 6 mm at the holder section tapering to 3 mm at the tip section. All the blades had double sided chamfers sharpened at 8 with a 1.5 mm cutting edge thickness. All blades were mounted with a rotor diameter of 306 mm. erwise specified) e front views: (a) conventional, (b) half- The conventional and half-width blades were fitted on the rotor with their blade tips bending towards the furrow centre. The cutting width of the four blade rotor was 50 mm (measured fromblade outside to outside surfaces, as shown in Fig. 2). The rotorwas operated at a depth of 50mmand run in a forward rotation so that the blades cut the soil downward, as per recommended practice for strip-tillage by Lee et al. (2003). The rotary tiller unit did not use a shield, as used on a commercial rotary till drill, as the study was concerned about the process of furrow cutting by the rotary tilling blades with no interaction from other equipment. The progression of soil cutting and throwing during the test was recorded by a high-speed camera (TroubleShooter TSHRMM, Fastec Imaging Corporation, San Diego, California, USA) recording at 1000 frames per second to capture the soil failure, throwing patterns, and creation of the seedbed furrow" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000957_1.1559581-Figure10-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000957_1.1559581-Figure10-1.png", + "caption": "Fig. 10 Attractive force to an operating point", + "texts": [ + " It consists of three kinds of force: attractive force to the point, frictional force between the point and a cursor, and elastic force of the point. The attractive force is the one to attract a cursor to the point with the force proportional to the distance when the cursor approaches to the point. The force is represented by the following equation: F5k~ uP02Pu2r0! P02P uP02Pu (9) where k is a scalar constant, P0 is a position vector of the specified point, P is a position vector of the cursor, r0 is the radius of the specified point, and r is the radius of the force field ~see Fig. 10!. Each operating point is a sphere that has a finite radius r0 and the cursor cannot go into it due to the stiffness of the surface. Therefore, the following equation is satisfied. uP02Pu>r0 This attractive force works when the cursor is within the force fields around operating points, so that the force from several points do not interfere each other. The frictional force and the elastic force are used in order to prevent the cursor slipping on the surface of the point. We determine parameters of point operation force, which are coefficient k in Eq" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003859_lawp.2021.3063369-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003859_lawp.2021.3063369-Figure1-1.png", + "caption": "Fig. 1. Proposed multifunction design of DDRA and independently controllable dual-band filter.", + "texts": [ + " Furthermore, the reflection ground is employed to obtain a microstrip filter function that provides the filter with the other independent passband. The DDRA and the dual-band filter share the same module but keep good isolation. The potential payoffs of the design are the following: 1) reduced number of microwave components in a system that needs an antenna and a filter; 2) independent operation without interaction effects between the two parts due to the separation of the virtual ground and the inherent isolation of the reflection ground; 3) the free and flexible design of the operating frequencies. Fig. 1 shows the configuration of the proposed multifunctional DDRA. The rectangular DR with the dimensions of a \u00d7 a \u00d7 h is constructed using ceramic material with relative permittivity \u03b5r1 = 38 and tan \u03b4 = 1.5 \u00d7 10\u22124. To take full advantage of 1536-1225 \u00a9 2021 IEEE. Personal use is permitted, but republication/redistribution requires IEEE permission. See https://www.ieee.org/publications/rights/index.html for more information. Authorized licensed use limited to: Dedan Kimathi University of Technology", + " 3(a) shows the structure for the first passband of the filter, which is designed on the virtual ground. Two U-shaped strip resonators with a width of w4 and a length of \u03bbd/4 (\u03bbd is the wavelength in the DR at the first center frequency f1) are coupled, and their coupling coefficient is controlled by the distance of d2. One of the ends of each strip line is shorted to the reflection ground. Each strip resonator is directly fed by another conformal strip etched on the side surface of the DR parallel to the xozplane with a length h2 and a width w2, as shown in Fig. 1. The resonator and its feeding strip are connected to each other with a short strip of length d1 and width w3. Fig. 3(b) displays the structure for the second passband of the filter, which is designed on the bottom surface of the substrate. Two \u03bbg/4 microstrip resonators (\u03bbg is the guided wavelength of the microstrip line at the second center frequency f2) are employed and coupled here that are, respectively, connected to the 50 \u03a9 microstrip lines. To introduce transmission zeros between the passbands of the filter, two open stubs are added to the feeding strips" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000644_apmc.2007.4554952-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000644_apmc.2007.4554952-Figure3-1.png", + "caption": "Fig. 3. Current distribution on the ground plane at the center frequency. (a) Magnitude of the current in x-direction. (b) Magnitude of the current in ydirection.", + "texts": [ + " To implement this idea, two coupling slots have been etched on the ground plane of the microstrip circuit and consequently on the top of the waveguide. In the optimization process, the larger separation between the two slots, the better matching was achieved. To match the waveguide to the microstrip line, both stubs were placed on the top of the slots. The structure in Fig 2(b) shows the final configuration in which either slots is placed close to the side walls of the waveguide. The magnitude of current on the ground plane and the microstrip lines are depicted in Fig. 3 and Fig. 4, respectively. It is obvious that the current and field vectors on the slots are in opposite directions. III. SIMULATION RESULTS After optimization of the transition, the simulations show over more than 22% of relative bandwidth from 16 to 22.2 GHz. Fig. 5 shows the return loss and Fig. 6 presents the coupling between two ports of the power splitter. It can be seen that both the coupling and return loss are acceptable in a 6 GHz bandwidth from 16 GHz to 22 GHz. The return loss is more than 15 dB in this range and the coupling coefficient is better than 4 dB for each port" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000092_8.247750-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000092_8.247750-Figure2-1.png", + "caption": "Fig. 2. Simpliifed model for the mutual coupling.", + "texts": [ + " V, and Vz are the feed point voltages of MSAl and MSA2, respectively, and they are given as (10) where po is the distance between the center of the MSA and the feeding point. In practice, antenna arrays are mounted on the faces of a polyhedron, so we will make an assumption to the diffraction coefficients as + = y ~ , +' = 0. Then they become Vi = Vz = ni(k11 Po)h, 0, = 0, (11) . (12) - sin( ?r/ y)e-'(\"/4) p@Zsin a cos(r/y) + 1 This assumption makes the analysis simple. Due to this simplification, the geometry of the problem becomes equivalent to Fig. 2. The geometry is interpreted as that of a mutual coupling between MSAs on planes separated by a fictitious wedge. The analysis of this model is possible because deviation of the diffracted field with w_edge angle y is contained in the diffraction coefficient D, and the other parts in J:, * Hd in (9) do not depend on y. 2 D, = We now get YZ1 as y,, = \"( J1( k,,ii) )), 4 v Jl(kl1 Po 5hg , - j k ( p + s ) , (13) (14) l T v d 4 1 /-=/+z d m g = -(GZ * A)( S l * K m l ) . We will now obtain an asymptotic expression for Y2,. It Employing some formulas from Appendix A, we finally will be assumed that the distances from the center of get the following result: very much greater than the equivalent radius. That is, YZI E PQ, Q, (29) d,,d2 >> Z. Next, we use in (14) the following identities obtained from Fig. 2: r l + ps^ = d,f + qt^, MSAl and MSA2 to the edge, d, and d,, respectively, are e - j k ( d , + d d ddldz(dl + 4 ) 1 Q = Ai COS 410 COS 4 2 0 + ---U2 COS +lo cos 420 + A , sin bl0 cos 420) (15) dl qt\u0302 + si = d 2 f + r 2 . (16) These identities give sin a. (d, - rl - 2 + cot a. * ( r , e j ) ) , (17) p=-zG- sin a. sin CY s = - {d, + r2 .P - cot a. ( r 2 v j ) ) , (18) p + s = I(dl + d2)f + r2 - r,l. (19) Now we can expand parts of the integrand as 1 - Ql (20) \u2019h &GGTJ= 4 m A 9 1 A = 1 + -{rl .P - cot a. - ( r , .E)> 2 4 1 d2 A , sin +lo sin 420 + - ( A 2 cos 4*o cos 4 2 0 + A3 cos 4 1 0 sin 4 2 0 ) + 7 (30) dl + d2 A , = . r r2Z2(Jo(E) - J2(kii)}2, L (31) It is noted that the above approximate formula of Y21 can be applied to H-plane coupled elements as well as E-plane coupled elements. The reader can see Y2, will not vanish when the MSAs are coupled by the H plane. That is because we have summed up every coupling between infinitesimal equivalent magnetic current elements, as can be seen in Fig. 2. It is also noted that the first-order term in the above expression of Y21 can be interpreted as the product of the radiation pattern of each MSA in the direction of the edge and the diffracted field due to a point source separated by the wedge. The mutual coupling coefficient S,, between the two elements is calculated as where Yll is a self-admittance of the MSA calculated by conventional methods [2], and Y is a characteristic admittance of the transmission line. It is appropriate to define antenna isolation SAi beplings between a radiating element i in one array and all In the above we have used the tween arrays on a polyhedron structure as mutual COUformulas: p1 = (s^" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002661_iros.2011.6094731-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002661_iros.2011.6094731-Figure7-1.png", + "caption": "Fig. 7. Laser beam director and tracking system: A camera mounted parallel to the laser monitors one bright LED on the vehicle. The tracking mirror is adjusted, so that the LED is always in the middle of the mirror.", + "texts": [ + " The Tracking System is a development of LaserMotive and not scope of this paper. Only a short overview is to be given here defining the interface to the onboard control system. The tracking is based on one camera tracking one red LED in the middle of the solar cell array, shown in Fig. 6. The optical axis of the camera is adjusted parallel to the laser beam and aimed to the vehicle by the tracking mirror. The tracking algorithms control the mirror so that the LED is always in the middle of the picture and therefore the power beam is directed to the vehicle (see Fig. 7). 7Joint Test Action Group. IEEE standard for a programming and testing architecture. It is used for comfortable programming and on-chip debugging A reference (O) frame was set centered over the laser beam. The vision system provides the tracking angles referred to a vertical axis as shown in Fig. 7. Using these angles and the altitude z measured by air pressure, the reference position can be calculated: x = sin (2 \u00b7 \u03b2) cos (\u03b1) \u00b7 z y = tan (\u03b1) \u00b7 z (2) The laser tracking provides its tracking angles at an update rate of 50 Hz. Despite the comparable high rate this information alone is not sufficient for precise position control because of the latency and precision. A fast data fusion is necessary to fuse vision and inertial measurements in order to achieve maximum dynamics, bandwidth and robustness" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000715_062035-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000715_062035-Figure5-1.png", + "caption": "Figure 5(a). Heat flux contours for EN8C Figure 5(b). Temperature contours for EN8C", + "texts": [], + "surrounding_texts": [ + "The boundary conditions followed in the present work is mentioned below and these points are considered in the design and analysis of the proposed work. (i) The model was meshed using tetrahedron method. (ii) The element was meshed with coarse type and medium smoothing. (iii) Number of nodes = 3986 (iv) Number of Elements = 2003 (v) After this, the teeth section of the rack was refined further to provide a more accurate simulation model and to give more accurate results. (vi) Number of nodes (after refinement) = 16380 (vii) Number of elements (after refinement) = 8266 To set-up the thermal-electric simulation, the number of time steps over which the calculation is done is take as 10 (not the default 1) time steps with 1 second per step and an update interval of 2.5 seconds kept as default. This is done so as to provide ample number of iterations and at the same time, to provide a 5 IOP Conf. Series: Materials Science and Engineering 263 (2017) 062035 doi:10.1088/1757-899X/263/6/062035 congruence point in a relatively quicker time [10-12]. Also, the radiosity controls that measure the relation between the radiation convergence and the surface parameters has been kept on with a convergence value of 0.0001 W/m2. Smooth contours have been set to provide details about the variations in the input and output results. The meshed model of the designed structure is shown Figure 1. The solver used here was of gauss-seidel iterative type. These are the FEM equations needed and used by the solver while analyzing the data given and these are the main equations used for determining the results. The resources for this data were taken from the help section in the project window main screen. Biot-Savart Law for finding the magnetic vector potential Gauss Law for determining flux density Faraday\u2019s law for calculating the electric field intensity. Ampere\u2019s Law for current density (not compulsory). Equation for determining Joule Heat Generated. Taking into account, the real-time conditions present, an input voltage of 525 volts was applied in increasing steps throughout each of the ten time steps with phase voltage set to zero. To improve the accuracy of the readings, the voltage was applied across the 28 teeth only, but the output was generalized throughout the entire surface of the rack. Also as part of the thermal input, the room temperature was maintained at 30 \u00b0C and the gear was initially placed at the room temperature. As the voltage was applied, the rack was simultaneously cooled by the surrounding air and the coolant. For this, the convection factor was considered between the outer surface and water. Also, the heat flow was assumed to flow through the rack body and between the rack and the surrounding medium and so the total heat flux was evaluated as part of the results. 3. Results and Discussion 6 IOP Conf. Series: Materials Science and Engineering 263 (2017) 062035 doi:10.1088/1757-899X/263/6/062035 In the figure 2(a) given above the step increase of the input voltage provided to heat the rack is given in a graphical format to a maximum value of 525 V. Figure 2(b) shows the variation of the Convection Coefficient of quench water with respect to the temperature. Maximum Value Over Time Electric Voltage Joule Heat Total Electric Field Intensity Minimum Value Over Time Minimum 100. V 8.0742e-015 W/m\u00b3 1.0846e-011 V/m Maximum 525. V 2.3646e-013 W/m\u00b3 7.8596e-011 V/m Maximum Value Over Time Minimum 100. V 0.56343 W/m\u00b3 1.7167e-004 V/m Maximum 525. V 15.531 W/m\u00b3 9.013e-004 V/m 7 IOP Conf. Series: Materials Science and Engineering 263 (2017) 062035 doi:10.1088/1757-899X/263/6/062035 Minimum Value Over Time Minimum 100. V 4.4055e-015 W/m\u00b3 1.0291e-011 V/m Maximum 525. V 1.1376e-013 W/m\u00b3 8.415e-011 V/m Maximum Value Over Time Temperature Total Heat Flux Minimum Value Over Time Minimum 265.38 \u00b0C 8.2511 W/m\u00b2 Maximum 589.72 \u00b0C 1266.3 W/m\u00b2 Maximum Value Over Time Minimum 421.04 \u00b0C 4.9745e+005 W/m\u00b2 8 IOP Conf. Series: Materials Science and Engineering 263 (2017) 062035 doi:10.1088/1757-899X/263/6/062035 Maximum 611.41 \u00b0C 9.3048e+005 W/m\u00b2 The tables given above 5 to 7, denotes the maximum and minimum value of the output parameters for all of the materials based on the iterations conducted for all the time steps. These values are taken as comparison against the standard values of EN8C. As It is noticeable, for the given input voltage the Joule heat value is the maximum for Cast Iron > EN8C > Structural Steel. This means that maximum heat is produced in Cast Iron for the same voltage. However, the electric field intensity is the maximum for Structural Steel > EN8C > Cast Iron. For Heat Flux Structural Steel > EN8C > Cast Iron. This implies that the rate of heat energy transfer is maximum in Structural Steel and minimum in Cast Iron. 9 IOP Conf. Series: Materials Science and Engineering 263 (2017) 062035 doi:10.1088/1757-899X/263/6/062035 Figure 3(a) shown above, is a close up view of the heat flux variation in the teeth section in the rack. As seen, it is not even around the teeth gaps and varies individually. The figures given above depict the contours obtained for the parameters Heat Flux and Temperature for all of the simulated materials. These are based on the iterative values whose boundary limits are mentioned in the tables 5-7. As seen from figures 3(b), 4(a), 5(a) we notice that the Heat Flux is maximum in the teeth region of the rack gear and starts decreasing as we move across the length of the rod. This is due to the dispersion by conduction and convection to the quench water. The same can be derived for the Temperature contours in 3(c), 4(b), 5(b). Here the temperature drop is noticeable right from the edge of the teeth section throughout the length of the rack. 1234567890 IOP Conf. Series: Materials Science and Engineering 263 (2017) 062035 doi:10.1088/1757-899X/263/6/062035 The Figure 6(a) provides the graphical representation of Joule Heat variation for 525 V. It is deduced that up to the 5th time step the values are almost similar but after this, the curve for Cast Iron rises almost exponentially when compared to the other materials. This is shown as the error percentage is between 2% and 7%. To test if this variation was constant, the input voltage was raised 615 V from the standard 525 V as shown in Figure 6(b). As expected, the values of Joule Heat for Cast Iron showed a similar rise after the 1234567890 IOP Conf. Series: Materials Science and Engineering 263 (2017) 062035 doi:10.1088/1757-899X/263/6/062035 4th time step while the values for the other materials remained constant with a small error percentage throughout the range of the time steps. This led to the conclusion that based purely on the results of this analysis, Structural Steel might be used as a substitute for EN8C. From the graph in Figure 6(c), we can determine that as the input voltage is lowered, the values of Cast Iron and Structural Steel obtain a similar pattern of results while those of EN8C show a considerable difference of almost around 2-4 W/m3. This voltage however does not provide a suitable threshold for the molecular structure to change to a martensitic structure and moreover shows us that the substitute materials used do not follow the same trends as the base material. From Figure 7(a) given above, we can determine that the rate of temperature drop is slightly more for Structural Steel = Cast Iron > EN8C. Upon comparison of the rate of change of heat fluxin Figure 7(b), it was found out that all the materials displayed a rise in the values of heat flux until the first-time step before falling steeply. Here also, the values for Cast Iron and Structural Steel were almost constant but the least heat flux drop was in EN8C. 1234567890 IOP Conf. Series: Materials Science and Engineering 263 (2017) 062035 doi:10.1088/1757-899X/263/6/062035 With slight variations to the voltage and electric field, the output values can be different for the materials. To further observe the changes in this process, the variables constituting the non-linear analysis pattern in the analysis section must be changed to include other analyses such as impedance analysis with an increasing accuracy of the finite element analysis itself. It was also found out that with a decrease in the input voltage, the case depth of the material decreased from the front (teeth) side and remained almost constant from the back side. This means that any change in the voltage should be undertaken only upon 1234567890 IOP Conf. Series: Materials Science and Engineering 263 (2017) 062035 doi:10.1088/1757-899X/263/6/062035 the specifications for the surface hardness and case depth being explicitly mentioned and tested after the process is completed. This matches with the inference made by Castallenos [11] \u201cThe overall efficiency of these systems is about 94% (near the resonance). The losses in the rectification stage can be assumed as 1% of the total efficiency. While remaining 5% may be related to inverter stage. In extreme cases, losses can be associated with an additional 5%\u201d. Furlani [12] has also verified these claims \u201cmaximum temperature values achieved in some nodes along a radius of the specimen and their correlation with the corresponding experimental final structures. In spite of the high heating rate at the surface of the specimen, the very high temperatures reached (about 1250 \u00b0C) in this zone are able to completely transform the external layer in austenite. The subsequent cooling phase induces in the material a complete martensitic micro-structure At intermediate zones, the maximum temperatures reached are in 600-700 \u00b0C interval, so that the transformation from perlite to austenite is almost complete (as well as the consequent austenite to martensite transformation), while the ferritic phase cannot transform itself into austenite.\u201d 4.Conclusions With increasing differences in the shape of the surface that has to be hardened due to the wide variety of gears used in the market today, a multiple frequency pattern of induction hardening would be considered more productive as the output time can be reduced thus pushing out more components per batch. This would mean the incorporation of more number of coils in the heating process with each coil being supplied a set voltage in a different frequency and used to heat a different section of the component. This would enable the different sections of the component to come out with different hardness values based on the demand of the customer. Also, due to the constraints of having properties similar to the standard material, only two alternative materials have been analyzed through this simulation. However, any other material with similar composition can be included in this analysis. To determine the viability of the substitute to the standard material, further testing has to be conducted and various other parameters have to be verified to put this substitute into large scale production. Some advances in IH process as mentioned by O. Lucia [13] such as \u201cOne of the issues for the future of IH is the load adaptive capabilities and some solutions have been proposed. An adaptive simmering control of the temperature for a domestic induction cooker is required. Parameters are updated online, depending on the estimates provided by a multiple-model reset observer (MMReO). This observer consists of a reinitialized reset observer and of multiple fixed identification models.\u201d" + ] + }, + { + "image_filename": "designv6_24_0001146_rfit.2007.4443911-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001146_rfit.2007.4443911-Figure6-1.png", + "caption": "Fig. 6. HFSS model of 5.8 GHz heatsink antenna with fins parallel to the radiating edges.", + "texts": [ + " Unlike a symmetrical pinfin heatsink, the extruded-fin heatsink can have two orthogonal orientations: 1) fins parallel to the nonradiating edges (fins PNRE) of the patch and 2) fins parallel to the radiating edges (fins PRE). This orientation must be considered. Two heatsink antennas (one of each orientation) were fabricated and simulated. These antennas use the same layout in Fig. 2. The length and width of the square patch are both ~11.8 mm. The heatsinks have 3 fins each having a height of 12.5 mm (not including the base thickness which is ~2.4 mm). The fabricated heatsink antenna with fins parallel to the non-radiating edges is shown Fig. 5 while the HFSS model of the opposite orientation is shown in Fig. 6. The measured return loss for the patch antenna and the two orthogonally-oriented heatsink antennas is shown in Fig. 7. The heatsink altered the input impedance of the basic patch, so a tuning stub was used to match the heatsink antennas. The figure shows considerable bandwidth improvement over that of the patch antenna. Table II compares the bandwidth and other parameters of the three antennas. As shown in Table II the peak directivity for the two different heatsink orientations is quite different" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000284_s10846-005-0932-y-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000284_s10846-005-0932-y-Figure4-1.png", + "caption": "Figure 4. Refraction phenomena due to beam-splitter.", + "texts": [ + " From the right triangle (O1AB), it results: tan a = AB O1A \u21d2 O1A = AB tan a \u21d2 OB + b 2 = b/2 tan a \u21d2 OB = b 2 tan a \u2212 b 2 . (27) In this part, the influence of refraction phenomena due to mirror (1) of PSVS is examined. It is desirable, the left and right view of a scene captured by means of PSVS to coincide and to have exactly the same magnification. The second virtual camera, due to refraction phenomena to beam-splitter, undergoes a parallel displacement to the optical axis of the real camera. Simultaneously the optical center O2 shifts on the virtual optical axis (Figure 4). To accurately calculate paths of a light beam in two different directions, created by these two virtual cameras, the displacement of the virtual axis and the shifting of the virtual optical center O2 must be calculated. Using Snell law [20] the refraction angle of a light ray from the optical center O, along the optical axis, is the following: \u03b8r = sin\u22121 ( nair nglass \u00b7 sin \u03b8i ) , (28) where \u03b8i is the incidence angle. If \u03c91 is the angle formed by the optical axis with mirror (Figure 4), then \u03b8i = 90\u25e6 \u2212 \u03c91. If d is the mirror (1) thickness the displacement m of the second virtual camera, after some simple trigonometric calculations, can be calculated as m = d cos(\u03c91 + \u03b8r) cos \u03b8r = d sin(\u03b8i \u2212 sin\u22121((nair/nglass) sin \u03b8i))\u221a 1 \u2212 (n2 air/n 2 glass) sin2 \u03b8i . (29) The shifting of the optical center O2, l is calculated as the difference in distance of a ray from the optical center O2 until mirror (4), when this ray is radiated through mirror (1) with refraction and without refraction", + " By the solution of the previous systems the following equations, providing coordinates of a point in a 3D space are calculated: x = (uR \u2212 uRo) auR (z \u2212 zoR) + xoR or x = (uL \u2212 uLo) auL (z \u2212 zoL) + xoL, (37) y = (vR \u2212 vRo) avR (z \u2212 zoR) + yoR or y = (vL \u2212 vLo) avL (z \u2212 zoL) + yoL, (38) z = auL(uR \u2212 uRo)zoR \u2212 auR(uL \u2212 uLo)zoL + auLauR(xoL \u2212 xoR) auL(uR \u2212 uRo) \u2212 auR(uL \u2212 uLo) . (39) To find the form of final equations, adapted to PSVS, some simplifications were made. These simplifications are: xoL = \u2212b 2 \u2212 m \u2212 l, yoL = 0, zoL = \u2212b 2 , xoR = b 2 , yoR = 0, zoR = \u2212b 2 , (40) where the parallel displacement m and the shifting l (Figure 4) due to refraction phenomena in mirror (1) are given from Equations (29) and (30). Substituting the equal values from (40), Equations (37)\u2013(39), are simplified to: x = 1 auR ( z + b 2 ) (uR \u2212 uRo) + b 2 or (41) x = 1 auL ( z + b 2 \u2212 l ) (uL \u2212 uLo) \u2212 b 2 \u2212 m, y = 1 avR ( z + b 2 ) (vR \u2212 vRo) or y = 1 avL ( z + b 2 \u2212 l ) (vL \u2212 vLo), (42) z = auRauL(b + m) + auRl(uL \u2212 uLo) auR(uL \u2212 uLo) \u2212 auL(uR \u2212 uRo) \u2212 b 2 . (43) Equations (41)\u2013(43) are more simplified, if the origins of images from the two different views coincide and the scale factors horizontally and vertically are equal" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002574_s0266-3538(02)00044-1-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002574_s0266-3538(02)00044-1-Figure3-1.png", + "caption": "Fig. 3. Pole prototype: geometry, boundary conditions and loading scheme for the run numerical simulation.", + "texts": [ + " Three orthogonal co-ordinate systems are adopted throughout the analysis, precisely: (X,Y,Z) denotes the external absolute reference, (r,s,t) and (a,b,c) denote two different local FE systems. The former is used to build the standard FE local operators, the latter to handle the orthotropic behaviour of all layers and it is strictly related to the fibre-orientation angle. In the 2D-FEs, used for the wound fibre layers, the c axis is always oriented tangentially to the fibre-axis, the other two being orthogonal to the former. Instead, in the 3D-FEs, used for the=0 fibres arrangement, the two local systems are obviously coincident. In Fig. 3 the averaged geometrical dimensions of the analysed specimens are given together with the boundary conditions and the loading scheme able to reproduce the experimental test (initial specific ring stiffness determination for the 80 mm pole specimen). The analysed pole specimen was modelled with 576 3D-solid elements and 384 2D-shell elements (refer to ADINA user manual [10] for details on the FE characteristics). The above mesh was obtained by 32 subdivisions along the hoop direction, 6 subdivisions along the pole-axis and considering, obviously, as many elements-layers as the pole plies, five in fact" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001939_2015-01-0088-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001939_2015-01-0088-Figure1-1.png", + "caption": "Figure 1. Topology Optimization Results", + "texts": [ + " The Optimization Problem Statement: \u2022 Objective (What do I want?) \u2022 Design Variables (What can I change?) \u2022 Design Constraints (What performance targets must be met?) Note: The functions f(x), gi(x), can be linear, non-linear, implicit or explicit, and are continuous To obtain equal stiffness on both the legs, topology optimization has been carried out using Altair OptiStruct FE Solver [2]. From the results of optimization it has been observed that the design proposed by the topology optimizer is not feasible for manufacturing as shown in Fig. 1. Hence, we have gone for below approach. Generally any transmission has 2 types of Forks i.e. Symmetrical and Un-Symmetrical Forks. For Symmetrical Forks, whose legs are same length - It's very simple to arrive at required Stiffness. For UnSymmetrical Forks, we have followed the below mentioned approach. From the above equation I2, will be obtained by modifying the cross section using CAD tool (Pro-E Creo 5), taking care of packing and manufacturability constraints of minimum thickness for PDC Casting of Aluminum" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002479_iccit.2009.24-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002479_iccit.2009.24-Figure3-1.png", + "caption": "Figure 3. Inter connection structure and graph of lantern application", + "texts": [ + " The computed NC Set are stored in a list for quickly been addressed. Requests sent to controller are from four directions. Every direction has a 3-bit request code. \u201c000\u201d means no request; \u201c111\u201d means broadcast; \u201c001\u201d,\u201d010\u201d,\u201d011\u201d means 3 kinds of unicast sending from this direction; \u201c100\u201d,\u201d101\u201d,\u201d110\u201d means 3 kinds of group cast sending from this direction. Directions are scheduled in round-robin way. Sometimes we need to deal with more complex situations. Then a special topology is built. We give a topology as shown in Figure 3. (a). It\u2019s converted into a FG shown in Figure 3. (b). Because of the graph\u2019s shape is similar to a lantern, we call this structure lantern application. The lantern FG can be expressed as: ( , )B B BG N E= , 0 1 0 1 2 3{ , , , , , }BN P P K K K K= , 0 0 0 1 0 2 0 3 1 0 1 1 1 2 1 3 {( , ), ( , ), ( , ), ( , ), ( , ), ( , ), ( , ), ( , )} BE P K P K P K P K P K P K P K P K = . From graph theory, we can get the subgraph number of a graph is 2n, n is the number of edges. So GB has 28=256 subgraphs. Every subgraph represent a kind of connection state in lantern structure" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002010_jae-2008-999-Figure14-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002010_jae-2008-999-Figure14-1.png", + "caption": "Fig. 14. Conceptual figure of the velocity at the rotating circle.", + "texts": [ + " This theory and result are verified from the experiment data as shown in Table 2. Also, this experimental data validates the reasonableness of the CM. Although, twenty-two numbers of teeth of comb-tooth structure were the fastest, we will deal with sixteen teeth for the comb-tooth structure from now. This is because over sixteen is hard to make and expensive. The relation between the velocity to the tangential direction at a circle v[m/s] and the rotating number of circle per one second f [Hz] is expressed by Eq. (89) from Fig. 14. f = v/ (2\u03c0 \u00d7 r) (89) where, r: the radius from the sampling point to the center of a circle [m] v: the velocity to the tangential direction at a circle [m/s]. Hence, Eq. (90) is proposed to simulate the rotating number of the motor per one second fM [Hz]. fM = (vx\u2018)contact 2\u03c0 \u00d7 r (90) where, r: the radius from the sampled node to the center of a motor [m]. As mentioned in Section 5.5, we sampled three peak points, which were the largest ones to the normal z direction. Hence, the rotary speed of a RUSM at k-th point is expressed by: (RPMM )k = (fM)k \u00d7 60 (91) where, RPM M : the revolution per minute of the motor [rpm] k: k-th the sampled peak point" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003168_6.1978-1006-Figure13-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003168_6.1978-1006-Figure13-1.png", + "caption": "Figure 13 Long Scarf Nozzle Weld Joint Configuration", + "texts": [ + " Any other joint configuration was discovered to be subject to material thicbess variations which were beyond the ability to be welded using any known technique. The resulting weld joint can have one of two configurations. Use of a preplaced segment of weld filler wire (consumable insert) is permissible for those joints which canuot be fit to a gap of less than .005 inches. For those weld joints which can be fit to gaps smaller than .005 inches, the preplaced weld filler wire insert is normally not employed. Figure 13 illustrates typical weld joints with and without an insert. Fitup of the seal cylinder to rim flange can only be accomplished after the nozzle is welded to the rim. The actual fitup to this joint proceeds in a manner which is very similar to the manner discussed previously and needs no further elaboration. Fitup of short scarf nozzles is somewhat easier due to the lesser scarf angle involved with these configurations. In this instance, i t is possible to keep the weld joint in the nominal plane of the exit rim and obtain a weld joint which is weldable using conventional techniques" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001573_j.apm.2014.11.058-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001573_j.apm.2014.11.058-Figure5-1.png", + "caption": "Fig. 5. Model of magnetic fluxes flowing into a node.", + "texts": [ + " For a permanent magnet with parallel magnetization, magnetomotive force E in radial direction Ei,j and circumferential direction Fi,j occurring in layers 2 and 6 can be respectively expressed as: Ei;j \u00bc lm 2 \u00bdHc cos hi ; \u00f08\u00de cite this article in press as: Y.-C. Wu, B.-S. Jian, Magnetic field analysis of a coaxial magnetic gear mechanism by two-dimensional lent magnetic circuit network method and finite-element method, Appl. Math. Modell. (2014), http://dx.doi.org/10.1016/ 2014.11.058 Please equiva j.apm. Fi;j \u00bc lf \u00bdHc sin hi : \u00f09\u00de According to Kirchhoff\u2019s Circuit Law, the sum of the magnetic fluxes flowing into a node is equal to the sum of the magnetic fluxes flowing out of that node due to the flux continuity, as presented in Fig. 5. The following equations can be obtained: /i 1;j \u00fe /i\u00fe1;j \u00fe /i;j 1 \u00fe /i;j\u00fe1 \u00bc 0; \u00f010\u00de /i 1;j \u00bc Qi 1;j \u00f0Ui 1;j Ui;j \u00fe Fi 1;j\u00de; \u00f011\u00de /i\u00fe1;j \u00bc Qi;j \u00f0Ui\u00fe1;j Ui;j Fi;j\u00de; \u00f012\u00de /i;j 1 \u00bc Pi;j 1 \u00f0Ui;j 1 Ui;j Ei;j 1\u00de; \u00f013\u00de /i;j\u00fe1 \u00bc Pi;j \u00f0Ui;j\u00fe1 Ui;j \u00fe Ei;j\u00de; \u00f014\u00de where /i,j is the magnetic flux flowing into a node, and Ui,j is the magnetic potential of a node. By substituting (11)\u2013(14) into (10), it yields: Q i 1;j \u00f0Ui 1;j Ui;j \u00fe Fi 1;j\u00de \u00fe Qi;j \u00f0Ui\u00fe1;j Ui;j Fi;j\u00de \u00fe Pi;j 1 \u00f0Ui;j 1 Ui;j Ei;j 1\u00de \u00fe Pi;j \u00f0Ui;j\u00fe1 Ui;j \u00fe Ei;j\u00de \u00bc 0: \u00f015\u00de From (15), we can get 7N linear equations with 7N unknowns of magnetic potential" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002148_kem.620.318-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002148_kem.620.318-Figure3-1.png", + "caption": "Fig. 3 Brake shoe force analysis(forward)", + "texts": [ + "51KN, the coefficient of friction between the brake shoe and wheel K=0.27, the angle of beam chute \u03b2=12\u00b0, wheel diameter OA=OB=420mm, AC=99.5mm, CD=15mm, CD\u2019=51mm (The length of CD in wheel backward revolving), the brake shoe arc L=362mm. According to the theoretical mechanical model, we can derive the pressure distribution acted on the brake shoe during the forward and backward direction in train braking. When the wheel forward rotates and the brake beam weight is excluded, the brake shoe force analysis shows in Fig. 3. According to the principles of force and moment balance, the equations can be shown as follows : 1 2 3 42 cos sin sin 0XF F F F F (1) 2 3 4sin cos cos 0YF F F F (2) 3 4 1 2 sin 0OM F OB F OD F OC (3) 3 2F F K (4) According to Eq. 1 to Eq. 4, we can obtain: F2=22.8903KN, F3=6.01804KN, F4=9.6568KN, \u03b8=8.3678\u00b0. It can be clearly seen that the acting point of force F2 and F3 deviates from the shoe center 3.63 \u00b0 upward, and the equivalent linear distance d is 26.62mm. Assuming the pressure between the wheel and the brake shoe is a linear distribution, the pressure distribution on the shoe can be shown in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003005_j.physe.2019.05.003-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003005_j.physe.2019.05.003-Figure1-1.png", + "caption": "Fig. 1. A schematic diagram of the proposed device.", + "texts": [ + " T effect of both induced ac-field and applied magnetic field. This paper is formed from four sections, first of all was the introduction, in Section 2, the proposed model and its theoretical formalism have been introduced. The numerical results for the spin transport characteristics of single layer graphene superlattice are presented in Section 3. Finally, the conclusion with a summary presented in Section 4. In this section, the schematic diagram of the present investigated spintronic field effect transistor (Fig. 1) is described as follows: a single unit cell of N superlattice periods that consist of (2N-1) single layer normal graphene strips with 2N single layer ferromagnetic graphene strips interlock. Graphene can be changed into a ferromagnetic graphene by depositing a series of ferromagnetic insulator, EuO strips on the top of the single layer graphene superlattice sheet with metallic gate above it (Fig. 1); magnetic exchange energy of 5meV can be induced into this graphene superlattice sheet [39\u201343]. Also these strips cause a proximity effect splitting of the electronic states in graphene superlattice [39\u201343] which originates a superlattice with a spin-dependent potential profile. The source and drain leads are metals and substrate is SiO2. The spin polarization transport is conducted in the presence of an induced ac-field and magnetic field. Photon-assisted conduction channels could be introduced by the oscillating ac-field that can be adjusted by the gate voltage to set it in the conduction window of that nanodevice [40,44\u201347]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000230_t-ed.1985.22174-Figure9-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000230_t-ed.1985.22174-Figure9-1.png", + "caption": "Fig. 9. Schematic diagram of the lambda lines and load lines in case C.", + "texts": [ + " The condition in subcase B1 is ZI > (VDD - Vp)/R, - ZpD. ( W Subcase B2: UcEAz > V p There is only one true intersection point of Class 111. The condition is 1 < ( V D D - VV)/R2 - 1,. 035) Subcase B3: Vp < VcEAI < Vv. The artificial intersection point is just the true intersection point of Class 11. Only one such true intersection point exists. The condition in subcase B3 is the same as that in subcase A3. 3) Case C: The magnitude of the slope in the load line is just equal to that of the X line, i.e., Rz = Iml. This case is indicated in Fig 9. i) ii) iii) Subcase C1: There is no artificial intersection point and the peak point is in the upper side of the load line. This results in a true intersection point of Class I, as shown in Fig. 9. The condition in subcase C1 is the same as that in subcase B1. Subcase C 2 : There is no artificial intersection point and the peak point is in the lower side of the load line. This results in a true intersection point of Class HI. The condition in subcase C.? is the same as that in subcase A 1 . Subcase C3: In this subcase, the peak point is just located in the load line. There are infinite number of artificial intersection points which are also the true ones of Class 11. The condition for subcase C3 can be written as Z1 (VDD - Vp)/R, - 1," + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000566_ijhvs.2019.102689-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000566_ijhvs.2019.102689-Figure7-1.png", + "caption": "Figure 7 Schematic of stroke amplification mechanism", + "texts": [ + "3, according to formula (9). In the case of no stroke multiplier, demand pressure value could not be reached in the limited vertical installation space of 25 mm. Besides, considering the hysteresis and nonlinear properties for a pneumatic system, the 25 mm compression process was not enough. Therefore, an amplification mechanism is designed for enlargement and to convert the vertical vibration into a horizontal movement for easier installation too. The designed stroke amplification mechanism is as shown in Figure 7. The upper part of the rod is connected to the frame with a sliding key, and the lower part takes the hinge connection with the slider. During the slider moving vertically, the crank pin in the curved hole drives the crank, the piston rod being driven to do compression movement. The body can avoid the \u2018dead centre\u2019 effectively. A multi-body simulation model of the mechanism is creating in the ADAMS software, as shown in Figure 8. According to Standard QCT545-1999, the rod is defined as a step movement of 25 mm within 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000206_tmag.2013.2242087-Figure16-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000206_tmag.2013.2242087-Figure16-1.png", + "caption": "Fig. 16. Open-circuit magnetic flux density distribution under three magnetization patterns: radial, parallel, and Halbach.", + "texts": [ + " Self-inductance and mutual inductance are functions of the rotor position in those electric machines with rotor saliency. Increasing implies decreasing rotor saliency, which results in limited variations of self-inductance and mutual inductance. Fig. 15 shows self-inductance and mutual inductance for three different values of . As shown in Fig. 15, self-inductance and mutual inductance become constant (independent of rotor position) in the case of machines with surface-mounted magnet, i.e., . Fig. 16 illustrates the analytical results of the open-circuit magnetic flux density distribution under three different magne- tization patterns: radial, parallel, and Halbach. The magnitude and direction of the magnetic flux density are indicated by color and arrows, respectively. As mentioned before, the FEM is employed to evaluate the proposed analytical expressions. Fig. 17 shows the open-circuit magnetic results using finite element analyses for the radial magnetization pattern. Considering identical assumptions for the analytical and numerical results, taking enough number of harmonics into account in the case of analytical calculations, and using the FEM package properly and efficiently should lead to similar analytical and numerical results" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002744_1.5122082-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002744_1.5122082-Figure5-1.png", + "caption": "FIGURE 5. Meridional shape and shape of the scapular apparatus 2D IMP low-flow high-pressure stage 2D IMP-0015-070-03", + "texts": [], + "surrounding_texts": [ + "When the speed 1w is minimized, the maximum possible blade height is obtained, but the scapular angles are small, the interscapular channels are long. Since friction losses are controlled by the ratio of the length of the channel to the hydraulic diameter [8], it is unclear whether such primary design is optimal.\nAn alternative approach is demonstrated by the flowing part of the low-cost impellers of the model stages of Clark (USA). The figure shows the meridional shape and shape of the RDC blading units of the low-flow highpressure stage of the model stage XXX3-Q from Clark.\nLow-flow high-pressure stage XXX3-Q has the parameters des = 0.015, T des , = 0.685, design constraints\nshaftD = 0.331 (shaft diameter of an impeller), b imp = 0.012. The characteristic dimensions of the 2D IMP XXX3-Q:\n1D = 0.481, 1b = 0.025, 2b = 0.025, impz = 15 (number blades), 1b l = 390, 2bl = 450. At the stage XXX3-Q, the\nblade height at the inlet is sharply reduced and the entrance angle of the blades is increased. Figure 4 shows the characteristics of the XXX3-Q stage according to the Clark firm data and the result of their simulation using the 8th version of the Universal Modeling Method model.\nOf the Universal Modeling Model (IDENT) program. uM = 0.463. The stage is tested at uM = 0.463 and 0.785. The characteristics of stage XXX3-Q at uM = 0.785 are simulated in a manner similar to those shown in Fig. 4. A comparison of the principles of the primary design of the Universal Modeling Method and Clark's firm is done with\n030032-3", + "the example of the stage XXX3-Q and the stage 2DI-0015-070-035 with the parameters des = 0.015, T des = 0.70,\nCharacteristic sizes of 2D IMP-0015-070-03: 1D = 0.451, 1b = 0.0447, 2b = 0.028 (width of channel), impz = 17, 1b = 20.50, 2b =470. Optimization of the shape of the blade grate on the analysis of velocity diagrams of an\ninviscid quasi-three-dimensional flow. Fig. 6 compares the speed diagrams of the impeller XXX3-Q and 2D IMP0015-070-035.\nFigure 7 shows the speed diagram of 2D IMP XXX3-Q at a flow rate ni =0.021. The ratio ni / des = 1.4 is a very large discrepancy. 2D IMP XXX3-Q has a much higher speed level and a very large blade load. In [9], the ratio of the average load to the average speed is recommended to be limited /mid midw w 0.45. In 2D IMP XXX3-Q, the load parameter is approximately /mid midw w 0.75. At the same time, the shape of the interlace channels of 2D IMP XXX3-Q is preferable to that of 2D IMP-0015-070-035 - both visually and in relation to the length to the hydraulic diameter. Fig. 8 compares the characteristics of the stage 2D IMP XXX3-Q and 2D IMP-0015-070-035 with stator elements of the stage XXX3-Q.\n030032-4", + "030032-5" + ] + }, + { + "image_filename": "designv6_24_0002010_jae-2008-999-Figure10-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002010_jae-2008-999-Figure10-1.png", + "caption": "Fig. 10. Rotation of the original coordinate axis to the tangential direction of a circle.", + "texts": [ + " The elliptical displacement was drawn according to divisional steps at one contacting point by: xi = |xcom| \u00d7 sin [ 2\u03c0 n \u00d7 i + rad (xcom) ] (80) yi = |ycom| \u00d7 sin [ 2\u03c0 n \u00d7 i + rad (ycom) ] (81) zi = |zcom| \u00d7 sin [ 2\u03c0 n \u00d7 i + rad (zcom) ] (82) where, xi, yi, zi: the value of x-, y-, z-directional displacement at the i\u2212th divisional step of an elliptical displacement xcom: the complex of a displacement to x-direction ycom: the complex of a displacement to y-direction zcom: the complex of a displacement to z-direction n: the total divided number of an elliptical motion i: from 0 to n rad(k): the radian of a complex k. The rotary speed of a RUSM is determined by the velocities of elliptical displacements, especially by the tangential ones about the circle of a motor. Hence, to simulate the RPM (Revolution Per Minute) of a RUSM, the original coordinate axis of each elliptical motion must be transformed to a new coordinate axis which is a rotated one to the tangential direction of a circle, as shown in Fig. 10. The general formulation of a transformed node value by coordinate axis\u2019s rotation is Eqs (83) and (84). Therefore, Eqs (85)\u2013(87) are derived for the calculation of transformed elliptical motion, of which coordinate axis is the rotated one to the tangential direction of a circle. X \u2032 = cos (\u03b8)X + sin (\u03b8)Y (83) Y \u2032 = \u2212 sin (\u03b8)X + cos (\u03b8)Y (84) where, X, Y : X,Y values about an original coordinate axis X\u2018, Y \u2018: the transformed X, Y values about a rotated coordinate axis. Hence, x\u2032 i = sin (\u03b8)xi \u2212 cos (\u03b8) yi (85) y\u2032i = cos (\u03b8)xi + sin (\u03b8) yi (86) zi = zi (87) where, x\u2032 i, y \u2032 i: the transformed xi, yi values about a rotated coordinate axis to the tangential direction of a circle, at the i-th divisional step of an elliptical displacement" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001026_tap.2015.2487512-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001026_tap.2015.2487512-Figure8-1.png", + "caption": "Fig. 8. Geometry of the cavity-backed slot with a non-rectangular cavity fed by the dielectrically loaded PPW (Figure 3) with Dielectric - TMM13i (\u03b5r = 12.85, tan \u03b4 = 0.0019), C = 10mils, SP = 100mils, H1 = 30mils, H2 = 270mils, and H3 = 50mils.", + "texts": [ + " Figure 7 gives some insight on the cause of this variation. Specifically, the element\u2019s resonance varies due to dielectric loading and the slot\u2019s input impedance. To reduce the power coupling sensitivity to the plunger movement, its Q value was reduced. This can be done by 0018-926X (c) 2015 IEEE. Personal use is permitted, but republication/redistribution requires IEEE permission. See http://www.ieee.org/publications_standards/publications/rights/index.html for more information. resorting to a non-rectangular cavity-backing as depicted in Figure 8. Figure 9 gives a comparison of the radiated power for the non-rectangular (\u03b1 = 45\u25e6) and rectangular cavity-backed slots (\u03b1 = 90\u25e6). It is clear from Figure 9 that the non-rectangular cavity delivers a nearly constant coupled and radiated power for g \u2265 150mils. Moreover, the percent of coupled power can be varied by adjusting the slot width, W , as depicted in Figure 10. That is, the slot width was chosen to control the amount of power coupling for the different slot antennas comprising the TWA. Figure 10 shows the desirable design properties for the cavity-backed slot antenna element" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002611_isocc47750.2019.9027698-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002611_isocc47750.2019.9027698-Figure6-1.png", + "caption": "Fig. 6. Configuration of Hall-effect sensors.", + "texts": [ + "org/publications_standards/publications/rights/index.html for more information. below the pole piece. In Fig. 5(a), all six coils were powered equally with 1-A dc current, so the Bmax point was located at the center of the working space. In Fig. 5(b), the location of Bmax was changed to the third quadrant of the coordinate system by increasing the current in coils 2, 3, and 4 (see Fig. 1), and decreasing the current in coils 1, 5, and 6, while keeping the total current \u22116 i=1 Ii the same as in Fig. 6(a). Therefore, the location of Bmax is related to the distribution of the magnetic field produced by electromagnets. If the magnetic field could be measured, the location of Bmax in the horizontal plane could be determined, which in turn determines the location of the levitating MGR. B. Installation of Hall-Effect Sensors The installation of Hall-effect sensors played a key role in the position determination technique. Fig. 6 shows three different configurations of sensor installation (inner, mid, and outer) where the distances between two Hall-effect sensors are 40, 88, and 120 mm, respectively. These three configurations were taken into consideration to optimize the performance of position estimation. The experimental results shown in Fig. 7 indicate that linear function can be fitted with \u201cinner\u201d and \u201couter\u201d installation strategy, and third-order nonlinear function can be fitted with \u201cMid\u201d installation. The results in [19] demonstrated that the \u201cinner\u201d installation strategy had the best performance in that it provided minimum position estimation error" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001721_ip-h-2.1990.0051-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001721_ip-h-2.1990.0051-Figure5-1.png", + "caption": "Fig. 5 (Lattice altered as Fig. I d ) Stability of tripole loop arrays as functions of lattice angle a1", + "texts": [ + " 5 and 6, the angle a2 is again 90\" but new a1 varies from 0\" to 30\" giving a square or rectangular grid when a1 = 0\". Here L1 = 2.5mm and L 2 = 0.6 mm. _ _ - - D1 = 6.1 mm, D2 = 3.5 mm \\\\\\\\\\\\\\ D1 = D2 = 6.1 mm, 45\" incidence 1 1 1 1 1 1 1 1 1 D1 =D2=6.lmm,80\"incidence Note change of scale for A < - 3 per cent Measurements : TM, orientation a or b 0 TE, orientation a or b 266 There are three main groups of results for 45\" incidence, together with a few results for the extreme case of 80\". 3.2.1 Lattice varied according to method ( i ) , Fig. I d : The broken curves in Fig. 5 are generated when 0 2 = 3.5 mm, giving close packing along y. To provide the comparison with the results of Fig. 6 below, D1 has been set to 6.1 mm which is the minimum feasible separation when a1 is zero. There is little variation of A with a1 but the values for TE and TM incidence for a given lattice orientation differ by about 3%. In this particular case however, the stability is lost if rotation between principal orientations at normal incidence is included. The obliquely hatched regions are generated when the lattice is again altered according to method (i) but now the element separations D1 and 0 2 are both 6", + " In contrast to the broken curves, there are now only minor differences between the results for the two orientations a and b, and these are represented by the spread of the hatched areas. The stability parameter A is everywhere negative and the curves for TE and TM intersect when a1 is about lo\" where A is less than -2%, giving a reasonably stable lattice, as confirmed by the experimental points shown. As a1 is increased the curves diverge. The TM performance deteriorates slightly to give A - -2.5% when a1 reaches 30\". Values for this limiting equilateral triangular lattice correspond to those plotted in Fig. 4 for D = 6.1 mm. 3.2.2 High angles of incidence: Fig. 5 also illustrates the effect of increasing the angle of incidence to 80\", the vertical hatching corresponds to the obliquely hatched areas for 45\". They now intersect near a1 = 25\" instead of lo\". The performance has deteriorated significantly, with IEE PROCEEDINGS, Vol. 137, Pt. H, No. 5, OCTOBER 1990 A reaching -7% for TE incidence for rectangular lattice cells where a1 = 0\". Equilateral triangular lattices corresponding to a l = 30\" give the optimum stability, with A typically - 5%, whereas at 45\" incidence small values of a1 are marginally preferable", + " However, at such high angles of incidence, the reflection band for TE is very wide (values of 25% were encountered for method (i) and up to 80% for method (ii)) and asymmetrical about the -10dB band centre. The concept of band centre stability has less significance than factors such as the width and depth of the nulls, grating responses, and the narrow TM resonance (for example, Reference 8). 3.2.3 Lattice varied as in Fig. le, method ( i i ) : a2 and 0 2 (3.5 mm) are constant but Dl now varies with al . It starts at 3.5 mm again, which is the closest practical spacing when a1 = 30\", but increases to 6.1 mm as a1 falls to zero. As for the hatched region in Fig. 5, in orientation a the curves for TE and TM incidence in Fig. 6 converge as a1 increases, now intersecting near a1 = 12\" and giving a highly stable lattice (A is almost zero). Here, Dl = 4.5 mm and 0 2 = 3.5 mm. (This intersection was also observed theoretically and confirmed experimentally at a1 = 10\" in the case of a more closely spaced array, with 0 2 = 3.0mm and L2 = 0.4mm.) But the per- IEE PROCEEDINGS, Vol. 137, Pt. H , No. 5 , OCTOBER 1990 formance is sensitive to the orientation parameter", + " For orientation b, the curve for TM incidence follows that for TEa until a1 N 12\", thereafter following TMa, while A remains at or below 1% for TEb for all a l . Nevertheless, Stability of tripole loop arrays as functions ofa2 (Lattice varied TM incidence, orientation a orb once a1 exceeds 7\", all the values of A in Fig. 6 lie within 2% of each other, and near a1 = 12\" they differ by only 1 YO. Until this latter critical angle is reached, the relationship between these four curves resembles that between the broken curves in Fig. 5 over the whole 30\" range of a l . At a1 = 30\u00b0, the unit cell is again an equilateral triangle, but the array is more closely packed than in Fig. 5 and the stability in TM is improved. Fig. 7 gives the results for the simple tripole element for the comparison with the curves for the tripole loop in Fig. 6. The lattices have again been varied according to method (ii) of Section 2. The solid curves refer to lattices which are very closely packed (02 = 3.0 mm) while the performance of arrays of tripoles set on lattices identical to those of Fig. 6 with 0 2 = 3.5 mm is shown by the broken curves. The TM response is relatively unstable at a1 = 30\" as would be expected from Fig", + "5 mm at a2 = 90\u201d to 6.1 mm at a2 = 60\u201d, while D1 is now constant at 6.1 mm. The values of A are broadly similar to those of Fig. 6, but the curves for TMa and TEb are now very close to each other, as are those for TEa and TMb. The stability of the latter pair declines as a2 increases and 0 2 decreases to give a more closely packed array. The equilateral triangular lattice at a2 = 60\u201d gives the maximum stability ( I A I < 2%) which is a marginally better performance than that shown by the hatched regions in Fig. 5 at a1 = 30\u201d, the alternative equilateral lattice which also has D = 6.1 mm. These two lattices are different. An equilateral triangular lattice with a1 = 10\u201d and D = 4.7 mm (12\u201d and 3.9 mm for tripoles) in fact gave arrays with the smallest drifts of all. 4 Discussion and Conclusions This study has concentrated on the reflection band centre stability at moderate angles of incidence as the indicator of the relative merits of a range of tripole and tripole loop FSS designs. The crosspolar performance, which would be important in some applications, has not been addressed here, although brief discussions of a limited number of cases for simple tripoles are given in References 5 and 6" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000008_icelmach.2018.8506886-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000008_icelmach.2018.8506886-Figure6-1.png", + "caption": "Fig. 6. Discrete 3D model", + "texts": [ + " The applied model includes: an aluminium support element with a water jacket (1), a simplified stator core (2), a simplified winding model (3), a thermally conductive resin filling the space between the winding and the supporting structure (4). In the CFD analysis program, the model (shown in Fig.5) was additionally supplemented with a cooling medium in the water jacket channels. The thermal resistance substitute parameters have also been assumed: Rs - thermal resistance corresponding to the pressure between the core and the water jacket construction, R\u017c - the thermal resistance corresponding to the groove insulation. Then the model was discretized. The discreet model is presented in Fig.6. All models and calculations were performed in Autodesk Inventor and Autodesk Simulation CFD software. Considering the stator design possibilities, different variants of its construction were analyzed. The analysis takes into account both the shape of the stator's supporting structure as well as the water jacket channel. For each variant an appropriate calculation model has been created taking into account the introduced changes. Fig. 7 presents calculation models of the analyzed stator support structures" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003611_is3c.2014.314-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003611_is3c.2014.314-Figure2-1.png", + "caption": "Fig. 2 Motions of the spherical motor", + "texts": [ + " The magnetic field is generated in the 3D actuator space by the PM poles mounted on the rotor\u2019s surfaces. It can interact with the current inputs in the stator coils, and produce torque vectors in the space to move the rotor in 3-DOF. One set of coils can be energized to create tilting motions of rotor in x direction, and another set in orthogonal direction to create the second tilting motions in y direction. Rotation 1 in Fig.4 presents the two similar tilting motions. The rest coils can be used to create spinning motions about the rotor axis, i.e., Rotation 2 in Fig. 2. Magnetic field of the spherical actuator is a vector field in 3D space. To formulate a vector field, we should have the divergence and the curl of the field, i.e., B and B , where is the magnetic flux intensity. The steady magnetic field is essentially a source free field, i.e., 0 B . )1( From the Helmholtz\u2019s Theorem, there must be a magnetic vector potential field A that can satisfy the equation of 0)( A . Hence, the magnetic flux intensity B can be expressed as AB . )2( Because the actuator has two layers of PM poles distributed in 3D space, the magnetic circuit of the motor is complicated" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000759_ijex.2016.075604-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000759_ijex.2016.075604-Figure1-1.png", + "caption": "Figure 1 Schematic diagram of the twin-spool turbofan engine (see online version for colours)", + "texts": [ + " CEP, overall efficiency, exergy destruction factor, thrust-specific fuel consumption (TSFC) and entropy generation rate of the engine have been used as the performance indicators. Effects of heat leakage on the variation of these performance parameters with respect to the design parameters of by-pass air ratio, fan pressure ratio, total compressor pressure ratio and maximum turbine inlet temperature have been investigated numerically. Twin-spool turbofan engine considered in this study is shown schematically in Figure 1. Processes, heat leakages, fuel, bleed air and cooling air mass flows are also illustrated on a T-s diagram of the model in Figure 2. 2.1 Processes and assumptions Processes 1\u20135 and 6\u20139 are irreversible adiabatic compression and expansion processes, respectively. Processes can be expressed as follows. \u2022 1\u20132 and 2\u20133: Air entering at ambient conditions is slowed down and compressed in a diffuser and then in a fan. \u2022 3\u20134: Part of this air is compressed in a low-pressure compressor (LPC). \u2022 Bleed air extraction: The air leaving the LPC is split into two parts: the first part (called bleed air) is taken out to provide anti-icing and in-cabinet air pressure, to drive the flap system and also for the environmental control unit requirements" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002278_iceaa.2010.5649999-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002278_iceaa.2010.5649999-Figure1-1.png", + "caption": "Fig. 1. Antenna layout, top view. The signal paths are shown in dark grey, while the shielding traces are light grey. The vias are not shown. The large rectangular frame around the patch array represents a via fence wall structure to reduce substrate wave excitation.", + "texts": [ + " Numerical full-wave simulations were performed with Ansoft HFSS [6] for rectangular slots, H-shaped and dog-bone shaped slots. Together with the proposed patch design, rectangular slots revealed the widest possible impedance bandwidth. As a result, the simulations exhibited the rectangular slot to be least sensitive to manufacturing tolerances. The array is fed by 36 \u2126 striplines, as the realization of 50 \u2126 striplines would result in structural resolutions of less than 60 \u00b5m, which proved not to be reliable enough for buried layers within the constraints of the materials and processes used. As illustrated by Fig. 1, all striplines are screened by shielding via fences, suppressing the excitation of unwanted parallel-plate modes as well as minimizing mutual coupling between adjacent transmission line sections. The maximum via distance was 350 \u00b5m, corresponding to about 1/5 of the guided wavelength. The layer configuration is depicted in Fig. 2. Impedance matching of the stripline to the slot was achieved by an open stub, according to the geometrical details in Fig. 3. Shorted stubs were simulated as well, however, they generally provide narrower impedance bandwidths. To surround the coupling slot entirely, the via fence had to be widened around it (Fig. 1). The resulting substrate-filled metal-shielded cavity exhibits a resonance near the target frequency band of the antenna. The dimensions of this rectangular cavity were carefully tuned to shift its resonance outside of this frequency range. Radiating patches on dielectric substrates tend to excite parasitic substrate waves, which generally 1-5 Institute for Micro- and Nanotechnologies, Ilmenau University of Technology, P.O.Box 10 05 65, 98684 Ilmenau, Germany, e-mail: 1frank.wollenschlaeger@tu-ilmenau", + " 4, the via fence was capable of keeping the antenna gain at levels above 10 dBi up to 65 GHz, while without the via fence the gain rapidly decreased at frequencies above 60 GHz. Thus, this structure effectively maximizes the gain bandwidth of the antenna. The most critical dimension of the via fence turned out to be its distance to the radiating edges of the patches, which was set to roughly a quarter of the free-space wavelength. To achieve the required gain of 12 dBi, a two-bytwo array was chosen. The distance between the patches was optimized to yield maximum gain for the overall structure. The resulting patch layout is depicted in Fig. 1. The feed network was designed for MMICs with differential microwave interconnects. Thus, a pair of patches located in the same H-plane is fed with 180\u00b0 phase difference with respect to the opposing pair, resulting in a constructive superposition of theradiated waves. For the power dividers, a quarter-wave matched T-junction structure was chosen. Wilkinson power dividers were considered initially but eventually abandoned, as the tiny increase of the impedance bandwidth did not justify the added complexity of the printing of resistive pastes in the buried stripline layers" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002224_3527608117-Figure11.2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002224_3527608117-Figure11.2-1.png", + "caption": "Fig. 11.2 Comparison of fracture toughness for conventional and spray formed alloys [3].", + "texts": [], + "surrounding_texts": [ + "P. Krug, G. Sinha 11.1" + ] + }, + { + "image_filename": "designv6_24_0000855_0954406220936732-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000855_0954406220936732-Figure6-1.png", + "caption": "Figure 6. Axial force of limbs of flexible force sensor.", + "texts": [ + " Therefore, the axial deformation of flexible link is caused by axial tension/ compression only. According to Saint Venant\u2019s principle, for the flexible links, the stress distribution near the flexure hinge is relatively complex. However, since the middle axial part of the link is far from the load area, the stress is only related to the load resultant force, that is the axial tension/compression. Therefore, the middle axial part of the link is the best potion to place the strain gages. The axial reaction force of the measuring limb of the flexible parallel force sensor is shown in Figure 6. Given the stiffness matrix of the force sensor K and the six-dimensional deformation of the moving platform X, the external force F can be calculated as follows F \u00bc KX \u00f018\u00de If K is invertible, then X \u00bc CF \u00f019\u00de where C\u00bcK 1 is the flexibility matrix of the force sensor. Based on the geometric compatibility condition of deformation between the dynamic platform and the end of each flexible series branch, the displacement\u2013 displacement relationship between the moving platform and the end of each series branch Xi \u00bc J 1i X \u00bc J 1i CF \u00f0i\u00bc 1, 2, " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003536_978-94-009-5063-4_1-Figure16-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003536_978-94-009-5063-4_1-Figure16-1.png", + "caption": "Figure 16. Stress distribution in plate before and after buckling (adapted from Brush and Almroth [4)).", + "texts": [ + " In tests of actual plates, which of course contain unavoidable imperfections, it is difficult to detect the onset of buckling because it happens gradually, as one might expect from Fig. 15 (b). The bifurcation point on the load-deflection curve for the perfect plate does not correspond to failure of the structure, but indicates a load at which the perfect plate starts to bow laterally. With further increase in uniform end shortening above the bifurcation value, the axial compressive stress resultant N x begins to redistribute, becoming .. more and more concentrated near the edge supports, as depicted in Fig. 16 (b). The stress resultant distributions across the width of the plate at a certain axial station x are shown for four values of axial compression by the curves 1-1,2-2,3-3, and 4-4 in Fig. 16 (b). At bifurcation the stress resultant is uniform and equal to Ncr. Near the edges the axial fibers are straighter than they are near the middle. Therefore, the end shortening is accomplished primarily by membrane compression, resulting ,in a large Nx ' Near the midwidth the same end shortening is accomplished primarily by bending, resulting in a small Nx . The behavior would be qualitatively similar if the plate were compressed by uniform axial load rather than uniform end shortening: The regions in the neighborhood of the ends x = 0 and x = a remain fairly straight because of the restraint against lateral displacement w there", + " As the post-bifurcation lateral displacement in the central region increases the edge regions at x = 0, a act as webs which, through shear, transfer the load away from the central region of the plate to similar effective axially oriented beams near the edges at y = \u00b1 b /2. These effective axial beams carry most of the compressive load. Approximate maximum loads for axially compressed stiffened plates are derived for design purposes from the so-called von Karman effective width formula [27] : (9) in which beff is an effective width shown in Fig. 16(c) over which one can assume that the load in the plate is carried, and N max is the maximum stress resultant that can be carried because of yielding or some other stress failure criterion. In the past 'the effective width formula (9) was used to calculate maximum bending moments carried by airplane fuselages and wings. Similar design procedures have been developed for plates subjected to inplane bending or shear loading. A comprehensive discussion of the ultimate strength of plates in bending, shear, and combined bending and shear is given in Chapter 5 of Ref" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002482_9783527631506-Figure5.23-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002482_9783527631506-Figure5.23-1.png", + "caption": "Figure 5.23 Current\u2013voltage characteristics of (a) BJT and (b) MOS transistors pertaining to implementation of resistors.", + "texts": [ + "22b, respectively, show the geometrical structures of epi and ion-implanted type of resistors. 5.10.1.4 Active Resistors Active resistance implies resistance obtained from active devices, that is, transistors. Considering the I\u2013V characteristics of well-known active devices, that is, BJT and MOS transistors, one can easily see that the devices exhibit the behavior of a resistance in certain areas of the I\u2013V characteristics. For the BJT, it is the saturation zone, while for the MOS it is the linear zone. This is illustrated in Figure 5.23. The advantages of an active resistance compared with a passive semiconductor resistance are twofold: (i) the available resistance can be very large without requiring large area of the substrate and (ii) the value of the resistance can be changed by a control voltage or current. In the modern MOS and CMOS technology, MOS transistors working in the linear region are frequently used to produce a resistance. Considering the N-type metal-oxide semiconductor (NMOS) transistor in Figure 5.24a, the instantaneous (i" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002755_tvlsi.2012.2227848-Figure10-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002755_tvlsi.2012.2227848-Figure10-1.png", + "caption": "Fig. 10. Bulk FinFET 6T SRAM (111) configuration. (a) (FEOL+BEOL) and (b) FEOL only. Dielectric regions are not shown.", + "texts": [], + "surrounding_texts": [ + "Owing to the width quantization property, multigate FETs with large electrical widths need to have multiple fins. We synthesized multifin FinFETs using the bulk and SOI FinFETs generated earlier at the 22-nm/14-nm/10-nm nodes. They consisted of four fins each, with shared raised source/drain epiregions that are via-contacted and connected using metal-1, as shown in Fig. 8(a) and (b). We varied the fin pitch, FP, which is the distance between the centers of consecutive fins, and computed the parasitic (FEOL+BEOL) capacitances for each layout using the setup described in Fig. 7(b). From Fig. 9(a), we can see that the trends in CDRAIN,TOT are in stark contrast to the single-fin results in Section III. While moving from SOI to bulk FETs, there is a 11.5%, 10.8%, and 8.8% increase in CDRAIN,TOT for the 22-, 14-, and 10-nm nodes, respectively, which can be attributed to the shared drain-to-bulk fin capacitances in bulk FETs. However, in the case of CGATE,TOT [Fig. 9(b)], there is only a 2%\u20134% increase from SOI to bulk FETs. An increase in FP from 40 to 70-nm results in a 20%, 31%, and 36% increase in CDRAIN,TOT for the 22-, 14-, and 10-nm nodes, respectively, while CGATE,TOT increases by 16%, 26%, and 28%, respectively. These results suggest that gate-toepi-source/drain/metal-1 capacitances begin to dominate as FP increases or the technology node decreases, and they highlight the need to model the entire (FEOL+BEOL) structure." + ] + }, + { + "image_filename": "designv6_24_0000807_msf.636-637.1105-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000807_msf.636-637.1105-Figure2-1.png", + "caption": "Fig. 2. Principal and cylindrical coordinates.", + "texts": [ + " Layers are assumed to be perfect and flawless and there is complete bond between them, i.e. the displacements are continuous across the layers. (r,\u03b8,Z) are the cylindrical set of coordinates and ur, u\u03b8 and uZ are radial, circumferential and axial displacements respectively (Figure 1). Fig.1. cylindrical vessel under internal pressure (r0,ra inner and outer radius respectively). The (x,y,z) coordinates called principal coordinates in which x is along the fiber in each layer and y and z normal to it with z along the radius of the cylinder (Figure 2). The governing equations of elasticity are written in the first coordinate while the mechanics of laminate are written in the principal coordinates for each layer.The notation and formulation are as in reference [7]. All rights reserved. No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of Trans Tech Publications, www.ttp.net. (ID: 128.6.218.72, Rutgers University Libraries, New Brunswick, USA-31/05/15,19:34:08) By these assumptions and the symmetry, the stress and strain components are independent of \u03b8 and Z", + " Assuming transverse isotropy for the lamina, we can write: )1(2 ,,, zy y xxyxzxzzyyzy E GGGEE \u03c5 \u03c5\u03c5 + ==== (4) The elasticity constants in principal coordinates ijC (i,j=x,y,z) can be expressed by: 1 1 1 1 \u2212 \u2212\u2212 \u2212\u2212 \u2212\u2212 = zy zy z zx y zy yx xy z zx x xy x zzyzxz yzyyxy xzxyxx EEE EEE EEE CCC CCC CCC \u03c5\u03c5 \u03c5\u03c5 \u03c5\u03c5 (5) It should be mentioned that equation (5) is expressed erroneously in reference [6] and the correct form above is from references such as [7]. The relation between elasticity constants in the two systems of coordinates is given by: { } [ ]{ })()( k ijkl k ij CAC = (6) where { } { })( 66 )( 55 )( 45 )( 44 )( 36 )( 33 )( 26 )( 23 )( 22 )( 16 )( 13 )( 12 )( 11 )( ,,,,,,,,,,,, kkkkkkkkkkkkkk ij CCCCCCCCCCCCCC = { } { })()()()()()()()()()( ,,,,,,,, k zz k yy k xx k yz k xz k xy k zz k yy k xx k ij GGGCCCCCCC = (7) and the transformation matrix [Akl] based on Figure 2 is given by: )}sin(),cos({ )(000020 0000000 0000000 0000000 0000000 000000100 2200000 0000000 4000020 2200000 0000000 400000 4000020 ][ 222222222 22 22 333333 22 222244 333333 22 22442222 222244 \u03c6\u03c6 == \u2212\u2212 \u2212 \u2212 \u2212\u2212\u2212 +\u2212+\u2212\u2212 \u2212+ = nm nmnmnmnm mn mnmn nm mnmn mnnmmnnmnmmn mn nmnmmn mnnmmnnmmnnm nm nmnmnmnm nmnmnm Akl (8) Boundary conditions include the conditions of traction on inner and outer surfaces: 00 )1( )( prr \u2212=\u03c3 , 0)()( =a n r r\u03c3 , 0)()( 0 )1( 0 )1( == rr zrr \u03c4\u03c4\u03b8 , 0)()( )()( == a zra r rr \u03c4\u03c4\u03b8 (9) (N= number of layers, p0=internal pressure), continuity of stress and displacements at the interface of the layers plus the two integral conditions of axial equilibrium and zero torsion for the cylinder: 0 1 2 0 )( )(2 1 prrdrr n k r r k z k k \u2211\u222b = = \u2212 \u03c0\u03c3\u03c0 , \u2211\u222b = = \u2212 n k r r k z drrr k k1 2)( 0)(2 1 \u03b8\u03c4\u03c0 (10) Based on these conditions, the circumferential displacement becomes: rzu k 0 )( \u03b3\u03b8 = (11) There are 2N+2 remaining unknowns, namely D(k),E(k) (k=1,2,\u2026,N),\u03b50,\u03b30" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002466_tcsi.2010.2046200-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002466_tcsi.2010.2046200-Figure6-1.png", + "caption": "Fig. 6. Effect of sampling the input signal.", + "texts": [ + " A digitally generated PWM signal on the other hand has a constant and relatively low switching frequency but requires a 4- to 16-times higher sample rate for comparable signal-to-noise ratio (SNR). Hybrid techniques can be used to reduce the switching frequency of SDM signals without increasing the sample rate [7]. However, in modern CMOS technologies, the sample rates required for straightforward digital PWM generation are easily achieved. Because the PWM signal is generated in the digital domain, the effects of sampling and quantization have to be dealt with. The effect of sampling can be modeled by inserting a sampleand-hold in series with the modulating signal, as shown in Fig. 6. As can be seen, the sampled signal crosses the reference triangle at different times than the original modulating signal, causing the edges of the PWM signal to shift. Fig. 7 shows the frequency spectrum that is obtained if the input signal is sampled at the PWM carrier frequency . The spectrum now fills up with harmonics of the modulating signal. Apparently, sampling the input signal causes distortion. In a digital PWM modulator, the edges are synchronized to a high-frequency bit clock, e.g., 256 " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002417_0956-0521(92)90017-d-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002417_0956-0521(92)90017-d-Figure6-1.png", + "caption": "Fig. 6. Three-dimensional cantilever subjected to an instantaneous load.", + "texts": [ + " The response of a single degree-of-freedom linear mass-spring system has to be the starting point for the examination of any new model. Exact agreement with the standard theoretical result for a suddenly applied constant loading was found using a transit time based on the speed of propagation of sound in the medium. Physical Process Model 667 Continuous systems may be represented as a series of finite elements and in the IOPM it is natural to treat each element as an interacting object. A three-dimensional finite element was used l\u00b0 for a continuous elastic rectangular cantilever as shown in Fig. 6. For a beam of span 12.0 cm, a breadth and depth of 1.0cm, an elastic modulus E of 100.0 kN/mm 2, density of 0.01, a Poisson's ratio of 0.0 and a suddenly applied point load at the end of 1.0 kN, the theoretical maximum deflection is twice the static deflection of 6.91 mm. The dilational wave has the highest velocity at 104mm/s and so the minimum time step is the shortest distance divided by the highest velocity and is 10/104 = 0.001 s. The time step was therefore chosen as 0.0008 s and the maximum deflection obtained by the IOPM during a 300 s simulation was 13" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001395_ijvp.2016.075351-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001395_ijvp.2016.075351-Figure3-1.png", + "caption": "Figure 3 ANCF beam element coordinates", + "texts": [ + " The FFR formulation will be discussed in more detail in the following section. While the FFR formulation can be used with finite elements that employ infinitesimal rotations, ANCF finite elements have no infinitesimal or finite rotations as nodal coordinates, and therefore, there are no restrictions on the amount of rotation or deformation within the element. Instead, element nodal coordinates are defined in the inertial frame and used with a global shape function matrix that has a complete set of rigid body modes. Shown in Figure 3 is a beam element consistent with this approach which shows the position vector gradients along each axis as well as the spatial coordinates associated with each node. This element is a fully parameterised element since it has a complete set of parameters, while other elements such as the Euler Bernoulli beam element are called gradient deficient elements since complete set of position vector gradients cannot be defined. Gradient deficient elements such as the Euler-Bernoulli beam element are simpler and more efficient in many applications including chain and belt applications, and for this reason, they will be used in the numerical study presented in this paper to determine the stresses of the track links of three-dimensional tracked vehicles", + " The ANCF, which is based on a global position field description, can be used to describe an arbitrary displacement including large rotation and large deformation. ANCF finite elements define a unique rotation field and can be obtained from the position field using a general continuum mechanics description. These finite elements allow for imposing continuity on higher order derivatives without increasing the order of the interpolation or the number of nodal coordinates. Furthermore, one can develop finite element meshes that have linear connectivity and constant inertia, as previously mentioned (Shabana et al., 2012). Figure 3 shows an example of the displacement field of an ANCF finite element which can be written as 1 2 3 1 2 3( , , , ) ( , , ) ( ),j j jx x x t x x x t=r S e where 1 2,x x and 3x are the element spatial coordinates; t is time; Sj is the element shape function matrix, and ej is the vector of element nodal coordinates. Using this displacement field, the equations of a pin joint between elements i and j can be written as ,i j=r r ,i j \u03b1 \u03b1=r r where is the coordinate line defining the joint axis; can be 1 2,x x or 3 ,x or any other coordinate line" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000143_isemc.1982.7567736-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000143_isemc.1982.7567736-Figure2-1.png", + "caption": "FIGURE 2. OPTICAL FIBER STRUCTURE", + "texts": [ + " The remainder, then, of this paper discusses a new technology in electromagnetic field sensors which shows great potential for non-perturbing measurements. This new technology has been named Fiber Optic Sensor Instrumentation Research and Development (FOSIRD). Theory of Fiber Optic Sensors The operation of fiber optic sensors depends on mode prop agation in optical fibers which, in fact, are cylindrical electromag netic waveguides. Optical fibers usually consist of two concentric dielectric cylinders, one an inner core and the other an outer clad ding (Figure 2). In addition, an outer jacket.is applied to protect the fiber from moisture and abrasion. As long as the dimensions o f the core and cladding in which the light waves travel are much larger than the optical wavelength, the light can be described by the propagation o f rays or beams that are reflected or transmitted at the various boundaries of the waveguide. The ray description, although not exact, is frequently used since it is more intuitive than a wave description. used as part of the phase-locked loop" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002323_antem.2012.6262422-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002323_antem.2012.6262422-Figure1-1.png", + "caption": "Figure 1. Patch antenna geometry", + "texts": [ + " ISSDPG tells up to what extent the gain deviates from this mean value in all directions within the sector, in other words whether the radiation pattern is flat or non uniform within the sector. In case the statistical distribution of the power gain within the sector turns out to be normally distributed, these two parameters are sufficient to describe the distribution completely. III. DIRECTIONAL ANTENNA-WALL INTERACTION In the first set of simulations, we designed an air patch antenna operating from 1.7 GHz to 2.1 GHz in proximity to a wall whose characteristics are shown in fig. 1-3. Given that the power radiation pattern is getting narrow at low frequency, the sector at 1.7 GHz is selected to cover the angles \u03c6 = [30-150], \u03b8 = [55-150] in degrees according to the definition in section II. Two cases were considered, with either the antenna\u2019s ground plane perpendicular or parallel to the wall (as shown in fig.4). The electromagnetic simulations have all been carried out using CST [10] at three frequencies (f = {1.71, 1.905, 2.05} GHz). In order to simulate a generic scenario for the antennas in proximity to a wall, three parameters (relative permittivity of the wall, distance between the patch antenna and the wall, wall thickness), as summarized in table" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001500_s11661-018-4589-0-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001500_s11661-018-4589-0-Figure4-1.png", + "caption": "Fig. 4\u2014The GBCD statistics of the deformed specimens.", + "texts": [ + " In this study, the tensile direction of the tensioned specimens was parallel to the rolling direction (RD) of the rolled specimens. For the rolling deformation, the grains were elongated along the rolling direction (RD) and flattened along the normal direction (ND). In the METALLURGICAL AND MATERIALS TRANSACTIONS A case of the tensile deformation, the grains were elongated along the tensile direction (or RD) and flattened perpendicular to the tensile direction (or RD). Figure 3 shows the grain boundary reconstruction maps of the deformed specimens, and the GBCD statistics are summarized in Figure 4. Compared to the BM specimen, the specimens of R5 and T6 revealed insignificant differences in the fractions of low-R CSL boundaries as the fraction of low-R CSL boundaries in the specimens R5 and T6 was 52.6 and 51.2 pct, respectively. Also, it can be observed from Figures 3 and 4 that the fraction of R1 boundaries increased at the expense of the fraction of low-R CSL boundaries with increase in the equivalent strain for both deformation modes. However, when the equivalent strain was increased from 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003029_2010-01-0639-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003029_2010-01-0639-Figure5-1.png", + "caption": "Fig. 5. A six-axis load cell prototype can be noticed and the characteristic three spoke structure is highlighted (.b). Concept design of the six-axis load cell suitable to measure the 3 forces (Fx, Fy, Fz) and 3 moments (Tx, Ty, Tz) acting on the cell itself (.a).", + "texts": [ + " The tested suspension is connected to the chassis via two elastic bushings at each wishbone arm (upper and lower) and at the top end attachment of the suspension strut. Each connection point has been equipped with a special high precision 6-axis load cell as shown in Fig. 2.b. In order to measure the resultant force and moment which are applied to the superstructure (i.e. vehicle chassis) the steering link can also be instrumented. The 6-axis load cells have been properly designed, modeled, calibrated and patented [4] (Fig. 5.a) Many commercially available multi axis load cells are able to measure the steady or dynamic forces and moments acting on a structure [5, 6, 6, 8, 9]. Nevertheless their use is not widespread in the research and development activities for their high cost. The 6-axis load cell has been conceived in order to have a simple device able to measure the three components of a force and the three components of a moment acting on the load cell itself but with good performances in terms of natural frequencies, limited mass and cost, high resolution and low cross-talk (for the technical specification, see Tab", + " The calibration method used for the six-axis load cell is described in the following. Known forces and moments were applied and measured by means of a single axis precision load cell. Two different calibration rigs have been used. The first calibration rig has been employed to apply pure axial loads (Fig. 6.a). The load cell was placed horizontally so as the loads acted along the sensor axis y. A second calibration rig has been used to apply forces and moments in the directions Fx, Fz, Tx, Ty, Tz (Fig. 6.b). The orientation of the reference system axes is reported in Fig. 5.b. According to [10], by applying more than 150 load combinations, via a minimization of the square errors between the applied loads and the load cell measures, the identification of the components of the calibration matrix (Mi eq. 2) has been performed. This procedure has been used to estimate the cross-talk of the device. The vector of the 3 forces and 3 moments acting at the centre of the sensor can be calculated simply by multiplying the calibration matrix Mi by the vector of the tensions \u0394V measured at the six half bridges [12]: SAE Int", + " Two different types of indoor test for suspension systems linked to the connecting structure above RuotaVia drum can been performed \u2022 Cleat test: where the wheel passes over an obstacle at different drum tangential speeds (simulated vehicle speed) up to 100 km/h to characterize the vibration performance \u2022 Flat surface test: speed ramp test from 0 to 250 km/h: the purpose of this test is to evaluate the tire/wheel unbalance effect. 1The stiffnesses parameters are referred to the load cell axis (see Fig. 5.b) in correspondence with the connection interface with the tested structure (see Fig. 5.a). SAE Int. J. Mater. Manuf. | Volume 3 | Issue 1 296 For each type of test, it is possible to measure \u2022 forces and moments at each joint connecting the car body to the suspension. 6-axis load cells (Fig. 2.b, 5) \u2022 tire forces and moments applied to the wheel centre. A measuring hub has been designed and used [16]; \u2022 displacements of the wheel centre in longitudinal and vertical directions by means of an optical device; \u2022 accelerations at various locations of the system (wheel carrier, suspension arms \u2026)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000271_971062-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000271_971062-Figure2-1.png", + "caption": "Figure 2 3 Critical Turn-away Distance with Circular Arc Path", + "texts": [ + ";umptions, and the path measures of performance used to discnrninate between the candidate paths. Itesults and discussion (of results are presented in Section 5. Section 6 presents conclusions. Appendix A summarizes the concept of critical speed and explains its relation to lane(: hange kinematics. 2. BACKGROUND l2.1 TERMINOLOGiY - Various vehicle maneuvers may 1x classified as lane-change maneuvers. Fricke [IS] called onehalf of a lane-change maneuver (at obstacle clearance) a criitical ,!urn-away maneuver. Figure 2.1 illustrates what Fancher et al. j 191 defined as a single /me-change maneuver, which is the ,action of applying ste:ering in one direction to displace a vehicle laterally followedl by counter steering to recover the original direction of .trzlvel, while maintaining directilonal control and minimizing displacement to approximately one lane width. The double /me-change maneuver is similar, but includes additional steering input to recover the original Similar performance measures have also been used direction of travel in the original lane of travel", + " Fancher et al. [19] report lane-change maneuver is sometimes referred to as a sinusoidal use of the following measure of performance ternled lanesteer maneuver. The foregoing maneuvers may be termed channe deviation: - collectively as lane-change maneuvers. However, this paper is concerned primarily with the features of the ideal single lanechange trajectory. A single lane-change trajectory describes the motion of a vehicle (center of gravit:y) along a path from one lane to another as shown in Figure 2.2 In this paper the terms curve, path, pathline, and trajectory are used interchangeably to describe vehicle position throughout a lane-change maneuver. In Figure 2.2, f(x) is a single-valued function describing the trajectory (path) of the vehcle center of gravity, x is the longitudinal distance, y is the lateral displacement, xe is the total longitudinal distance:, ye is the total lateral displacement, and @x) is the tangent dirtxtion angle. 2.2 EVALUATING 'VEHICLE PERFORMANCE USING IDEAL TRAJECTORIES - According to Fancher et al. [19], the three basic reasons for conducting lanechange maneuver tests are: (1) the maneuver's geometry emphasizes important facets of vehicle control or performance- where the chosen period 3", + " Pathline deviation rrlay be measured by various error performance indexes that have been proposed in the literature. The integrals of EQ (2.1) give examples of generic penalty or cost functions [22], which have their origin in the calculus of variations. 2.3 KNOWN DESIRED LANE-CHANGE TRAJECTORIES - Various functions have been used to describe ideal lane-change paths. Some notable trajectory description schemes aue listed in Table 2.2. Circular Arc2 Fett [23], Fricke [IS], and Daily [14] provide derivations and geometry for the circular arc representation of an icleal lane-change path. Figure 2.3 depicts the critical turn-awa.y distance for obstacle avoidance. Fricke [I51 prese:nts a formula for one-half of the single lane-change maneuver for the purpose of determining the critical turn-away distance, d,. He does so using the geometry of Figure 2.4, from which EQs (2.4) and (2.5) are derived for the purpose of determining the center of gravity path radius of curvature, p [24]. where e = e(t) is error or deviation as a function of time, t, The radius can also be described as shown in EQ (2.7), in with e(t) = y,(t) - y(t), e is the rate of change of the error with terms of the critical speed formula - EQ (A.2). respect to time, y,(t) is the: objective or desired path, y(t)is the , - 7 actual or simulated path, and the upper time limit, T, is chosen sufficiently longer than the maneuver period", + "1 1) echoes Limpert's [16] expression for lane-change distance under limit conditions because i t is a critical turnaway equation where only 67 percent of the maximum lateral acceleration is used. where p, is the maximum value for the tire-road coefficient of friction. Daily [I41 takes this approach one step further by connecting the critical turn-away path segment to its mirror image, doubling the distance:, d,, to x, = 2d, , thereby completing the geometry of a single lane-change maneuver as shown in Figure 2.5. In contrast to Fricke's definition of d, as the lateral displacement to avoid an obstacle, Daily considers the lateral distance, ye = 2dY , as the total lateral distance for completion of a single lane-change. A modification of EQ (2.10) to account for the total lateral displacement yields EQ (2.12). EQ (2.1 2) is a common form of the total lane-change distance equation. where xe = the total longitudinal distance ye = the total lateral displacement Fett [23] described the geometry for the circular arc path of a single lane-change maneuver. He noted that the radius of curvature is a function of speed and friction, but he also contended that it is a function of the prescribed lateral displacement and the lane-change duration, 1,. Assuming constant velocity, the longitudinal position at time t is x=Vo t and the lateral position is given by where, in accordance with the geometry of Figure 2.4, the constant radius of curvature is described by For the circular path, I'ett relied on experimental data and assumes the duration of a \"forced lane- change to be about .3.0 st:conds. A forced lane-change is effectively one in which an oncoming obstacle applrolaches rapidly or where headway dtzcreases quickly. There:fc~re, at constant velocity, he considiers the total longitudinal distance for the lane-change maneuver to be x, = Vo te. There are two other wiiys to set up the circular trajectory", + "3 is provided by Nelson [25]: A comparison of curvatures and trajectories for both the 5th and 7th degree SCPs is shown in Figures 2.6 and 2.7. ram^ Sinusoid -. Several authors [16][18][23] have recommended the use of a ramp sinusoidal function to describe the ideal lane-change path. Fett [23] held that a sinusoidal trajectory was preferable to connected circular arcs because it provides a continuous function of continuous curvature and i t accommodates changing velocity. The geometry for ramp sinusoidal analysis is shown in Figure 2.8. Zellner et al. [18] proposed the following function for a sine-shaped lane-change path. where x = x(t) = Vot is the linear displacement in the direction of travel, y ( x ) is the lateral displacement (usually equivalent to ihe lane width), and lateral acceleration, a,, takes the form: Substituting x for Vot and EQ (2.20) for a, reduces EQ (2.19) to a more familiar form [ 1151. For the path function, y(x), the minimum radius is given by: Limpert identified four levels of lateral acceleration for lane-change distance and time estimation formulas: (1) nomzal describes relatively low lateral acceleration, (2) severe describes moderately high lateral acceleration, (3) limit indicates high acceleration, and (4) rnaxirrzurn is reserved for very high lateral acceleration", + " Limpert [16] cites experimental results that yield a distribution of values for the constant coefficient, C:, , with modes at 2.6 and 2.67. Based solely on algebraic manipulation, C, is approximately equal to 2.5 1. S ~ i r a l s - Kar~ayama and Miyake [26] proposed a novel lane-change trajectory while researching smooth local path planning for auto!nomous vehicles. They maintain that Eulerian, or Cornu, spirals are preferable to straight lines and circular arcs because of continuous and smooth pa,ths and curvature. As shown in Figure 2.9, we construct a lanechange path by specicying two endpostures where the curvature is null. The endpostures are a combination of posi~.ion and orientation information; for example: ( p , p, ) + po=:(xp yo, 8,) and pe=(xe, ye, 8, ), where position coordinates (x,, y,)*x,, ye) and 8, is the tangent direction of the path at each posture point. The piecewise continuous curvature for the Cornu spiral path varies linearly. Taking this feature into account, lane-change: path curvature has the fc~llowing piecewise structure: Where 1 is the totall positive length of the curve, K, are curvature constants", + " For example, the deflection for a left turn is ar = ~ 1 2 , and the deflection for a lane-change is a = 0. For a description of lane-change paths of constant velocity vehicles, the parallel case (80 = 8,, so a = 0) of the Cornu spiral is used. The path inflection point is given by the intermediate posture that Kanayama and Hartman [27] refer to as the least cost split point, q. The split point used in conjunction with each endposture yields two clothoid pairs that describe the ideal lane-change path. The minimum radius of curvature occurs at the turning points B and D as depicted in Figure 2.9 and 2.10. Acceleration Profiltg - Prescribing a lateral acceleration profile enables analysis of the effect of lateral acceleration on a lane-change trajectory. Figure 2.1 1 compares three examples of acceleratior~ profiles where a , , = p,g. The first is sinusoidal with a maximum amplitude equal to a,,, [12], the second is trape;:oidal with maximum amplitude equal to 0.85 a,, ,, [ I l l , and the third, the crudest, is represented by a square wave with a maximum amplitude of 0.67 a,, , 1161. Chovan et al. [12] prescribed a sinusoidal lateral acceleration (a,) profile as a function of time: a ,. = a,,,, sin@?) := ( 2 e ) ( ) - 12.32) where a , , E maximum lateral acceleration = 0", + "35) Limpert [ 161 used a quare wave acceleration profile with a.n amplitude equal to 67 percent of the nominal maximum as an average lateral acce1,eration. This acceleration profile is represented by one period of a square wave as shown in Figure Z ! . 14. Bezier-S~l ine -, In the Bezier spline approach, the spline algorithm interpolates only the prescribed endpoints; the other points are used to control the shape of the spline curve [28]. According to Ape:taur and Opicka [8], a simple fomn of the Bezier Spline equatiorr is: In Figure 2.15, the spline trajectory is constructed using lanechange dimensions h=3.h6 m and x, = 62.14 nl. These dimensions are inputs to the following knot algorithm: where n = 7 is the numbel- of knots (points) and n- 1 is number of equal segments along the x-axis. Again, the vehicle speed is assumed to be constant; therefore, the knot locations are symmetrical about the path center (x,/2, yJ2). In Figure 2.15, the Bezier spline curve is tjisplayed along with the cubic spline curve using the same knots. For the given number of knots, the Bezier spline path is more desirable. Other Lane-Change Traiectories - Several other trajectories, not presented in detail here, are discussed in the literature. Nelson [25] proposed replacing the circular arc with polar polynomials with continuous curvature. Guldner et al. [29] used curvature profiles to design reference paths for assessing automatic steering controllers", + "6) Lising the same procedure, a similar expression for circular arcs is, determined as For the more complicated expressions for minimum radius; of curvature, iterative techniques are required to solve for x,. Rearranging EQ (A l), the maximum curvature of a lanechange trajectory is limited by the following: Therefore, if x, is specifie:d along with ye, and p,, Vo may be determined and vice versa. Then, by rearrangement, the total longitudinal distance is determined as APPENIDIX B APPENDIX C FIGURES meter:; trajectory/path - 7.4 - Road's Edge - 6.2 ---t - - 1 - -t y displacement I- / i - - 0 - Road's Edge 30.7 46.2 61.5 x distance (m) -- C Figure 2.1 Typical Single Lane-Change Evasive Maneuver I --__if x distance (nzeters) Figure 2.2 Single Lane-(Change Trajectory Lane-Change Trajectory Using a Seven-Degree Cubic and Bezier Splines 1 I Vehicle sick- , I slip angl'e I 1 I I I I I I I 1 I I ! I , I I I I I : TO instantaneous I r path center i Figure A.l Oversteer Yaw Motion" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001424_j.1600-0846.2012.00652.x-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001424_j.1600-0846.2012.00652.x-Figure2-1.png", + "caption": "Fig. 2. (a) pQCT image of the forearm of a single subject. The image was used to measure areas of different tissues. Pixel size was 0.2 mm 9 0.2 mm. (b) Image after tissue segmentations with arm support removed and (c) the simplified geometry used for FE simulations.", + "texts": [ + " Subsequently, after a 5 min break, venous occlusion (VO) was created by using a pressure cuff at 8 kPa (60 mmHg) for 12 min to induce swelling of the forearm soft tissues. During each protocol, 20 repeated measurements were conducted, and the mean responses (deformation vs. suction pressure) of them were used as a target for model responses. pQCT measurements The pQCT imaging of the forearm was performed at the sites of mechanical testing (Stratec XCT 2000, Stratec Medizintechnik GmbH, Pforzheim, Germany). One image slice (thickness = 2.3 mm, pixel size = 0.2 mm 9 0.2 mm) was taken from each subject [Fig. 2(a)]. Different tissue layers were manually segmented from the images, and the tissue areas were calculated using Matlab. The location of tissue midpoint was also based on pQCT measurements. Using the pQCT images the model geometry was constructed by first simplifying each tissue cross-section to be of circular shape [Fig. 2(b)]. Then, a circular shaped cross-section area, defined by the circle radius, of each tissue in the model was matched with the measured cross-section area (Table 1). Finally, the 2D geometry was extruded to 3D, achieving a cylindrical shape for the model geometry [Fig. 2(c)]. Three different models were created; one was based on the mean experimental results of all subjects (N = 11), while two were based on measurement results of two individuals with clearly different (maximum, minimum) tissue responses. The FE model (Fig. 3) was created and simulations were conducted using ABAQUS 6.9 (SIMULIA, Dassault Syste\u0300mes, Providence, RI, USA). In the model geometry, the forearm was 30 cm in length, long enough to exclude any edge effects. The suction head consisted of 3063\u20136774 rigid 4-node surface elements (R3D4) and the tissues of 24860\u201332976 hexahedral 8-node elements with reduced inte- gration points (C3D8R)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003461_j.camwa.2016.08.018-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003461_j.camwa.2016.08.018-Figure2-1.png", + "caption": "Fig. 2. (a) The flexible beam and body coordinate system attached to rigid position. (b) Beam element. (c) Part of the beam on right side of the beam element.", + "texts": [ + " The passive prismatic joint is attached to end of the rigid first link and therefore has the same motion as the rigid first link. It is assumed that there is no friction between the prismatic joint and the flexible link. In addition, a horizontal driving force is applied on the left end of the flexible link to provide its axial motion. It is also assumed that the flexible link has zero initial conditions. Additionally, to focus on the vibration behavior of the flexible link, the mass of the first link is assumed to be zero. An element of the beam is considered and shown in Fig. 2(a). Effective forces applied to the element are shown in Fig. 2(b). Effective forces applied to right hand side of the beam is shown in Fig. 2(c). Euler\u2013Bernoulli theory is used for the flexible link. Therefore, shear deformation effects, rotational inertia and axial deformation are neglected. Additionally, slope angle due to link deflection is assumed to be small; therefore sin (\u03b8) \u2248 tan (\u03b8) \u2248 \u03b8 \u2248 \u2202w/\u2202x and cos (\u03b8) \u2248 1. The assumption of linearizing the model is reasonable as we intend to investigate vibrations with low amplitude. As shown in Figs. 1 and 2, beam lateral vibration is in z direction. Axial motion of the beam is in the x direction", + " This results in additional and complicated derivative terms in the differential equation of motion. Therefore, its solving process and vibration analysis become difficult. Specifically, the beam lateral displacement, w, is a function of the element distance, X , and time, t , therefore, the total time derivative of the beam lateral displacement is w = w (X, t) \u21d2 d2w dt2 = \u22022w \u2202t2 + 2 \u2202X \u2202t \u22022w \u2202X\u2202t + \u2202X \u2202t 2 \u22022w \u2202X2 + \u22022X \u2202t2 \u2202w \u2202X (1) in which we can write \u2202X \u2202t = v(t). (2) However, in this research, coordinate system xz is attached to the beamas shown in Fig. 2(a) and the element is at distance x. Consequently, the coordinate system has the same rigid body motion as the beam (body coordinate system). With this treatment, the coordinate system moves in the x direction with velocity v (t). Therefore, derivative of axial position of the element with respect to the xz coordinate system is zero (\u2202x/\u2202t = 0) and the total time derivative reduces to a much simpler form as w = w (x, t) \u21d2 d2w dt2 = \u22022w \u2202t2 . (3) In Fig. 2 parameters T , V ,M and \u03b8 are axial force, shear force, bendingmoment and slope angle of the beam, respectively. Additionally, parameter L and \u03c1 represent the beam length and mass per length of the beam, respectively. In general \u03c1 can be a function of x. Consider Fig. 2(b). Using Newton\u2019s law for the beam element in w direction we can write \u03c11x d2w dt2 = \u2212 \u2202 (V cos (\u03b8)) \u2202x 1x + \u2202 (T sin (\u03b8)) \u2202x 1x \u2212 \u03c1g1x. (4) In addition, in the x direction we have \u03c11xv\u0307 (t) = + \u2202 (T cos (\u03b8)) \u2202x 1x + \u2202 (V sin (\u03b8)) \u2202x 1x. (5) Dividing Eqs. (4) and (5) by 1x and assuming small value for \u03b8 so that sin (\u03b8) \u2248 \u03b8 and cos (\u03b8) \u2248 1, we can simplify Eqs. (4) and (5) as follows \u03c1 d2w dt2 = \u2212 \u2202V \u2202x + \u2202 (T\u03b8) \u2202x \u2212 \u03c1g (6) \u03c1v\u0307 (t) = + \u2202T \u2202x + \u2202 (V\u03b8) \u2202x . (7) We assume shear deformation and rotational inertia effects are negligible" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001177_journal.202.2014.3.202-3950-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001177_journal.202.2014.3.202-3950-Figure3-1.png", + "caption": "Figure 3. State trajectories of the chopper.", + "texts": [ + " Thus, by using (1) the following state space model may be obtained for the first subinterval: \u23a1 \u23a3 x\u03071 x\u03072 \u23a4 \u23a6 = \u23a1 \u23a2\u23a2\u23a3 \u2212 1 L ( R1 + RRc R+Rc ) \u2212 R (R+Rc)L \u2212 Rc (R+Rc)C \u2212 1 (R+Rc)C \u23a4 \u23a5\u23a5\u23a6 \u23a1 \u23a3 x1 x2 \u23a4 \u23a6 + \u23a1 \u23a2\u23a2\u23a3 1 L RRc (R+Rc)L 0 \u2212 R (R+Rc)C \u23a4 \u23a5\u23a5\u23a6 \u23a1 \u23a3 v1 ig \u23a4 \u23a6 (3) vo = [ RRc (R+Rc) R (R+Rc) ]\u23a1 \u23a3 x1 x2 \u23a4 \u23a6 + [ 0 \u2212 RRc (R+Rc) ]\u23a1 \u23a3 v1 ig \u23a4 \u23a6 where X= [x1 x2] is the state space vector of the system and ig is a DC current injected at the output of the regulator as a DC control signal required to determine the regulator impedance (Fig. 3). The second subinterval, during which Q is off, may be achieved by setting all values of V1 in (3) to zero, therefore: \u23a1 \u23a3 x\u03071 x\u03072 \u23a4 \u23a6 = \u23a1 \u23a2\u23a2\u23a2\u23a3 \u2212 1 L ( R1 + RRc R+Rc ) \u2212 R (R+Rc)L \u2212 Rc (R+Rc)C \u2212 1 (R+Rc)C \u23a4 \u23a5\u23a5\u23a5\u23a6 \u23a1 \u23a3 x1 x2 \u23a4 \u23a6 + \u23a1 \u23a2\u23a2\u23a3 1 L RRc (R+Rc)L 0 \u2212 R (R+Rc)C \u23a4 \u23a5\u23a5\u23a6 \u23a1 \u23a3 0 ig \u23a4 \u23a6 (4) vo = [ RRc (R+Rc) R (R+Rc) ]\u23a1 \u23a3 x1 x2 \u23a4 \u23a6 + [ 0 \u2212 RRc (R+Rc) ]\u23a1 \u23a30 ig \u23a4 \u23a6 Cuk model claims that if a unidirectional low-frequency non-harmonic signal exists in the dynamic system, the steady state of such systemmay be described using a single, weighted sum of both modes: dX dt = A2X+B2U + (A1 \u2212A2)XD + (B1 \u2212B2)UD (5) Y = C2X+ E2U + (C1 \u2212 C2)XD + (E1 \u2212 E2)UD The matrices A1, B1, C1, E1 and A2, B2, C2, E2 represent the coefficients in (3) and (4), respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003727_j.2042-3306.1989.tb02089.x-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003727_j.2042-3306.1989.tb02089.x-Figure2-1.png", + "caption": "Fig 2 . The ,?ate designed tofit on (I standurcl truck is illu.strateclfrorn the fop. the side andfrom behind. Reference distances are I .25 m ahoire thcj ground nmhen plrmd on a truck. The distance between the markers i s one nietrefrnni thc side andfioni behind", + "texts": [ + " Ontario) were glued to the skin as illustrated in Fig I , The markers for lateral filming were placed on the cranial aspect of the dorsal tuber coxae, the lateral tailhead, the lateral tuberosity of the proximal tibia, at the level of the distal calcaneus of the tarsus and at the midpoint of the lateral condyle of the third metatarsal bone. For filming from behind (caudal) the markers were placed on the tailhead, on the calcaneal tuber and over the ergot of the fetlock. In order to control the horse's speed strictly, facilitate left and right lateral filming, caudal filming and have reference points in all planes, a gate which fits on a standard half or three-quarter ton truck was designed and manufactured (Fig 2 ) . The standard heights and widths of the gate used for reference values were calibrated daily using the adjustable system incorporated in the design. A Locam I1 16 mm camera with a 22.5-90 mm zoom lens (Redlake Corporation, Morgan Hill, California) was used to film the horses from a mini-truck. The film used was Eastrnan Ektrachrome high speed daylight 7251 and the camera was operated at a frame rate of 200 frames/sec for standardising procedures and 100 frames/sec for the eight trials. Filming was SlLk VIEW SO EQUINE VETERINARY JOURNAL performed on the backstretch of a racetrack designed for Standardbred racing (Mohawk Raceway, Campbellville, Ontario)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002686_12.328516-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002686_12.328516-Figure1-1.png", + "caption": "Figure 1. The Space Infrared Telescope Facility (SIRTF)", + "texts": [ + "org/ss/TermsOfUse.aspx tmiling the earth, and will performbackground-limited imaging and spectroscopic measurements ofcelestial objects in the 3-180 im spectral range. SIRTF carries an 85 cm aperturetelescope at 5.5 K delivering 6.5 jim diffraction-limited imaging to three science instruments (SI) with focal plane detectors cooled as low as 1 .5 K. The observatory cryostat carries 360 liters of superfluid helium cryogen, which is expected to last in excess of 2.5 years. The overall observatory is shown in Figure 1. Past cryogenic space telescope designs, such as the Infrared Astronomy Satellite (IRAS) launched in 1981, and the Europeansatellitelnfrared Space Observatory (ISO) launched in 1995, had enclosed both the telescope and the science instruments inside the ciyostat. The vacuum shell resulting from such an arrangement is relatively large and massive. SIRTF uses Warm Launch Architecture[3} where only the science instruments are enclosed in the cryostat; the telescope is mounted externally to the vacuum shell, as shown in Figure 2" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000431_j.engappai.2004.09.003-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000431_j.engappai.2004.09.003-Figure2-1.png", + "caption": "Fig. 2. Stator and rotor laminations of a UM.", + "texts": [ + " Korous\u030cic\u0301-Seljak / Engineering Applications of Artificial Intelligence 18 (2005) 47\u20135548 basic construction resembles the design of a DC series motor. A UM performs like a series motor\u2014the same current, regardless of the power supply, passes through both the armature (rotor) windings and the fieldexcitation (stator) windings via the brushes in one continuous path. Fig. 1 shows the rotor and the stator parts of a UM. The rotor-and-stator unit of a UM is constructed by stacking the rotor/stator iron laminations (see Fig. 2). The shape and the profile of the rotor/stator lamination are described by several two-dimensional (2D) geometrical parameters. There are two types of parameter: the invariable and the variable. Invariable parameters are fixed; they cannot be altered, either for technical reasons or because of the physical constraints of the motor. See Table 1 for details of geometrical parameters. In our case, there are 12 mutually independent variable parameters that we can optimize. However, some important problem constraints have to be taken into account: The parameters should be changed simultaneously (both independent and dependent parameters) to achieve proper electromagnetic conditions in the material" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001824_022050-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001824_022050-Figure3-1.png", + "caption": "Figure 3. Loading conditions.", + "texts": [ + " Lateral load, load 3rd International Conference on Advances in Mechanical Engineering (ICAME 2020) IOP Conf. Series: Materials Science and Engineering 912 (2020) 022050 IOP Publishing doi:10.1088/1757-899X/912/2/022050 transfer and longitudinal load of 2130N, 1031N and 686N respectively was applied at upper and lower wishbone ball end joint. The front uprights also experienced a bump load of 1560N which is transferred to the chassis through damper. A fixture was set up at bearing surface of the upright. Figure 3 shows all the force acting on the upright, simultaneously, so that the upright can be tested for the worst case scenario. . During braking, throttling and steering simultaneously the upright experiences maximum forces, so the upright is designed for a condition in which the car corners while braking and also encounters a bump. So all the before mentioned forces act simultaneously on the upright. The maximum stress and total deformation were obtained as 57.95N and 0.05mm respectively. Figure 4 and figure 5 depicts the stress flow and deformation flow path respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003354_s10015-016-0339-9-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003354_s10015-016-0339-9-Figure4-1.png", + "caption": "Fig. 4 Lift force part", + "texts": [ + " Part A generates thrust force mainly, and part B generates lift force mainly. We set the airfoils (NACA2415) at wing\u2019s links and wing root, and covered them by sill nylon film. In flapping-up, the wing generates not only contra-lift force but also large thrust force. In flapping-down, the wing generated not only thrust force but also lift force. We set the airfoils (NACA2415) at wing\u2019s links and use a thread to fix at the certain position. We covered the airfoils by sill nylon film and the front cover to generate lift force (Fig. 4). 1 3 We fixed the bio-inspired position of the nine feathers at part A shown in Fig. 5. In flapping-down, the nine feathers acted as one large feather. On the other hand, in flappingup, the nine feathers moved up separately and air went through between feathers shown in Fig. 6. We measured lift force and thrust force when the MAV wings were flapped. The detail of experimental method is shown in Fig. 7. We fixed MAV at the end of beam and measure the displacement of beam by laser displacement sensor" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001106_tia.2013.2253079-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001106_tia.2013.2253079-Figure1-1.png", + "caption": "Fig. 1. Double-layer PM EVT machine.", + "texts": [ + " Then, unique power conversion characteristics of the EVT-based system are investigated and compared with the conventional DFIG system and PMG system, including the improved power conversion coefficient, maximum power acquirement, and fault ride through (FRT) capability. A controller with maximum power point tracking (MPPT) for the EVTbased wind power generation system is developed and implemented in Matlab/Simulink. Computer simulation and system experiment results are finally provided to verify the working principle and the unique power conversion characteristics. 0093-9994/$31.00 \u00a9 2013 IEEE Fig. 1 shows the schematic of the EVT with dual mechanical ports and dual electrical ports, which consists of three parts, namely, a three-phase stator, a wound three-phase inner rotor, and a PM outer rotor. Several types of outer rotor for the EVT have been proposed in recent years, namely, the squirrel-cage outer rotor, double-layer PM outer rotor, and single-layer PM outer rotor. Because of high power density, high reliability, high efficiency, and easy control, the double-layer PM outer rotor is adopted in the EVT-based wind power generation system [12]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003464_ijvd.2017.082579-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003464_ijvd.2017.082579-Figure2-1.png", + "caption": "Figure 2 Initial design domain and boundary conditions of an automobile part (see online version for colours)", + "texts": [ + " The use of the proposed hybrid approach improves the convergence rate by computing the best value 1.7248 with the smallest function evaluation 20.000 and standard deviation 0.00041 values. As can be seen from Tables 1 and 2, HCSSNM gives the best results reported in the literature for welded beam design problem. The hybrid approach proposed in Section 3 is applied to optimal structural design of an automobile suspension arm taken from automotive industry. Initial design domain and boundary conditions of the automobile part is given in Figure 2. Minimisation of weight is chosen as objective function. Fatigue life is chosen as constraint function in this problem. Input variables for the meta-models are the five design variables which are x1, x2, x3, x4 and x5 as shown in Figure 3. Initial values, lower and upper limits of the design variables are provided in Table 3. In this study, LHS was used to sample the design space for a total of 50 training. Weight and fatigue life are calculated for each 50 experiment. The optimisation problem is formulated as follow: 1Min ( ) ( )F x f x= (13) 2( ) ( ) 1 6g x f x e= > , 1, ,l u i i ix x x i NDV\u2264 \u2264 = where f1(x) and f2(x) represent the weight and fatigue life values as objective function and constraint, respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000745_2000-01-0498-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000745_2000-01-0498-Figure2-1.png", + "caption": "Figure 2. Structure of the Burex 9 decoupling element", + "texts": [], + "surrounding_texts": [ + "The function of a decoupling element is to connect the \"moving\" engine section of the exhaust gas system with the \u201crigid\u201d area of the vehicle. By doing this, the aim is primarily to compensate for the relative (tilting) movements of the engine which occur especially in the case of transverse fittings. With the decoupling element, the gas-conducting components such as the manifold and the down pipe are relieved from mechanical tension to the extent that the principle of lightweight construction can be implemented here. The generally increasing requirements with regard to driving comfort concern not just a reduction in the level of vibration in the vehicle but also a drop in the level of noise. By employing a decoupling element, the transmission of structure-borne noise (e.g. noises generated by the charger and the engine) to subsequent areas of the exhaust gas system is reduced. This decrease in the medium- to high-frequency radiation of the components results in improved acoustics within the passenger section. Thus costly encapsulations or shieldings in these areas can be dispensed with. In today's applications, decoupling elements are installed at various positions of the exhaust gas system separate from the converter and silencer arrangement. Fig. 1 shows existing arrangements of decoupling elements. Figure 1. Arrangement of decoupling elements in the exhaust gas system" + ] + }, + { + "image_filename": "designv6_24_0001733_icelmach.2018.8507255-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001733_icelmach.2018.8507255-Figure6-1.png", + "caption": "Fig. 6. View of the main machine parts", + "texts": [ + " The main dimensions of the prototype are given in table II. The two stator cores are made of ATOMET EM1 soft magnetic composite. The main physical properties of this material are described in [24]. Each of the stator is equipped with a quite common 2-layer tooth-concentrated winding where the number of slots per pole and per phase is spp = 0.5. This tooth-concentrated winding provides a fundamental winding factor of about 0.866 [25]. The ironless rotor integrating the magnets is located in the center of the system as illustrated by figure 1. Figure 6 shows a snapshot of the machine different parts. Rotor is located on the upper left corner of the picture and the two stators are located in the center (in up) of the picture. The distance between the two stators can be set with a screw system located in the machine. Figure 7 shows a picture of the prototype in the test bench. The black scroll wheels located in each side of the structure allow to tune the values of the air gaps. This prototype machine is mechanically coupled to a DC motor which for load or lead purpose" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002850_12.599519-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002850_12.599519-Figure3-1.png", + "caption": "Figure 3. All side clamped plate", + "texts": [ + " The beam material is steel and the cross-section area is 30 mm \u00d7 5 mm. The strain is measured at position 72 mm, 288 mm and 320 mm. The beam is modeled using five equal finite beam elements. Two displacement measurements are taken in node 3,4 or 5 of the finite element mesh (see Fig. 5.1). The contact point is at the end of the beam and is realized by a steel tip. Additionally, the beam can be excited by a modal hammer which measures the force. The experimental setup of the elastic plate is given in Fig. 3 (dimensions in mm). The scheme of the test rig with the positions of the actuators and the sensors is shown in Fig. 3(b). The plate measures 780 \u00d7 780 mm and has the thickness dp = 0.7 mm. Two piezo actuators (PZT patches) are bonded on the plate as shown. The displacements are measured at two points with non-contacting optical measurement systems and the strains are measured at the illustrated positions in length direction of the drawn rectangles (see Fig. 3(b)). The plate 340 Proc. of SPIE Vol. 5764 Downloaded From: http://proceedings.spiedigitallibrary.org/ on 06/22/2016 Terms of Use: http://spiedigitallibrary.org/ss/TermsOfUse.aspx is excited by a hammer, which directly measures the contact force. The plate is modeled using 64 equal plate elements (with different boundary conditions). The modeling of the introduced plate is very difficult because of the changing properties of the environment (temperature). The first eigenfrequency of the plate e", + " (1) is considered without unknown inputs (n(t) = 0) and transformed by x(t) = \u03c6 x\u0303, where \u03c6 is the modal matrix of the system. The system is described in modal coordinates by \u02d9\u0303x(t) = A\u0303x\u0303(t) + B\u0303u(t) , (29) y(t) = C\u0303x\u0303(t) . (30) From matrix B\u0303 = \u03c6\u22121B it can be seen if and how strong the modes of the system can be influenced. The same approach is used in reference17 to place the actuators. From practical point of view only few of the first modes has to be considered. Considering the matrix B\u0303 for the system given in Fig. 3 with 169 equal finite plate elements, the positions of the PZT patches shown in Fig. 3 are chosen and present a good combination to excite the first 4 modes of the system. The location of the sensors is also important for the observability of the structure. The strain gages are positioned where the stress is maximal for the first and second mode, in order to get a high resolution for the measurements. The states of the finite element model of the plate contain displacements w(x, y, t) and bendings (\u2202w(x,y,t) \u2202x , \u2202w(x,y,t) \u2202y , \u22022w(x,y,t) \u2202x\u2202y ) of the finite element nodes and their time derivatives", + "org/ on 06/22/2016 Terms of Use: http://spiedigitallibrary.org/ss/TermsOfUse.aspx where wij(x, y) is the plate displacement modal amplitude and qi,j(t) are the generalized coordinates. More detailed descriptions are given in reference.18 Using \u22022w(x, y, t) \u2202x2 = \u221e\u2211 i=1 \u221e\u2211 j=1 \u22022wij(x, y) \u2202x2 qij(t) , (32) \u22022w(x, y, t) \u2202y2 = \u221e\u2211 i=1 \u221e\u2211 j=1 \u22022wij(x, y) \u2202y2 qij(t) (33) the displacements can be calculated by the strain gage measurements. In this experiment, four strain gage measurements are used (see Fig. 3), so only w11(x, y), w21(x, y), w12(x, y) and w22(x, y) can be considered. For the beam, Eq. (31) has only to be considered for the x direction. The beam gets in contact with the contact surface after a short excursion of the beam (dashed position in Fig. 2(b)). The displacement is measured in node 4 and 5. In Fig. 5(a) and 5(b) the measured and the estimated force are compared. In Fig. 5(a) the design parameter q = 103 is too small, the observer can not follow the fast dynamic of the contact force", + " This only works if one of the measurements is in node 5, to get the entire information about the deformation of the beam. Proc. of SPIE Vol. 5764 343 Downloaded From: http://proceedings.spiedigitallibrary.org/ on 06/22/2016 Terms of Use: http://spiedigitallibrary.org/ss/TermsOfUse.aspx The displacement measurements for the plate are taken in two different positions. In measurement position 1 the first displacement is measured in (x = 390 mm, y = 585 mm) and the second in (x = 585 mm, y = 585 mm, see coordinate system in Fig. 3(b)). The measurements are collocated with the nodes of the finite element mesh. The plate is divided into 64 equal plate elements. In measurement position 2 the first displacement is measured in (x = 390 mm, y = 390 mm) and the second in (x = 585 mm, y = 390 mm). A contact in two different positions, contact 1 (x = 390 mm, y = 390 mm) and contact 2 (x = 585 mm, y = 390 mm) are applied to the plate. In Fig. 9(a) the measured and the estimated contact force are compared. The measurement of displacements is taken in measurement position 1 and the contact is applied in position contact 1", + " The contact can be detected but after the contact the estimated force oscillate with the same amplitude as the contact itself. The model of the plate seems not to be sufficient. In Fig. 10 the estimation is shown for the displacement measurement in measurement position 2 and the contact is applied in position contact 2. The results are similar as presented before. In Fig. 11 a simulation result of the proposed control approach is shown. Displacements of the plate are measured at two different positions, y1(x=360 mm,y=540 mm), y2(x=600 mm,y=540 mm) and the two PZT patches are bonded in positions as shown in Fig. 3(b). The reference signal is chosen to be w(t) = 150 V (V = I2\u00d72). The state feedback matrix is calculated via the linear quadratic optimal control design approach. In this case, the parameters of the weighting matrix are chosen in such a way that the actuating signal stays between 100 V and 200 V, so they are not optimized from a theoretical point of view. The disturbance rejection part is not applied in this simulation example, since in the real experiment the PZT patches have not enough power for static deformations and only vibration control is possible" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000982_cp.2018.0938-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000982_cp.2018.0938-Figure2-1.png", + "caption": "Fig. 2: Multilayered structural design of CBCPW feed ACSP Antenna", + "texts": [ + " 5) The LTCC package walls provide a support structure for injecting a molding compound into the package cavity and thereafter curing it. This methodology is similar to the dam and fill process which is commercially used for precise encapsulation of chips with high I/O count [3]. In compraison to glob-top encapsulation, the volume covered by the molding compound is fixed; hence, the AiP performance can be precisely estimated beforehand by means of electromagnetic simulations An aperture-coupled stacked-patch (ACSP) antenna, shown in Fig. 2 has been used in this work. The assembly consists of seven gold metal layers, numbered top to bottom and five layers of LTCC dielectric substrate DuPontTM GreenTapeTM 9K7V (\u03b5r = 7.0, tan\u03b4 =0.0012 at 60 GHz) with an average fired thickness of \u2248 106.7 \u03bcm per layer. The antenna feed is provided by an asymmetric stripline (ASL) in metal layer 5 (ML 5) and the ground vias on either side ensure excitation of the dominant transverse electromagnetic mode (TEM) only. A rectangular slot placed in the intermediate ground plane (ML 3) is used to couple the electromagnetic energy from the ASL to the antenna radiating elements namely the two stacked patches and the double via fence (ML 1 & ML 2)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000951_tmag.2017.2668442-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000951_tmag.2017.2668442-Figure3-1.png", + "caption": "Fig. 3. The analysis model. (a) 3D model.(b) three-layer field model.", + "texts": [ + " According to reference [4], the rotational speed of a 1000 MW hydraulic turbo-generator is 107 rpm. To simplify the supply power, direct current and a 50 Hz power frequency supply are implemented in the device. Therefore, the pole-pairs p is designed as 14. When the power frequency is 25Hz, the rotational speed of primary traveling-wave magnetic fields is about 107 rpm. Table I list the basic parameters of the proposed device. A. 3D Finite element model analysis 3D transient finite element method (FEM) is implemented to analyze the proposed levitation device. Fig. 3 shows the 3D model and three-layer sub-domain field model of the proposed levitation device. Based on the theory of referenced [5], the convective Maxwell equation used for the finite element analysis (FEA) can be expressed as follows. Concerning the secondary (region I), 1 ( ( ) ( ) 0 [ ( )] 0 A v A v A A) (1) Concerning the superconducting primary (region II), 1 ( ) ( ) sA A J (2) Concerning the air-gap (region III), 1 ( ( ) 0 A A) (3) Concerning the air balloon (region IV), =0A (4) where , A , , are the magnetic permeability, magnetic vector potential , penalty factor, and electric scalar potential, respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001994_j.rcim.2007.04.003-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001994_j.rcim.2007.04.003-Figure1-1.png", + "caption": "Fig. 1. Normal and tangential stiffness of a deburring tool.", + "texts": [ + " One is to reduce chatter, and the other to reinforce the fault, which is the weak fixation between the robot and deburring tool. In contrast to machine tools, whose structural stiffness is high in all directions, the relatively low stiffness of robot arms allows large-amplitude resonances, which cause chatter. Asada and Goldfine [44] have analyzed the mechanics of robotic grinding, and shown that chatter is reduced when the normal and tangential stiffness differ by a factor of 10, as shown in Fig. 1. This difference in relative magnitudes and direction of stiffness reduces the strong coupling between the normal and tangential directions. Because the overall stiffness depends on both the robot and the tool, a tool should have an axis with the stiffness less than a tenth of the stiffness of the robot. Deburring irregular edges requires higher accuracy than polishing and milling machined edges. The normal direction must be made compliant, so that the cutter will remain in contact with the edge and sustain small normal forces needed for edge breaking" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002105_ias.2003.1257748-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002105_ias.2003.1257748-Figure4-1.png", + "caption": "Figure 4. Finite element simulation of the flux switching motor design used in the experimental tests.", + "texts": [ + " Since the field and armature windings are performing different functions within the machine they do not need equal slots [11]. Furthermore, the flux behind a field winding slot is substantially constant and unipolar in nature. The back iron behind a field slot can therefore be thinner without causing excessive iron losses. A field slot is therefore designed to be narrower than the armature slots but slightly deeper. The armature slot area can also be increased slightly to provide additional room for the bifilar armature winding. The motor design used for the experimental prototype tested in this paper is shown in Fig. 4. The outside diameter of the lamination is 90 mm. At the time of the lamination design a full dynamic simulation did not exist and so the following design procedure was adopted : (i) Static FEA was used to alter the tooth shapes and rotor radius to maximise the coupling of flux between the field and armature windings in the aligned positions. The variation in reluctance seen by a fully pitched field winding was minimised though not completely eliminated. A small variation in this reluctance was retained to facilitate braking of the motor through energisation of the field winding alone" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002584_1999-01-0782-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002584_1999-01-0782-Figure5-1.png", + "caption": "Figure 5. Hub movement modeled with angular deflection in response to the applied load.", + "texts": [], + "surrounding_texts": [ + "Epilogics became interested in the Kuhl Wheel as a way of potentially reducing the weight of the steel wheel for original equipment suppliers to the automotive industry. Using the shape flexibility of the Kuhl Wheel spokes, the presumption was a weight savings of 10%-20% could potentially be achieved over existing stamped steel wheel designs without a strength penalty if a suitable means of attaching the spokes to the hub could be determined. The reasoning was that the Kuhl Wheel had two advantages to its design that made it well suited to light weight. The primary advantage was that the spokes of the Kuhl Wheel were deep in the axial direction which provides good stiffness to resist deflection of the hub during cornering as well as low stresses. Secondarily, the spokes of the Kuhl Wheel could be designed to be loaded in tension when a torsional load is applied to the hub. This would make steel a good choice of materials for fabrication due to its excellent tensile strength-to- mass ratio. These two factors made the Kuhl Wheel seem like an excellent starting point to reduce the weight of the conventional stamped steel wheel." + ] + }, + { + "image_filename": "designv6_24_0002050_biorob.2012.6290760-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002050_biorob.2012.6290760-Figure1-1.png", + "caption": "Fig. 1. The elliptical microrobot in the 3D Cartesian frame and in the 2D axisymmetric cylindrical frame. The \u03bd coordinate represents the position on the ellipse and values \u2013 \u03b3/4 and + \u03b3/4 respectively at the top and bottom extremities.", + "texts": [ + " Such an approach enables the design of the proposed mechanism of propulsion for swimming microrobots, which is much simpler than that of the living microorganisms. The model is described in detail in the following. The swimming microrobot was considered as a prolate spheroid moving along its major axis in unbounded water. The prolate spheroid is a good approximation of the shape of the body of many ciliates, such as those of the genus Paramecium [19, 22]. The surface of the spheroid can be expressed in Cartesian coordinates as 1 2 2 2 22 a z b yx (1) where a and b are, respectively, the semi-major and the semi-minor axis (Fig. 1). The metachronal waves were modeled as wave deformations of the surface of the spheroid that travel parallel to its long axis. The model was thus developed exploiting the axial symmetry of both the geometry and the propulsive actions. The problem was represented in a cylindrical frame by means of a 2D axisymmetric model (Fig. 1). In the 2D axisymmetric cylindrical frame the prolate spheroid is represented by a hemi-ellipse and can be expressed as 1 2 2 2 2 a z b r (2) In order to model the deformation of the surface we also adopted a coordinate defined as c pp D pp FF F FF 84 (3) where pF+ and pF\u2013 are the distances of a point on the ellipse from the two foci, DF = 2c is the distance between the foci and \u03b3 is the length of the ellipse. The \u03bd coordinate thus represents the position on the ellipse and varies between \u2013 \u03b3/4 and + \u03b3/4 (Fig. 1). Assuming a simple sinusoidal waveform, the propagating wave of deformation can be expressed as b r tf A 2cos 2 (4) where A is the peak-to-peak amplitude, \u03bb is the wavelength and f is the frequency. The deformation was assumed to be perpendicular to the surface of the microrobot. The r/b factor was introduced for obtaining deformation amplitude that is null on the extremities of the spheroid and maximum on the equatorial circumference. This accounts for the lost of coordination between cilia in the proximity of the extremities of the body of the actual microorganisms", + " For this reason, the resulting hydrodynamic force can be obtained by integrating the stress tensor of the fluid on the surface of the microrobot S dragprop dSpFFF nuuI ))(( T (7) where n is the normal vector on the surface S of the microrobot [13]. This resulting force becomes zero once the steady velocity has been reached. Because of the axial symmetry of the model, only the zcomponent of the force is not null. Therefore the motion of the microrobot can be described by the following FFzm zm (8) where mm is the microrobot mass. The microrobot thus moves at a velocity zv m relative to the fluid. The model was defined using the reference system of Fig. 1, which is fixed on the swimming microrobot. Consequently, with respect to this moving frame, the microrobot is steady while the fluid goes towards it at a velocity vf = \u2013vm (Fig. 2). Hence, the following volumetric force was applied to the fluid: Moreover, the boundary to which the microrobot moves was treated as an inlet with prescribed fluid velocity vf. The model was thus implemented in a Finite-Elements simulation software (COMSOL Multiphysics 4.2). The implementation of the model through the Finite Elements Method will allow us to easily modify it for designing microrobots with different and more complex shapes and features" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000974_2002-01-1347-Figure19-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000974_2002-01-1347-Figure19-1.png", + "caption": "Fig. 19. Milwaukee Cylinder, MilCAD, the C dimension is 7/8\u201d long. The customer needs a 2.25\u201d length.", + "texts": [ + " This feature is controlled by rules established by the product manufacturer. If the manufacturer will allow the extension of rod stick outs or additional thread lengths, or variations on mounting hole diameters then the user can click onto these items and input their required dimension within the manufacturers established envelop. The customer now has an instant drawing for printout and CAD file for insertion into their MCAD system of the \u201cstandard special\u201d allowed by the manufacturer. A record of the changes is automatically created. Fig. 19 thru 21. Generating a CAD file from this system is instant. There is no downloading of the drawing files or there is no waiting on compressed drawings to be sent via e-mail. Once the drawing is generated on the screen, the drawing can be saved directly to the users system, in a variety of 2D, 3D and Solid Model file formats. Security on this system is controlled by Verisign. This certifies the source of the data generated to ensure CAD files saved to the users computer from this system is safe and clean of any virus\u2019s" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002628_0954411021536414-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002628_0954411021536414-Figure3-1.png", + "caption": "Fig. 3 Procedure for monitoring walking frame interface loads using the video vector generator. The vector generator produces a white line on the video screen that is an accurate representation of the vector of ground reaction force, as the subject walks over the force plat-", + "texts": [ + " This during walking showed that changes in vertical aligninvolved the use of the ORLAU video vector generator ment of patient truncal posture could be identi ed visu[8, 9 ], a device that takes the signal outputs from a ally and that the rear wheel on either side of the walking Kistler force platform and converts them into a line rep- frame could become clear of the ground as the patient resenting the vector of ground reaction force, accurately reciprocated their gait. Observation made it clear that superimposed on the video image of the subject, in the these phenomena were related to the exibility of the selected camera plane. Patients were asked to ambulate structure in the sagittal and coronal planes. The test along a path of the walkway that should produce a single procedures shown in Fig. 5 were used to monitor applied foot strike on the force platform (Fig. 3). Sagittal and loads and consequent deformation or de ection in the coronal planes of ve patients using the prototype walk- relevant planes. A solid strut (common to both devices) ing frame were monitored, enabling the vectors of was used to replace the gas spring located between the ground reaction force to be digitized for those walks hinged boom and the main frame. This was done in when suitable strikes occurred. This permitted sensible order to eliminate the legitimate movement essential to approximations of the maximum loads that would be the patient support mechanism" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002245_aim.2005.1511046-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002245_aim.2005.1511046-Figure4-1.png", + "caption": "Fig. 4 Major mechanical structures for a feed drive system.", + "texts": [ + " In recent years, along progress of material science and technology, the granite material is gradually used for the machine structure. It appears that the granite makes machine structural element, not only lighter, but oscillation amplitude smaller than iron. This improvement on material dynamic performance makes the granite good for the high speed machine structural material. Based on the design process described before, the mechanical structure of an experimental feed drive system is designed as shown in Fig. 4. The masses of the base, the saddle, and the table are 570 Kg, 140 Kg and 60 Kg, respectively. And the result of the first resonance frequency for mechanical structure is shown in Fig. 5. The objective at the stage of the design is to get a simple but accurate enough model to predict the machine performance by computer simulation. As shown in Fig. 2, the mechanism components of the feed drive system include ball screw, nut, support bearing, linear bearing carriage, coupling and a motor. In order to analyze dynamic performance of the feed drive system, a multi-body model of the feed drive system is established as illustrated in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002678_j.cad.2011.03.008-Figure13-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002678_j.cad.2011.03.008-Figure13-1.png", + "caption": "Fig. 13. Objects with circular arc edges.", + "texts": [ + " Edge length preservation is a regularity called \u2018\u2018isometry\u2019\u2019 in Lipson\u2019s optimization-based method for 3D object recovery [5]. Even though we can obtain sensible results for inaccurate inputs, further investigation is required to ensure the best result, such as in the ill-conditioned situation above and when there is gross inaccuracy in the corner chosen. Another issue is the extension of the algorithm to deal with curved edges in 2D. Our mathematically deterministic algorithm lends itself well to recovering simple surfaces such as cylinders and spheres. Our work in this area has begun, and Fig. 13 shows some preliminary results in which the curved 3D edges are merely circular arcs, which are projected into ellipses. More work is needed to deal with other curve types. It is desirable to eliminate the need to identify the cubic corner by the user. Perkins [8] has identified the geometric conditions of the edges at a vertex for a cubic corner to be likely. But ambiguities exist, as the example in Fig. 14 shows [9]. Automatic detection of cubic corners remains one of our goals. It is not reasonable, however, to expect automatic generation of the data required for non-cubic corners; the three 3D angles or the lengths of the three edges need to be specified by the user" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001842_icemi.2011.6037951-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001842_icemi.2011.6037951-Figure5-1.png", + "caption": "Fig. 5 Force acting on coil in magnetic field", + "texts": [ + " Lastly, achieve the magnetic field generated by Halbach PM array by summing all of PM together based on linear character. 1 1 1 2 2 2 1 1 1 ( , , ) ( , , ) ... ( , , ) ( , , ) halbach n n n B x y B x x y y B x x y y B x x y y (14) Force acting on coil current in magnetic field is presented as Eq. (15). If the area of coil\u2019s section is S and surface current density is ( , )J x y respectively, we can calculate Lorentz force in following steps. Firstly, partition the section into n cells whose area is ids whose center coordinate is ( , )i ix y , presented in Fig. 5. Then, the force acting on coil can be considered as the sum of force acting on all cells shown as Eq. (16). Conventionally, the stator coil current is twisted around the stator iron yoke and it will result in fluctuation of actuating force and y-dircetion force much bigger than x-dircetion force. If the stator is air-cored long coil without iron yoke, the y-dircetion force will equal to x-dircetion force approximately and the fluctuation of force will decrease[8,9,10,11,12]. We construct linear morotr as presented in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000932_icuwb.2016.7790496-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000932_icuwb.2016.7790496-Figure3-1.png", + "caption": "Figure 3. Branch-line directional coupler with a short stub on the coupling port.", + "texts": [ + " 2, of which the input port (the transmitter port of RFID reader) is port 1, the direct port (antenna port) is port 2, the coupled port (with 50 matched impedance) is port 3 and the isolated port (the receiver port of RFID reader) is port 4. As a 3dB branch-line directional coupler, the characteristic impedances of main lines and branch lines should be Z0m = 35.4 , Z0b = 50 [6]. The coupler is designed on the Rogers RT 5880 substrate with relative permittivity = 2.2 and thickness h = 0.254mm. Its performance is shown as the solid line in Fig. 4, which has a isolation of 55.9dB at 2.45GHz. As shown in Fig. 3, to improve the isolation of coupler, a /4 transmission line is appended to the coupled port of a conventional branch-line directional coupler with a distance of /8 to the branch line. The input signal can transmit to the isolated port through four routes. Assuming that the phase of input signal is 0 on the transmitter port, the phase changes of the transmitted signal is shown in Table I. 978-1-5090-1317-3/16/$31.00 \u00a92016 IEEE On the isolated port, the phase difference between the strong signal leaked from the transmitter to the receiver and the signal transmitted through the other three routes is ", + " Finally, the detailed dimensions of the modified coupler are: wm = 1.25mm, lm = 22.04mm, wb = 0.76mm, lb = 22.4mm, ws = 0.3mm, ls = 22.887mm, ld = 12.065mm. As the dashed line It proves that the isolation can be highly improved by using a quarter-wave length transmission line on the coupled port of directional coupler. Considering the performance, the processing difficulty and the whole layout of coupler, \u201c \u201d lumped-element equivalent network to a transmission line is chosen to realize the coupler shown in Fig. 3. According to [7], the element values of the lumped equivalent network are given by L = (Z0sin )/(2 f0), C = tan( /2)/(2 f0Z0). (1) The lumped equivalent circuit of the modified coupler can be simplified to the circuit shown in Fig. 5, where C1 = Cm + Cb, C2 = Cm + Cb + Cd, C3 = Cd + Cs, Lg is the parasitic inductance caused by via hole. Assuming that Lg is 0.3nH, and inductances in the proposed coupler have a Q of 25 at 2.45GHz. All the circuit values in Fig. 5 can be evaluated using (1), and can be optimized by simulation with the results of Lm = 2" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001310_6.1995-1532-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001310_6.1995-1532-Figure3-1.png", + "caption": "Fig. 3: General layout of the PS", + "texts": [ + " All these elements including the moving parts of the anchor brackets are delivered as a kit ready for use (fig. 2). u The canister is made of welded aluminium skin panels with a beam reinforced bottom and 4 connection brackets to the first ring frame at the top section. Four struts located in the middle of the shell and connected to the third ring frame prevent any lateral motion under the 18 g maximum ascent and re-entry flight loads. Both anchor brackets are machined with titanium alloy to save mass. They have to transfer shock loads up to 1320 kN at parachute opening (fig. 3). The nose cone release system uses an expandible tube which breaks the connection ring to the cone and prevents any fragment generation toward the parachute pack. The cone is jettisoned at 27 mls by the help of 4 pyro-pistons (ref.3). The pistons are installed on the first ring frame of the booster. The pistons and the tube are activated simultaneously by a pyro chain. The pyro command is generated by the control system and secured by a safe and arni device (BSA). The safe position is kept until the end of the ascent flight to bring the risk of inadvertent separation at a negligible level of occurrence (fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002821_mop.30303-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002821_mop.30303-Figure4-1.png", + "caption": "Figure 4 Radiation patterns of the proposed antenna at 920 MHz. (a) XY-plane. (b) XZ-plane. (c) YZ-plane. [Color figure can be viewed at wileyonlinelibrary.com]", + "texts": [ + " Figure 2 shows the simulated and measured reflection coefficients of the proposed antenna with regard to the operating frequency. The simulated result is in good agreement with the measured result. The impedance bandwidth based on 10 dB return loss covered the authorized UHF RFID frequency band in Korea. The simulated peak gain and antenna total efficiency of the proposed antenna are shown in Figure 2(b), and are 4.5 dBi and 80% at the center frequency (920 MHz), respectively. As shown in Figure 3, the proposed antenna can be frequencytuned by controlling the antenna sizes (L1 and L2). Figure 4 Figure 3 Frequency tuning characteristics with regard to the L1 and L2. (a) Reflection coefficients. (b) Impedance variations at operating frequency (f0, 920 MHz) on Smith chart. [Color figure can be viewed at wileyonlinelibrary.com] 440 MICROWAVE AND OPTICAL TECHNOLOGY LETTERS / Vol. 59, No. 2, February 2017 DOI 10.1002/mop represents the simulated and measured radiation patterns of the proposed antenna at 920 MHz. Figure 4(a) shows the radiation pattern in the azimuth plane (XY-plane), and has the null at the edge of the antenna elements. It is similar to a radiation pattern of the dipole. The elevation planes [XZ and YZ planes in Figs. 4(b) and 4(c), respectively] of the proposed antenna have the boresight direction (high directivity), and the maximum front-toback ratio is 17 dB. In this article, we proposed a miniaturized high directive printed Yagi-Uda antenna with a meandered structure and a printed microstrip balun for long-range handheld UHF RFID reader systems" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001832_b978-0-08-097123-0.00015-0-Figure15.2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001832_b978-0-08-097123-0.00015-0-Figure15.2-1.png", + "caption": "FIGURE 15.2 Forces and moments acting on an azimuthing thruster in uniform flow.", + "texts": [ + " The resulting thrust from an azimuthing thruster is the sum of three components: T \u00bc TP \u00fe TD \u00fe TG (15.1) where TP, TD and TG are the component thrusts from the propeller, duct and the pod, respectively, and T is net unit thrust. As with any other propulsion device, the effective thrust acting on the ship is the net thrust adjusted by the augment of resistance (thrust deduction factor) induced by the unit on the vessel. These types of unit experience a complex system of forces and moments which are strongly dependent on the relative alignment of the unit to the incident flow as seen in Figure 15.2. The principal forces and moments which occur are Fx the longitudinal force in the propeller shaft direction. Fy the transverse force perpendicular to the propeller shaft. Q the propeller torque. Mz the steering or turning moment of the unit. All of these forces and moments are dependent both upon the inflow incidence angle d and the magnitude of the inflow velocity Va. In general, however, six components of loading {Fx, Fy, Fz, Mx, My and Mz} will be present. For design purposes two specific sets of model test data are commonly used" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002458_tvt.2012.2186991-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002458_tvt.2012.2186991-Figure2-1.png", + "caption": "Fig. 2. Tire slip angle.", + "texts": [ + " Using (11), an alternative expression for e2 is given by e2 = sin (\u03b8r \u2212 \u03b8) cos \u03b8r (13) and for e3, given that \u03b8 = \u03b8r (i.e., e2 = 0), one has e3| e2=0 = f \u2032\u2032(x) cos2 \u03b8r \u2212 tan \u03c6 L cos \u03b8r . (14) From (13), it is clear that e2 is the vehicle heading error (the error related to \u03b8(t)), and from (5) and (14), it is easy to see that e3 is the error of the vehicle front steering angle [\u03c6(t)]. OF SLIDING EFFECTS The kinematical model (1) is based on the common assumption of rolling without slipping. A tire slip angle is notated by \u03b1(t) and illustrated in Fig. 2, where vw(t) is the wheel linear velocity. It is the angle that is formed between the direction of wheel travel and the line of intersection of the wheel plane with the road surface [28]. When rolling without slipping is assumed, it means \u03b1(t) = 0,\u2200t. This assumption is reasonable for small-scale indoor mobile robots. However, in case of a four-wheel ground vehicle moving on a curved trajectory, tire slip angles always exist and have meaningful influence on the vehicle behavior in high speed maneuvers" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000858_aps.2015.7305629-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000858_aps.2015.7305629-Figure3-1.png", + "caption": "Fig. 3. (a) Top side of 8x8 Butler matrix antenna array; (b) Bottom side of 8x8 Butler matrix antenna array", + "texts": [ + " This Butler matrix was realized in the microstrip line form to facilitate easy integration with the tapered slot antenna adopted in the beam-switching array. In this paper, we adopted the tapered slot antenna as the antenna element in the beamswitching array. The structure of the antenna is shown in Fig. 2. A transition from the feeding 2481978-1-4799-7815-1/15/$31.00 \u00a92015 IEEE AP-S 2015 microstrip line at the backside of the substrate to the slotline was used. The realized circuit of the beam-switching tapered slot antenna array with the 8 x 8 Butler matrix is as shown in Fig. 3. The top side is the realized 8 x 8 Butler matrix circuit and the bottom side is the tapered slot antenna array. For 5G applications, the center operating frequency is set at 11 GHz. The circuit was fabricated on a Rogers RO4003 substrate, which has a dielectric constant of 3.55 and a thickness of 0.508mm. The spacing between the antenna elements was chosen appropriately to avoid grating lobes and to obtain better mainbeam patterns. The size of the integrated beam-switching array is 160 mm x 195 mm" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002055_sice.2014.6935219-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002055_sice.2014.6935219-Figure2-1.png", + "caption": "Fig. 2 Auxiliary linear sliding surface.", + "texts": [ + " thus, control law which constrain the state to the ellipsoidal sliding surface and lead the state to equilibrium point is derived as follows u = \u2212 {( b2 a2 \u2212 \u03b2r ) e\u2212 \u03b1r e\u0307\u2212 b2 a + r } \u2212Ksgn { e\u0307 ( (e\u2212a)2 a2 + e\u03072 b2 \u2212 1 )} (17) Since the ellipsoid is closed loop curve, the state is perturbed by the chattering that occurred near the equilibrium point, and not stay the equilibrium point. To make the state stay the equilibrium point, We defined auxiliary linear sliding surface in the ellipsoidal area which is near the equilibrium point. ellipsoidal area is given as follows (x\u2212 a) 2 a2 + x\u03072 b2 = q2 (18) Moreover, to reduce the jerk at the equilibrium point, We adjust the linear sliding surface from L0 to L1 in the area which is near the equilibrium point as shown Fig.2. In this section, we derive the convergence time from arbitrary initial state to equilibrium point along with the ellipsoidal sliding surface. Here, we represent the equation of ellipsoid Eq.(2) with sinusoidal function as follows.{ x = a (cos \u03b8 + 1) x\u0307 = b sin \u03b8 (19) From the differentiation the first equation of Eq.(19) with respect to time and the second equation of Eq.(19), we can derive Eq.(20) \u2212a sin \u03b8 d\u03b8 dt = b sin \u03b8 \u21d4 d\u03b8 dt = \u2212 b a (20) Now, we derive the traveling time from initial state P to equilibrium point O along with ellipsoidal trajectory" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002453_978-3-030-20131-9_174-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002453_978-3-030-20131-9_174-Figure4-1.png", + "caption": "Fig. 4. Flexible beam model with displacements.", + "texts": [ + ", no preload in the beam as the idler travels across the top-most point of the ellipse), the ellipse axes are given by be = Lbsin( ) (1) ae = (c2 + be 2)1/2 (2) According to the widely adopted pseudo-rigid body model of a compliant fixedpinned beam of length Lb under tip loading with large deflections (nonlinear beam) [17], it has been shown that the beam tip traces an approximately circular deflection path, with the center of that path located a distance Lb(1 \u2013 ) from the base of the beam, and an equivalent torsion spring K located at the circle\u2019s center (see Figures 2 and 4). In other words, the equivalent rigid-body beam has length Lb ( = 0.85 gives a relatively accurate approximation for a range of loading orientations [17]). The undeflected beam is initially oriented at an angle , and the tip of the deflected beam under tip loading is at an angle of , as in Figure 4. If the beam has an initial undeflected orientation (see Figure 4), and the beam/idler carriage is displaced a distance dl along the x-axis (see Figures 3 and 4), then the center of rotation of the equivalent beam is given by: xc = dl + Lb(1 \u2013 )cos( ) (3) yc = Lb(1 \u2013 )sin( ) (4) The idler location P (x,y) is then found as the intersection of the circle with the ellipse by solving the set of equations (x \u2013 xc)2 + (y \u2013 yc)2 = ( Lb)2 (5) (x/ae)2 + (y/be)2 = 1 (6) This allows approximation of the (negative-valued) beam deflection angle using x = xc + Lbcos( + ) (7) The orientation of the force applied by the belt on the idler F is then easily found by averaging the angles of the two (left and right) belt segments (i" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002889_aero.2014.6836466-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002889_aero.2014.6836466-Figure1-1.png", + "caption": "Figure 1 \u2013 Baseline SEP TDM SV delivers 20.7 kW (end-of-life) transiting from LEO to GEO and back. SEMS is accommodated on the RCS Module top deck and booms.", + "texts": [ + " [1,2] All non-payload subsystems have direct flight heritage and large performance margins to reduce risk, given risk posture of this Category 3, Class D mission. To reduce cost, Space Vehicle hardware is flight-ready through protoflight testing in flightlike environments or flight-qualified at the component level. The flight segment is built up from Bus, Reaction Control System (RCS), and SEP (Solar Electric Propulsion) Modules to streamline integration and test (I&T). Flight-proven, SV Bus and hydrazine propulsion Modules effectively support the SEP TDM. The Bus/RCS Modules mount on top of the SEP Module (Figure 1). Various technology enhancement options are possible for the SEP TDM. These options range from the inclusion of additional simple sensors, to experimental hardware enhancements, to operational enhancements up through full-up 978-1-4799-1622-1/14/$31.00 \u00a92014 IEEE Autonomous Rendezvous and Docking (AR&D) concepts that include separate host and docking vehicles (beyond our present scope). Per the NASA BAA requirements [3], mission enhancements are constrained to a maximum of an additional $100M. To fully understand the system impacts of electric propulsion operations in space, it is extremely important to characterize the environment around the SV to understand how the propulsion system may perturb measurements of the ambient plasma, ion and interplanetary magnetic field [5-7]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000673_tie.2018.2842736-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000673_tie.2018.2842736-Figure5-1.png", + "caption": "Fig. 5 FE model of 10pole/12slot and 10pole/30slot MFM-BDRMs: (a) 10pole/12slot MFM-BDRM; (b) 10pole/30slot MFM-BDRM", + "texts": [ + " Based on the above analysis, we can see that, for the MFM-BDRM with the stator pole pair being 4, the 8pole/24slot MFM-BDRM is obviously superior to the 8pole/9slot MFM-BDRM in no-load and load back EMF, torque ripple, maximum torque outputting capability and power factor. B. Specific FSCW of MFM-BDRM with 2 2sQ p To validate the above analysis, a specific FSCW of MFM-BDRM with 2 2sQ p is also designed. The stator pole pair is 5, and the corresponding stator slots are 12. Its finite element model is shown in Fig. 5(a). Meanwhile, to comprehensively understand the performance of specific FSCW of MFM-BDRM with 2 2sQ p , a integer-slot winding of MFM-BDRM with the same size is design to compare their electromagnetic performances, as shown in Fig. 5(b). Specific parameters of two schemes are listed in Table III. TABLE III PARAMETERS OF FOUR FE MODELS OF MFM-BDRMS Parameters 10pole/12slot MFM-BDRM 10pole/30slot MFM-BDRM Stator pole pair 5 5 Stator slots 12 30 PM rotor pole pair 18 18 magnetic blocks 23 23 Outer diameter of stator (mm) 216 216 Axial length (mm) 70 70 Thickness of PMs (mm) 4 4 Series Turns per phase 80 80 Maximum current (A) 28.5 28.5 Maximum voltage (V) 350 350 According to (2) and (3), when 12Q , 2N . According to (5) and (12), the stator winding coefficient can be expressed as \u03c0 ( \u03c0) 6sin 2\u03c0 \u03c0 \u03c02sin cos( ) \u03c012 2( \u03c0) 62sin 2 dp p d v k k k (18) Based on (18), the calculated absolute value of winding coefficient of any order of harmonic magnetic field is listed in Table II" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001046_062024-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001046_062024-Figure3-1.png", + "caption": "Figure 3. Front sensor module. (a) Overall appearance drawing. (b) Sectional view during treatment. (c) Sectional view during inspection. (d) Exploded view.", + "texts": [ + " Figure 2(c) is a crosssectional view of the connection between the middle module and the front module. The two modules are connected by threads. The stop part is used to avoid the thread looseness caused by vibration and improve the reliability of the connection. ISCME 2020 Journal of Physics: Conference Series 1748 (2021) 062024 IOP Publishing doi:10.1088/1742-6596/1748/6/062024 The function of the front sensor is to judge the stiffness of the patient's muscles by reading the value of the sensor 2 during detection. In figure 3(a), the thread of the front sensor outer casing is used to connect with the middle module, and the center hole is used to pass the iron core that applies impulse force. The two symmetrically arranged synapses are the circuit connection points between the front module and the middle module. In figure 3(b), when treating a patient, the iron core of the middle module hits the central slider and the sensor, and the central slider transmits the pulse force F1 to the human body, triggering the resonance of the human tissue to achieve the therapeutic effect. In figure 3(c), when detecting a patient, the right end of the center slider is close to the patient through the massage head. When the iron core hits the center slider, the center slider slides to the right under the impact force F2. The resultant force generated by the body tissue and the compression force of the spring form a resultant force F3, which forces the central slider to slow down. Sensor 2 judges the stiffness of the patient's muscles by detecting the high and low strength caused by the deceleration of the center slider. The appearance and dimensions of each part are shown in figure 3(d). ISCME 2020 Journal of Physics: Conference Series 1748 (2021) 062024 IOP Publishing doi:10.1088/1742-6596/1748/6/062024 The design purpose of the middle working module is to provide the impulse force required by the treatment device when working, and to connect the front module and the rear module. In figure 4(a), the bracket is subjected to the rightward thrust F exerted by the rear module during assembly, and this force cooperates with the limit slot to maintain the bracket at the right end of the middle module in the non-working state" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001244_tasc.2016.2543267-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001244_tasc.2016.2543267-Figure7-1.png", + "caption": "Fig. 7. (a) Case separation standard (b) Case 1 (C) Case 2", + "texts": [ + " In the skew model, the same amount of axial force is produced in the opposite direction when the rotating direction is changed. The distribution of the magnetic flux density is upside down, however, the performance characteristic is technically identical regardless of the rotating direction. Otherwise, the ends of the rotor have different magnetic flux distributions according to the rotating direction in the V-skew model. Consequently, the performance and loss characteristics differ according to the rotating direction. The case varies with identical motors according to the rotating direction as shown in Fig. 7. In this paper, characteristic analysis on every case is conducted based on FEA. Magnetic flux distribution, torque, and axial force distribution are considered. Since the input condition and the stator are identical, copper loss is the same for all cases. Therefore, iron loss is considered as well. Magnetic flux density distribution is shown in Fig. 8. The legend of the contour plot is from 0 to 2 T. According to the rotating direction, skew angle is changed as shown in Fig. 11. The more saturated portion is located at the ends of the motor in Case 1 and at the center of the motor in Case 2" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000734_measurement.2017.7983556-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000734_measurement.2017.7983556-Figure1-1.png", + "caption": "Fig. 1. a) The concept of Yagi-Uda antenna. R \u2013 reflector, DE \u2013 driven element, D \u2013 director, DR \u2013 distance between the reflector and driven element, DD \u2013 distance between the director and driven element, LR \u2013 length of the reflector, LDE \u2013 length of driven element and LD \u2013 length of the director. b) Technical", + "texts": [ + " The passive elements are divided into two types. The first type of passive elements causes focusing of the electromagnetic wave radiated from driven element. These elements are called as directors. The second one of passive elements causes reflection of the electromagnetic wave radiated from driven element. These elements are called as reflectors. Different features of directors and reflectors are given by distance from driven element and length of particular elements. The concept of Yagi-Uda antenna is in Fig. 1. a). The number of directors determines the features of radiation. The more directors cause improvement of directivity and gain. The simplest Yagi-Uda antenna consists of three elements, one driven element, one reflector and one director. The maximum directivity obtainable from this antenna is about 9 dBi (7 dBd) [1]. The tested antenna has been designed as three elements Yagi-Uda antenna for 1 GHz and was printed on printed circuit board because of used frequency. The Yagi-Uda antenna, which has been implemented, is in Fig.1. b). implementation of designed three elements Yagi-Uda antenna for 1 GHz. The parameters, which have been measured in particular, are bandwidth, VSWR (voltage standing wave ratio), directivity, gain and efficiency of the antenna. The bandwidth is frequency range, where antenna maintained desired features. The bandwidth is joined with VSWR. VSWR equals to 1 means, that antenna is perfectly matched to the transmission line and no power is reflected. It means that it is reasonable to choose small values VSWR" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001030_20.703862-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001030_20.703862-Figure2-1.png", + "caption": "Fig. 2. Coordinate relationships of the fixed and moving coordinates.", + "texts": [ + " In order to limit the mathematical efforts and yet gain physical insight regarding the problem, the following assumptions are made in this study. First, both the rotor and stator surfaces are perfect cylinders, and the axis of the rotor is parallel to the axis of the stator, i.e., only translatory eccentricity is considered so that the problem is still twodimensional. Second, the effects of slotting are neglected. Third, the permeability of the rotor and the stator back iron is infinite. And last, eddy current and saturation effects are neglected. An arbitrary point in the air gap or the permanent magnet region in Fig. 2 can be expressed in terms of either \u2013 or \u2013 coordinates, which can be transformed into each other as follows: (3a) (3b) Throughout the following analysis, only linear components in perturbation equations are retained, which is not unusual for most perturbation analyzes. The detailed derivation of the transformation equations is given in Appendix I. A static field is a field in which the effects of inertia can be neglected. In a magnetic field, a static condition occurs when the propagation time for an electromagnetic wave within a device is small compared to the characteristic time associated with the motion of the device", + " The governing field equations and the associated boundary conditions are formulated and solved by the aid of a perturbation method which is introduced to treat the nonlinear boundary conditions caused by rotor eccentricity. The perturbation analysis is validated by the corresponding finite element analysis. A regular first order perturbation solution is sufficient for obtaining good results compared to those of the finite element analysis. Comprehensive understanding of the magnetic field induced by rotor eccentricity will help to develop motors with better design and performance. APPENDIX I DERIVATION OF COORDINATE TRANSFORMATION (3a) AND (3b) Consider the coordinate relationship detailed in Fig. 2 in the domain of the air gap or the magnet pole region. In the figure, let be a point defined by the polar coordinates and with respect to the origin in the fixed coordinate system. It is necessary to be able to express and as a Fourier series of and . By drawing the perpendicular line from upon the line , the following expression is produced: or where Rewriting the above equation yields or simply Taking logarithms on both side of the above equation yields The right-hand side can be expanded in a convergent series" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002888_s11665-018-3265-2-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002888_s11665-018-3265-2-Figure3-1.png", + "caption": "Fig. 3 (a) FEM model of ECSEE, and (b) the coordinate axis systems of samples and a crystal", + "texts": [ + " If the value of m were much too large, the ellipse would not be transformed back to a round shape for the cold extrusion forming. If the value were much too small, the deformation efficiency would be reduced. Moreover, the contact friction factor of die, k, is cardinal to the shear strain. Therefore, the friction factor k, torsional angle u, and ratio m were picked out as the design objects. The simulation software DEFORM 3D v6.1 was used for the reliability evaluation of the ECSEE technique (see Fig. 3(a)). The diameter and length of the as-deformed pure copper sample are 10 mm and 35 mm, respectively. The related parameters of ECSEE die were set as follows: (1) k = 0.008, 0.2, 0.3, 0.5, 0.7 for u = 90 , m = 1.5; (2) u = 30 , 45 , 60 , 75 , 90 for k = 0.1, m = 1.5; and (3) m = 1.1, 1.2, 1.3, 1.4, 1.5 for k = 0.1, u = 90 . The structure parameters of the ECSEE die are as follows: L1 = 7 mm, L2 = 10 mm, and L3 = 7 mm. Journal of Materials Engineering and Performance Die and punch were assumed as rigid bodies, whereas the copper billet was considered to be a deformable one", + " More information regarding the details of setting the parameters is available in the literature (Ref 5). Pure copper prepared by annealing treatment was selected as the ECSEE experimental samples. The extrusion process was performed on the YA32-315-type hydraulic machine, with an extrusion rate of 1 mm/s. The main structural parameters of the ECSEE die cavity were chosen as follows: u=120 , m=1.55, L1 = 7 mm, L2 = 10 mm, and L3 = 10 mm (Ref 5). After a single-pass extrusion test, the samples were cut into 12 samples along the longitudinal section (see Fig. 3(b)). The received samples were first mechanically polished and then electrolytically polished. Electro-polishing was carried out in 500 mL distilled water + 250 mL phosphoric acid at room temperature for 6-8 min at 4 V. An EBSD analysis test was conducted on a TOUANATA400-type field emission SEM. The PASS 5 software was used for data handling. In this study, the textures are represented in the TD-ND-ED reference system (see Fig. 3(b)). The normal direction (ND, perpendicular to the extrusion direction) and the extrusion direction (ED) are represented as X0 and Y0. As illustrated in Fig. 1(a), the extruded sample is distorted on the ED\u2013ND plane during the ECSEE deformation in Channel L2. The TEM samples were prepared in the following sequence: sectioning, manual grinding, punching (wafer with a diameter of 3 mm), grinding pits, and ion milling. TEM observation was conducted on a Tecnai G2 F30 high-resolution transmission electron microscope at the accelerating voltage of 300 kV. A series of positions were selected in the periphery regions of the received sample for recording the microstructure evolution (shown in Fig. 3(b)). Journal of Materials Engineering and Performance Vickers micro-hardness tests were employed by using 100 g loads and a dwell time of 15 s with a HXP-1000TM microhardness tester. The micro-hardness measurements at 50 selected points were investigated along the lateral line and the central line of the longitudinal sections. For each selected point, three measurements were performed, and the average values were then used. Normally, the friction in the plastic deformation is detrimental to the billet forming process (Ref 12)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003423_ecce.2019.8913146-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003423_ecce.2019.8913146-Figure6-1.png", + "caption": "Figure 6: Single stator single rotor AF-DFIG", + "texts": [ + " [18] presents the value of for a dual stator single rotor (DSSR) permanent magnet AFM. The power density can be expressed as (3), where is the outer diameter of the machine including the end windings. On determination of the effective stack length of the machine, the stator and rotor core lengths can be figured out as described in [18-19] by considering end winding lengths. The optimal air-gap length can be determined after careful consideration of manufacturing tolerances, the outer dimension of the machine and power requirements. (3) Fig. 6(a) shows an isometric view of the modeled structure in 3D FEA, while Fig. 6(b) shows a top view of the full machine with the windings going into the slots of both stator and rotor. Fig. 7 shows the winding configuration used in the machine. Initial design characterization can be carried out with a 2D model of the AF-DFIG considered at the mean air-gap diameter, as shown in Fig. 8, for fast analysis. Though 2D modeling enables faster optimization and easier sensitivity analysis, loss prediction might be inaccurate. The authors figured 3D analysis would better suit the need of the work which is to compare the performance of AF-DFIG to its\u2019 radial flux counterpart" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001568_s13246-016-0502-6-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001568_s13246-016-0502-6-Figure3-1.png", + "caption": "Fig. 3 View of length adjustment mechanism designed in CATIA software", + "texts": [ + " The structure of the designed exoskeleton was made of aluminum with the mass of 5\u00a0kg. This frame had a male and female connection to adjust the length of thigh and calf in a way that female aluminum proile was designed with dimensions of 35 \u00d7 5 \u00d7 3 cm3 for the femur and 25 \u00d7 5 \u00d7 3 cm3 for the tibia. Also, the dimensions of the male proile were considered as 2 \u00d7 4 \u00d7 20 cm3 which was the same for both bones and facilitates the adjustment of length [9]. To simulate the extension and lexion movements of the leg joint, two bearing were used at the end of these joints (Fig.\u00a03). A compact plastic was embedded in exoskeleton in order to remove the load from the human sole. The end side of the plastic was curved that let the person to continue his movement in stance phase. The sole was attached to the ankle by hinge that can simulate abduction and adduction movements of the ankle (Fig.\u00a04). Human gait cycle begins when the toes leave the ground and ends when heels touch the ground (swing phase). The motion span of the hip joints varies from \u221215\u00b0extension to 30\u00b0 lexion. Also, this span is between 0\u00b0 extension and 60\u00b0 lexion in the case of the knee joint (Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002724_12.575416-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002724_12.575416-Figure2-1.png", + "caption": "Fig. 2 the parallel beam scanning mechanism", + "texts": [ + "1 optical sysytem scanning mechnaism detector preamplifier video processor indicator synchronizing mechnaism refrigerator object radiation Fig.1 The schematic diagram of thermal imager system Infrared thermal imager can change infrared radiation into distributed image, so an observing tool were gotten, which can 254 Proc. of SPIE Vol. 5640 Downloaded From: http://proceedings.spiedigitallibrary.org/ on 07/15/2016 Terms of Use: http://spiedigitallibrary.org/ss/TermsOfUse.aspx This paper used rotating mirror to do line scanning and swing mirror to do the frame scan. Take Fig.2 for example, this mechanism use parallel beam to scan. Reflecting mirror is an octahedral prism. The basic measurement of this mechanism depends on the efficient width of light beam and the viewing field of system. The schematic diagram of the mechanical system is presented in Fig.2. This system includes four parts: the condenser system and opticmechanical scanning system, the detector array, the light-emitting diode array and signal process system, image pick-up and display system. The optical system is presented in Fig.3. 3. THE INFRARED DIAGNOSIS OF TRANSFORMER FAULT The faults of the transformer include fault connecting of conductor, the eddy current of housing, which is caused by magnetic leakage, and the fault of the cooling apparatus. 3.1 The infrared diagnosis of the fault connecting of exterior part of conductor When the transformer and the exterior current-carrying conductor is fault connecting or loose, the local part will become overheat because of the increase of the resistance" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003504_icwt.2015.7449220-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003504_icwt.2015.7449220-Figure1-1.png", + "caption": "Fig. 1. Microstrip [6]", + "texts": [], + "surrounding_texts": [ + "A curved microstrip line can be modeled as a cascade of sections of microstrip lines with chamfered bends. Illustrated in Figure 3a is a typical bend in a microstrip line for an arbitrary bend angle also shown in the same gure are the reference planes that dene the edges of the bend. The equivalent circuit of the bend, in the region restricted to the connes of the reference planes, is shown in Figure 3b [2]. Figure 3 shows the equivalent-circuit representation of the bend in the corner of resonator, the inductance L represent the association of discontinuity. For optimum chamfer, the ration of the width of the chamfered region to the width of the microstrip line is approximately 0.5 [2]. Figure 4 explain the effect of the various chamfer side from 0 mm to 8mm toward reflected power. The 1.3 mm chamfer side and 1 mm resonator width generates the optimum magnitude response with return loss \u00a1 27dB. Inductance affected by resonator length and characteristic impedance [7]. L = 0.0847Z0l nH (5) The chamfer side inversely proportional to the inductance effect that generates higher reflected power." + ] + }, + { + "image_filename": "designv6_24_0000535_0954406214544726-Figure10-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000535_0954406214544726-Figure10-1.png", + "caption": "Figure 10. Typical temperature contour on the outer surface of rotating bowl at 45 s.", + "texts": [ + " It is proposed to divide in to three segments spacing 120 each as shown in Figure 8. A quantitative measurement of experimental transient temperatures histories at three thermocouple locations starting at 0 and ending 120 is presented in Figure 9(a). Similarly, during welding of the second segment in addition to those three thermocouples, previous thermocouple data (which is at 90 from weld starting point) are also plotted as shown in Figure 9(b). A typical result of thermal simulation is presented in Figure 10, in which the 3D temperature profile on the outer surface at 45 s for the simulation condition of point\u2013point clamping condition with 0.5mm/s weld speed for the optimum sequence of Case I mentioned in Table 4. Figure 11 shows a typical contour of longitudinal distortion of the bowl cooled to room temperature after welding. The comparison of numerical results to those experiments before and after welding are presented in Figure 12 showing reasonably a good agreement between them. It can be observed from Figure 13(a) that the distortions are high at higher welding speeds for an optimum sequence and clamping conditions" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001472_9781118354179.auto089-Figure21-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001472_9781118354179.auto089-Figure21-1.png", + "caption": "Figure 21. Racing friction disc with\u2014sintered paddles. (Reproduced by permission of Alcon Components Limited)", + "texts": [ + " They typically use composite organic friction materials, phenolic resin bonded with special fibers, fillers, and frictional modifiers. Such materials provide good torque control, stable coefficient of friction within low to medium temperature range, long lives, and refined driving. They work well with gray cast iron or spheroidal iron pressure plates and a spheroidal iron or steel flywheel. For racing applications, where performance has priority over refinement characteristics and long lives, sinter friction materials are sometimes used, with the friction disc shown in Figure 21 (Alcon). Sinter inserts (\u201cpaddles\u201d) can be either bonded or riveted to the disc plate. In such vehicles, flywheels and pressure plates are typically made of steel, with the emphasis on high rotational speeds, low mass, and good thermal stability of the coefficient of friction across wide range of temperatures. At the very top of motor racing, multi-plate clutches are used, as shown in Figure 9. In such designs, both friction discs and pressure plates are manufactured from special carbon fiber materials, able to withstand high temperatures, providing high coefficient of friction values" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000827_6.iac-03-v.3.03-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000827_6.iac-03-v.3.03-Figure5-1.png", + "caption": "Figure 5. Interior Views of the OSP Crew Cabin (Top: Five Crew Upright During Ascent, Center: Three Deconditioned Crew Laying Down For Re-entry, Bottom: Seven Deconditioned Crew in an Emergency Return Configuration)", + "texts": [ + " The BLB is very similar to a lifting body, but with many additional advantages. The dominant feature of the BLB is that the wing and fuselage are blended together. This blend provides ample unpressurized volume for subsystem packaging while contributing to the aerodynamic performance of the vehicle. Blended Lifting Body Volume The scale and shape of Orbital\u2019s OSP enable a large crew cabin that easily accommodates five astronauts sitting upright during ascent, and seven astronauts on the return mission as shown in Figure 5. The OSP crew cabin has enough additional volume to carry 16 middeck lockers for the storage of crew supplies. The BLB shape also provides sufficient unpressurized volume for subsystem packaging, at a reasonable scale and with sufficient margin. Blended Lifting Body Shape Performance Over the course of the shape trade study, the design team encountered many technical issues associated with various shapes and developed possible solutions for each. The features of the BLB, shown in Figure 6, were selected based on the solutions found to best enable an OSP" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002162_icecs.2008.4675079-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002162_icecs.2008.4675079-Figure1-1.png", + "caption": "Fig. 1. Typical SC integrator.", + "texts": [ + " the relaxed requirements on the antialiasing filter, the possibility to interchange speed for accuracy and the robustness against capacitor mismatch. This paper studies the impact on \u03a3\u0394-modulators of another parasitic effect, i.e. dielectric relaxation. A simple way to think of this phenomenon is to imagine that after a short discharge of an initially charged capacitor a voltage will build up if you leave it floating. Looking at a typical circuit implementation of a \u03a3\u0394-modulator with switched-capacitor (SC) integrators, like the one shown in Fig. 1 1 it is clear that the capacitors Cin and CFB are constantly charged and discharged. Hence, there could be an impact of dielectric relaxation on a \u03a3\u0394-modulator. We have found references to research performed on other type of converters like the SAR and pipeline ADC [3]\u2013[5], but to the author\u2019s knowledge there is no previous research reporting on the influence on \u03a3\u0394-modulators. This paper is structured as follows. After this introduction we will set up a test-case \u03a3\u0394-modulator, to simulate the impact of dielectric relaxation", + " The system level model of this (ideal) \u03a3\u0394-modulator is shown in Fig. 2. Like (almost) all (lowpass) \u03a3\u0394-modulators it contains a low resolution quantizer (in our case 1-bit) and discrete time integrators which make up the loopfilter of the modulator. The design and optimisation of the NTF(z) was done through a proper selection of the coefficients ai and g in Fig. 2, using standard design techniques [6], [7] for an OSR of 32. A typical circuit implementation of the integrators of Fig. 7 uses a the SC-integrator of Fig. 1. As can be seen there are two phases in the operation of the circuit. During a first phase, \u03c61, a voltage v1(t) is sampled onto capacitor Cin. During a second phase, \u03c62, the charge on capacitor Cin is transferred to the feedback capacitor (CFB) where it adds to the charge already present on CFB . With the charge conservation rule we find: 978-1-4244-2182-4/08/$25.00 \u00a92008 IEEE. 1221 Cinv1(kT ) + CFBvout(kT ) = Cinv2((k + 1)T ) + CFBvout((k + 1)T ) (1) Or, in the z-domain, we derive: Vout(z) = Cin CFB 1 1 \u2212 z\u22121 ( z\u22121V1(z) \u2212 V2(z) ) (2) Hence, from a system level point of view, a SC-integrator can be modeled like shown in Fig", + " The actual values of the resistors Ri and the capacitors Ci were based on the extracted parameters for the MIM capacitors in the AMIS I3T technology [2], [8] and hence the setup is realistic. Finally, with a last degree of freedom, we ensured that the time constants of the RiCi branches were chosen such that all the important timings in a \u03a3\u0394-modulator were covered: from the low frequencies of the input signal to the higher frequencies of the clock. For our simulations all integrators of Fig. 2 were replaced by the circuit implementation of Fig. 1 and the resulting SCnetwork was next simulated in the spectre environment. All circuit components but the capacitors were ideal, including the OTAs, switches, 1-bit quantizer, non-overlapping clock generation circuit and voltage sources. We also made sure to have complete settling of the voltage within half a clock period of time. The input signal of the modulator was a sine wave at 1/8 of the frequency bandwidth of the \u03a3\u0394-modulator and an amplitude of 0.2 times the reference voltage of the quantizer", + " Only for the aggravated case of 100 times, one can see two smaller effects: the NTF-zero shifts to lower frequencies and the Q-factor is smaller. Even for this case for the important low frequencies the spectrum still rises with 20db/dec. In order to be able to generalise the results on our test case modulator to other \u03a3\u0394-modulators we are now going to investigate the impact of dielectric relaxation from a more theoretical perspective. One way could be to try and find the transfer function of the SC-integrator circuit of Fig. 1 with the ideal capacitors replaced by the model of Fig. 4. However, due to the resistors in the model of Fig. 4 the analysis of the resulting SC network is very difficult. Therefore, we need to use another approach. The approach taken here is similar to the one described in [3]\u2013[5] where the dielectric relaxation is modeled as a pure memory effect: 1) a capacitor charged to a voltage Vinit 2) shorted from t = 0 to t = t0 3) and left floating from t = t0 to t = tf , will show at t = tf a voltage over its two terminals equal to: VC = \u03b2 ln(tf/t0)Vinit = \u03b3Vinit (3) In fact this model corresponds quite well to how the dielectric relaxation intuitively works", + " It seems indeed that the classical charging and shorting story cannot directly be applied, since one terminal is always connected to a floating node (inverting node of the amplifier). Nevertheless dielectric relaxation is to be expected since the voltage on the feedback capacitor is changing all the time and dielectric relaxation is the phenomenon working against such kind of variations. Hence, the situation for this capacitor was investigated more closely by performing spice simulations on the SC-integrator of Fig. 1 with the capacitors replaced by the model of Fig. 4. We found out that during \u03c61, when the capacitor is put in feedback around the operational amplifier, part of the charge that is on the feedback capacitor leaks away. Moreover, we found that the charge on the feedback capacitor at the beginning of the charge transfer phase \u03c62 equals: CFBvout(kT ) \u2212 CFB\u03b3FB(vout(kT )\u2212 vout(kT \u2212 T )) (4) instead of CFBvout(kT ), like for the ideal case. 2E.g. with \u03b1 = 1, for our case given the Ri and Ci we could derive that this \u03b3 is approximately equal to 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001244_tasc.2016.2543267-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001244_tasc.2016.2543267-Figure8-1.png", + "caption": "Fig. 8. Magnetic flux density distribution", + "texts": [ + " Consequently, the performance and loss characteristics differ according to the rotating direction. The case varies with identical motors according to the rotating direction as shown in Fig. 7. In this paper, characteristic analysis on every case is conducted based on FEA. Magnetic flux distribution, torque, and axial force distribution are considered. Since the input condition and the stator are identical, copper loss is the same for all cases. Therefore, iron loss is considered as well. Magnetic flux density distribution is shown in Fig. 8. The legend of the contour plot is from 0 to 2 T. According to the rotating direction, skew angle is changed as shown in Fig. 11. The more saturated portion is located at the ends of the motor in Case 1 and at the center of the motor in Case 2. Higher axial force is obtained according to (8) with a more magnetically saturated model, since more flux flows out from the core with a saturated core. (a) Conventional skew (b) V-skew case 1 (c) V-skew case 2 Fig.11. Magnetic flux density distribution according to the skew angle Numerical analysis on each model is conducted with a sinusoidal current source" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001402_j.mechmachtheory.2013.12.016-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001402_j.mechmachtheory.2013.12.016-Figure2-1.png", + "caption": "Fig. 2. T time t. T", + "texts": [ + " FromR \u00bc I \u00fe \u03b5e\u03c1 Q results that the determinant of R is given by detR \u00bc detQdet I \u00fe \u03b5e\u03c1 , which combined with detR = 1 leads to detQ = 1. \u25a1 he motion of the rigid body. The body fixed reference frame in the initial position and orientation at time t0, and in the current position and orientation at he point P is attached to the rigid body. Remark 2. Based on the construction of SO3 and the multiplication of dual tensors (Eq. (C.3)), a direct conclusion is the Lie group structure of Eq. (36). This Lie group parameterizes all rigid motions. Thus, a rigid body motion (Fig. 2) [23] can be modeled using Eq. (38). Based on Theorem 4, a representation of any dual tensor from SO3 can be given: Theorem 5. For any orthogonal dual tensor R defined as in Eq. (38), a dual number \u03b1 \u00bc \u03b1 \u00fe \u03f5d and a dual unit vector u \u00bc u\u00fe \u03b5u0 exists in order to have the following expression where From which leads The du while R \u00bc I \u00fe sin\u03b1eu \u00fe 1\u2212cos\u03b1\u00f0 \u00deeu2 ; \u00f040\u00de u and \u03b1 are recovered from the linear invariants of Q, while d = \u03c1 \u22c5 u and u0 \u00bc 1 2 \u03c1 u\u00fe 1 2 cot \u03b1 2 u \u03c1 u\u00f0 \u00de \u03b1\u22600; 1 2 \u03c1 u \u03b1 \u00bc 0; : 8><>: Proof" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003087_temc.2008.2004052-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003087_temc.2008.2004052-Figure2-1.png", + "caption": "Fig. 2. (a) Monopole antenna mounted on a metal box positioned next to a flat phantom and (b) geometry of a dual-band meander monopole antenna operating at 900 and 1900 MHz. The flat phantom dimensions are shown for frequencies of 900 MHz and above, according to [36] and [37]. h is the distance from the phantom material to the antenna feedpoint.", + "texts": [ + " 1) Dipole Antennas: Linear wire dipole antennas of lengths \u03bb/15, \u03bb/8, \u03bb/4, and \u03bb/2 were studied at 300, 450, 900, 1450, 1900, 2450, 3700, and 6000 MHz in free space and next to the flat phantom. The distances between the antenna feedpoint and the phantom material are h = 5, 10, and 20 mm (see Fig. 1). The dipole wire radius is 1.8 mm. 2) Monopole Antennas: Quarter-wave linear monopole antennas at 300, 900, 1900, 3700, and 6000 MHz were mounted on the center of the top face of a 100 \u00d7 40 \u00d7 19 mm3 metal box resembling a conventional portable wireless device [see Fig. 2(a)]. Helical monopoles (14 turns, 48-mm axial length, 4 mm diameter) and printed meander monopoles (7 mm pitch and 51-mm axial length, printed on RO 4003 c substrate), both operating at 900 MHz, were also investigated. A dual-band meander antenna operating at 900 and 1900 MHz was also analyzed [see Fig. 2(b)]. The meandered branch is responsible for the resonance at the low frequency while the straight strip is responsible for the resonance at the high frequency. Antennas were studied in free space and next to the flat phantom (h = 12 and 20 mm). Due to the thickness of the metal box and the phantom shell, the smallest distance that could be used was h = 12 mm. 3) Planar Antennas: All planar antenna models were mounted on a metal box of dimensions 100 \u00d7 40 \u00d7 10 mm3 . They were modeled and measured both in free space and next to the phantom" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001993_pccon.2007.372915-Figure15-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001993_pccon.2007.372915-Figure15-1.png", + "caption": "Fig. 15: Application of the indirect air cooling in a 5kW telecom power supply with lOkW/dm3 and 1U height.", + "texts": [ + " Based on this concept a transformer for a 5kW telecom power supply is redesigned what results in an three times increase of the power density of the transformer and an improved efficiency. With a HTC made of heat pipes a maximum power density of 55kW/dmi3 for the transformer with cooling system are achieved without cooling via the surface of the transformer. Based on the indirect air cooling system a redesign of the telecom supply has been performed, where also the topology and the switching frequency have been optimised. There, a power density of l0kW/dmi3 and an overall height of 1U (40mm) could be achieved (cf. figure 15). [1] J.T. Strydom, \"Electromagnetic Design of Integrated ResonatorTransformers,\" Ph.D. Thesis, Rand Afrikaans University, South Africa, 2001. [2] I.W. Hofsajer, \"On electromagnetic Integration in Hybrid Electronic Structures,\" D. Eng. Thesis, Industrial and Electronics Research Group, Faculty of Engineering, Rand Afrikaans University, May 1998. [3] J. Biela and J.W. Kolar, \"Analytic Design Method for (Integrated-) Transformers ofResonant Converters using Extended Fundamental Frequency Analysis,\" Proceedings of the 5th International Power Electronics Conference (IPEC), Niigata, Japan, April 4 - 8, 2005" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003717_induscon.2016.7874600-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003717_induscon.2016.7874600-Figure2-1.png", + "caption": "Figure 2 - Structure of the machine designed.", + "texts": [ + " (3) With: = rotor velocity [rpm]; = electric frequency [Hz]; = number of poles; = motor electric position; = motor mechanical position; The three-phase motor developed for this study has an axial structure. The stator has 9 windings connected in series three by three, which possess plastic supports placed in a circle. The wire used in the windings is the 18AWG, 0.815mm2. The rotor is formed by 16 Neodymium magnets fixed in two metallic bases using epoxy resin and glass fiber. At each base, the poles are arranged alternately, as shown in Figure 1. Figure 2 illustrates the structural scheme developed. Each stator phase is represented by a different color. Figure 3 shows the final assembled machine in laboratory. In order to supply the machine with a sinusoidal wave, it is necessary that each branch of the three-phase inverter \u2013 shown in Figure 4 \u2013 to be triggered independently. Analyzing one of the inverter branches, S1 and S2 for example, each switch receives a complementary command signal, which avoids a short circuit on the power supply Vcc. At least three current sensors must be used because a direct measurement of current in each motor phase is required" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002202_ecctd.2007.4529551-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002202_ecctd.2007.4529551-Figure1-1.png", + "caption": "Fig. 1. Bonding Wire in IC package.", + "texts": [ + " Attendant upon a scaling of semiconductor process in recent years, an operating voltage of a signal processing integrated circuit such as CPU has decreased steadily and finally it became lower than 1[V]. Because of this, a demand for a new output voltage precision appeared vis-a-vis the power circuit. For example, if the power supply voltage is 0.9[V], 2% precision of an output voltage of a LDO is permitted only to 18[mV]. With a high load current flows through a LDO which uses a power MOS transistor, it\u2019s impossible to ignore a resistance of a bonding wire. Fig. 1 shows the Au (Gold) bonding wire in SOT-23 package. It lies between an IC pad and a package pin, and supposed having an impedance of 0.556[nH] and 92m\u2126 [1]. Supposed that a LDO is designed for a load current of 300[mA], then with the above bonding wire impedance, a drop between the output pin and the IC pad will be 27.6[mV]. This drop voltage preponderantly exceeds the permitted range of 2% fluctuation of the LDO output voltage. A compensation for this drop voltage by some ways becomes necessary" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001188_speedam.2012.6264408-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001188_speedam.2012.6264408-Figure3-1.png", + "caption": "Fig. 3. Basic machine design with surface mounted magnets", + "texts": [ + " -5 -3 -1 3 5 -5 30 6,24 12,18 -12,-18 -6,-24 x -3 6,24 6 6,12 -6,-12 x -6,-24 -1 12,18 6,12 6 x -6,-12 -12,-18 1 -12,-18 -6,-12 x 6 6, 12 12,18 3 -6,-24 x -6,-12 6,12 6 6,24 5 x -6,-24 -12,-18 12,18 6,24 30 As the magnitude of one particular magnetic field component can either have a positive or a negative sign, the table entries which show combinations of two magnetic field components of same order and different signs are irrevelant and therefore marked with \"x\". All cogging torque harmonics are related to the frequency of one electrical period. The sign of a cogging torque harmonic determines, if cogging torque harmonics of same order are in phase or in antiphase. When analyzing one magnetic field harmonic by itself, it can be seen that only one cogging torque harmonic arises, while two cogging torque harmonics are generated, when two different magnetic field harmonics are considered. III. ApPLICATION EXAMPLE 1: BRUSHLESS PERMANENT In Fig. 3, a motor with an interior rotor with surface mounted magnets is shown. The stator consists of twelve teeth (Ns = 12) while the rotor features eight poles (pz = 4). By using Table I, it can be seen that the magnetic field harmonics of order 1 and -3 of the magnetic field each generate cogging torque harmonics of order 6. However, the combination of those two field harmonics gener ates a cogging torque of order 6 with opposite phase (negative sign in Table I). As the arising cogging torque harmonics are caused by differ ent permeance harmonics, it could be possible to reduce the overall cogging torque peak-to-peak value by adding a mag netic field harmonic of order -3 to the magnetic fundamental wave (order 1)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003464_ijvd.2017.082579-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003464_ijvd.2017.082579-Figure3-1.png", + "caption": "Figure 3 Design variables for shape optimisation (see online version for colours)", + "texts": [ + " As can be seen from Tables 1 and 2, HCSSNM gives the best results reported in the literature for welded beam design problem. The hybrid approach proposed in Section 3 is applied to optimal structural design of an automobile suspension arm taken from automotive industry. Initial design domain and boundary conditions of the automobile part is given in Figure 2. Minimisation of weight is chosen as objective function. Fatigue life is chosen as constraint function in this problem. Input variables for the meta-models are the five design variables which are x1, x2, x3, x4 and x5 as shown in Figure 3. Initial values, lower and upper limits of the design variables are provided in Table 3. In this study, LHS was used to sample the design space for a total of 50 training. Weight and fatigue life are calculated for each 50 experiment. The optimisation problem is formulated as follow: 1Min ( ) ( )F x f x= (13) 2( ) ( ) 1 6g x f x e= > , 1, ,l u i i ix x x i NDV\u2264 \u2264 = where f1(x) and f2(x) represent the weight and fatigue life values as objective function and constraint, respectively. The surrogate models of the objective function and constraint function are constructed by radial basis function (RBF)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002584_1999-01-0782-Figure10-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002584_1999-01-0782-Figure10-1.png", + "caption": "Figure 10. First prototype undergoing in house cornering fatigue test simulation.", + "texts": [ + " A weld fixture for accurate positioning of spokes during welding was fabricated using a combination of wood and metal parts. This approach allowed the construction of wheels with sufficient dimensional accuracy for standard corner- ing fatigue testing and very limited on-vehicle evaluations. TEST RESULTS \u2013 A range finding cornering fatigue test was carried out on a wheel prototype of this design to assess the validity of the approach and the analytical work. Ultimately, one prototype was fabricated and tested using a modified gap bed lathe to simulate a dynamic cornering fatigue test as shown in Fig 10. Testing to 18,000 cycles revealed cracking in the aluminum clamping blocks at the hub. It became obvious from limited testing that if a weight reduction could be had over a stamped steel wheel, further innovation and refinement at the hub spoke joint were going to be necessary. Based on what we had found during the testing, we decided to proceed toward a more production worthy design. DESIGN CRITIQUE \u2013 The decision to develop a different version of the wheel resulted from a review of the state of the design" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000827_6.iac-03-v.3.03-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000827_6.iac-03-v.3.03-Figure4-1.png", + "caption": "Figure 4. Examples of Shapes Studied for SLI (From Left to Right: Capsule, Lifting Body, Wing Body, SHARP Shape)", + "texts": [ + "0 3 OSP VOLUME AND FUNCTIONALITY Designing an OSP with the necessary volume to accommodate the required number of crew for the required time on-orbit, all within the desired weight target, is a considerable challenge. The OSP must have sufficient pressurized volume to accommodate the crew and sufficient unpressurized volume for the packaging of subsystems. The vehicle shape selected should not require a large-scale airframe to package subsystems, as this will drive up the weight. At the same time, a shape resulting in a small-scale vehicle that packages with a high density may not be able to achieve the desired planform loading and landing speed. As shown in Figure 4, the Orbital design team studied capsules, lifting bodies, wing bodies, and shapes employing sharp leading edge technologies and evaluated the pros and cons of each. Using the results of these trade studies, the outer mold line features were selected that were shown to best enable the OSP to meet its performance requirements. A new shape was synthesized from the lessons learned during the wide variety of system trade studies. This shape is called a Blended Lifting Body, or BLB. The BLB is very similar to a lifting body, but with many additional advantages" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002504_amm.137.95-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002504_amm.137.95-Figure2-1.png", + "caption": "Fig. 2: (a) Load time history, (b) sorted load time history - \"Duration Curve\" [2]", + "texts": [ + " R means capacity of component in safety or serviceability limit state, S means effect of actions in this states. The calculated probability of exceeding Pf is compared with target probability Pd, which has been appeared in some of standards (see Table 1). Actions in SBRA. Each action is expressed by a load duration curve and a bounded histogram with the design load values in SBRA method. Corresponding bounded histograms are developed on the basis of correlation between action F(t) and time of the service life span (see Fig. 2a). This \u201cLoad time history curve\u201d is rearranged to the \u201cLoad duration curve\u201d - LDC (see Fig. 2b). Than the \u201cLoad duration curve\u201d is transformed into relevant bounded histogram (see Fig. 3) [2], [3]. The load combination effect and moisture on the timber strength is expressed by the Madison curve (\u201cM-curve\u201d) \u2013 see Fig. 4. Material properties in SBRA. Material properties are expressed by bounded histograms obtained from laboratory tests. Tested material. Cement-splinter boards VELOX WS with a thickness of 35 mm and basis weight of 25 kg.m -2 (see fig. 5, 6). Fasteners. Dowels with a diameter of 6 mm made of steel, strength class S235, with tensile strength fu = 360 MPa were selected as fasteners for testing" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000289_bf00772952-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000289_bf00772952-Figure6-1.png", + "caption": "Fig. 6. Temperature distr ibut ion over the radius of a f ive-dimensional sphere for the instant of dimensionless t ime Fo = 0 .2 wi th di f ferent values of Blot criterion and r2/r 1 = 2.", + "texts": [], + "surrounding_texts": [ + "34G finite number of terms of the series in the case of six terms in the series. With the passage of time this error decreases rapidly. For example, the amplitude of the sixth term falls in proportional to exp[-(37.5)2Fo], i.e. with Fo = 0.01 the error is reduced by more than a factor of 106. These temperature distributions are observed in annular and rod models for a five-dimensional sphere with n, = 3 with the same ratios of r2/r 1, and Bi and Fo numbers. Similar temperature distributions to these are also noted in annular and rod models with other values of n, and with values of r2/r l, Bi, and Fo corresponding to them (from the condition for conformity of rod model profiles). Thus with n, = 2 we have r2/r 1 = 1.58, Bi --- 0.705, Fo = 2.01 or with n, = 1 we have r2/r 1 = 2; Bi = 0.414; Fo = 5.828. In the case of n. = 1 and n. = 3 it is possible to calculate an accurate value for temperatures, for example with n, = 3 Bi = 1, Fo = 0.2 the temperature at the inner surface will be 0 = 0.268; with n, = 1 Bi = 0.414, Fo = 0.2.5.828 = 1.1656, 0 = 0.244. The approximate temperature value with n, = 2 may be estimated by using precise calculations with n, = 3 or n. = l, and also by interpolation: / ' I , - - 1 On.=z = On.=, + (On.=3 - On.=, )g__- 1 = 0 ,256. Thus, by using annular specimens in thermal fatigue tests with thermally insulated parabolic end surfaces it is possible very effectively to vary simultaneously the thermal and mechanical stresses operating in a specimen. It is comparatively easy and quite accurate to calculate the temperature fieId in a specimen with heat exchange at the inner surface. . 2. . 4. 5. 6. 7. 8. LITERATURE CITED M. M. Khrushchov, Fatigue in Babbits [in Russian], Izd. Akad. Nauk SSSR, Moscow (1943). G. N. Tret'yachenko and V. G. Barilo, Inventor's Cert. USSR 1539590 G 01 N 3/60, \"Thermomechanical fatigue test specimen,\" Otkrytiya. Izobret., No. 4 (1990). G. N. Tret'yachenko and V. G. Barilo, Inventor's Cert. USSR 1381372 G 01 N 3/60, \"Thermal fatigue test specimen,\" Otkrytiya. Izobret., No. 10 (1988). V. G, Barilo, \"Use of annular specimens of wedge-shaped section for studying the thermal fatigue of materials,\" Probl. Prochn., No. 2, 60-64 (1989). A. D. Kovalenko, Circular Plates of Variable Thickness [in Russian], Fizmatgiz, Moscow (1959). A. D, Kovalenko, Plates and Shells in Turbomachine Rotors [in Russian], Izd. Akad. Nauk UkrSSR, Kiev (1955). N. N. Malinin, \"Design of circular and annular symmetrically loaded plates of variable thickness,\" in: Strength Analysis, Brittleness, and Creep of Engineering Structural Elements [in Russian], Mashgiz, Moscow (1953). G. N. Tret'yachenko and V. G. Barilo, Inventor's Cert. USSR 1620916 5G 01 N 3/60, \"Thermal fatigue specimen for testing materials,\" Otkrytiya. Izobret., No. 2 (1992)." + ] + }, + { + "image_filename": "designv6_24_0001805_cem.2013.6617125-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001805_cem.2013.6617125-Figure1-1.png", + "caption": "Fig. 1. Antenna model view in the simulation program.", + "texts": [], + "surrounding_texts": [ + "The procedural steps to build up a simulation scenario are the modeling of the antenna, of the car body and of the environment in terms of an asphalt ground." + ] + }, + { + "image_filename": "designv6_24_0001826_870305-Figure18-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001826_870305-Figure18-1.png", + "caption": "Fig. 18", + "texts": [], + "surrounding_texts": [ + "870305\n9\nTrends: See chart Also higher degree of integration: Formation of large pre-assembled units, consisting of instrument panel. heater or air conditioner, steering, pedal~, fuse-and-relay board.\nHaterial: Trends:\nPP, ashtray: FS 31 (phenoplast, saw-dust-filled) - Multi-color injection-molded - Textile decor - Softer surface", + "10\no Heater and air conditiorer\n870305\nHaterials:\nTrends:\n- Fousing parts: PP TF - Blower wheel: PON - Thin-layer technique for - Flaps hard/soft\n- Flilps: - Controls:\nhousing parts\nNetal/PU foam ABS", + "870305\n11\nHaterials:\nTrends:\n- Housing parts: PP TF - Blower wheel: POM - Thin-layer technique for - Flaps hard/soft\n- Flaps: - Controls:\nhousing parts\nHetal/PU foam ABS" + ] + }, + { + "image_filename": "designv6_24_0002075_jrfid.2020.3010196-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002075_jrfid.2020.3010196-Figure4-1.png", + "caption": "Fig. 4: (a) Simplified spiral microstrip resonator. (b) Network model. [22]", + "texts": [ + " To the authors\u2019 knowledge, most of the time, the results from simulation using this model are not showcase. Furthermore, there is no transparent reporting of how the computation of these values is done. Taking into account the few reportings of the simulated data using this model, the results of the lumped-element model are satisfactory. This method is rather lengthy as it requires much computation and is strongly nonintuitive and, thus, it is not of great help when trying to design gap-coupled microstrip resonators. In [22] the simplified spiral microstrip resonator (SSMR), portrayed in Figure 4(a), is modeled as a two-port network using ABCD parameters. This model uses a coupled line with two terminals linked by a transmission line and another two terminals open-circuited. The transmission line has a characteristic impedance of Zt and an electrical length of \u03b8t. The even and odd mode characteristic impedance of the coupled line is Ze and Zo, with an electrical length of \u03b8e and \u03b8o. Assuming that the SSMR is excited at point A, the network model of SSMR may be derived. The results for the fundamental and first spurious resonant frequencies of the SSMR show a good agreement between the full-wave simulated and the network model" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002611_isocc47750.2019.9027698-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002611_isocc47750.2019.9027698-Figure2-1.png", + "caption": "Fig. 2. Magnetic drive unit with six electromagnets.", + "texts": [ + " Horizontal motion was realized by making the horizontal magnetic flux gradient nonzero at the levitation point. Therefore, based on the discussion above, the levitation was determined by two factors: the produced magnetic field and the magnetization of the magnetized object. B. Magnetic Field in the Working Space The controlled external field, in this study, was produced by six identical electromagnets which were evenly distributed on a disc (see Fig. 1). According to the Biot\u2013Savart\u2019s law, the magnetic field produced by six multilayer finite length solenoid with iron core (see Fig. 2) at a random point in free space is [25]: Bz (x, y, z) = 6\u2211 i=1 Bzi(x, y, z) (6) where i indicates the ith electromagnet, i = 1, 2, . . . , 6, Bi(x, y, z) is the magnetic flux density of the ith electromagnet at point P (x, y, z). Bzi(x, y, z) is the z-component magnetic flux density of the ith electromagnet at point P (x, y, z): B(x, y, z) = \u2212\u03b2 \u03bc0\u03c3Ii 4\u03c0 ro u t\u222b ri n L/2\u222b \u2212L/2 2\u03c0\u222b 0 r[{y \u2212 yi \u2212 rsin\u03c6}sin\u03c6 + {x \u2212 xi \u2212 rcos\u03c6}cos\u03c6] |Ri |3 d\u03c6dzdr (7) where \u03b2 is the magnetic flux density factor of the soft iron core effect in enhancement of the magnetic field, \u03bc0 is the relative 1083-4435 (c) 2015 IEEE" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003233_cp.2014.0319-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003233_cp.2014.0319-Figure1-1.png", + "caption": "Fig. 1: Topology of investigated model.", + "texts": [ + " During the operation, the LCF magnets can be demagnetized or remagnetized in order to However, different excitation sources will always influence each other during operation, such as armature reaction in PMSM [11], and combination of magnets in multi-layer PMassisted synchronous machine [12]. Therefore, the cross coupling effect may also exist in HMMM and needs to be investigated in detail. In this paper, a 24-slot/4-pole HMMM model is built according to [9]. Based on that, the cross coupling effects between two types of magnets considering armature reaction will be investigated, including the machine performances such as open-circuit back EMF, back EMF after load and magnetization capability. Fig. 1 shows the topology of analysis model. The NdFeB magnet is placed near rotor shaft and parallel magnetized, while the LCF magnets are placed at the side of NdFeB magnet with circumferential magnetization direction. Notably, the LCF magnets are designed as trapezoidal shape, which helps to partially demagnetize or remagnetize the LCF magnet in order to achieve different magnetization levels. Basic geometric parameters of the analysis model are shown in Table 1. Fig. 2 shows the demagnetization curves of two permanent magnets which are the same as the original paper about HMMM [9]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003554_jrfid.2019.2926194-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003554_jrfid.2019.2926194-Figure4-1.png", + "caption": "Figure 4. (a) Dimensions of the SRR and slotted patch antenna (b) Fabricated prototypes of the wearable RFID reader antennas.", + "texts": [ + "org/publications_standards/publications/rights/index.html for more information. sufficiently small size and flexibility are a priority. Moreover, the wearable patch antenna we presented in [19], was considered a good candidate because it possesses a ground plane, which acts as an insulator between the patch of the antenna and the human body. On the other hand, the single layered SRR presented in this work, is easier to integrate with a glove. Both of the wearable RFID reader antennas are matched to 50 Ohm input impedance. Fig. 4 shows the geometry of the slotted patch and SRR antenna and Table 3 lists the dimensions of both antennas. ANSYS HFSS v.19.1 was used in the optimization of the antennas. The 7-layer planar human hand model discussed in section III was used to tune the antennas at a resonance frequency of 866 MHz with realized gain as high as possible. The antennas are fabricated using conductive nickel and copper plated electro-textile (sheet resistance 0.16 Ohm/Sq and thickness 0.17 mm) on EPDM foam material having the relative permittivity of 1" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003553_iciap.2003.1234118-Figure3.1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003553_iciap.2003.1234118-Figure3.1-1.png", + "caption": "Figure 3.1: Scheme of Peering Behavior of Praying Mantis and the implementation of the Miniature Mantis Head Camera Platform, which utilizes peering behavior for distance estimation", + "texts": [], + "surrounding_texts": [ + "A miniature video camera was mounted on a specially designed micro-translation platform, which provides precise periodic side-to-side peering movements of the camera with constant speed. When an electromotor of the platform is activated, the camera translates in the direction that is parallel to the image plane. This behavior simulates the peering behavior of the praying mantis. The video output signal of the camera is connected to the miniature wireless video RF transmitter, which broadcasts video signal remotely, thus enabling autonomic usage of the device on mobile robot. Both camera and transmitter are operated from single 9v battery. The total size of the platform with camera and transmitter is [10cmx5cmx2cm]. The video signal is then received by an RF video receiver that is connected to the PCI frame grabber located inside a Dual Pentium III workstation, which performs the image processing of all the incoming frames. In addition, based on the incoming image analysis, the workstation could send action commands back to the remote robot, supporting the peering platform. Varying the target distance and peering velocity parameters, performance of the system was measured. Targets were placed at various distances in front of the camera: 5, 6, 7, 8, 9, and 10 cm. Peering velocities of 1.5cm/sec and 2cm/sec were used. [For Mantis Religiosa individuals 50 to 70 mm in size, peering amplitudes are approximately 2 to 10 mm and peering velocities approximately 6 to 18 mm s-1.]" + ] + }, + { + "image_filename": "designv6_24_0001305_024-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001305_024-Figure1-1.png", + "caption": "Figure 1. Elements of the electrostatic ultrasonic motor.", + "texts": [ + " On the contrary, piezoelectric micromotors display better theoretical characteristics (high torque, low speed), but the relatively small number of papers concerning this subject reveals the difficulties in obtaining good quality piezoelectric thick layers [3]. With the aim of overcoming the limitations inherent in both these types of micromotors, we describe a new type of micromotor based on ultrasonic electrostatic actuation. The stator is composed of four elements: (i) excitation electrodes; (ii) a membrane; (iii) a spacer to create a gap between the membrane and the electrodes; (iv) a rotor (figure 1). The electrodes are used to generate a flexural travelling wave on the membrane by means of electrostatic forces. This wave results from the superposition of two stationary waves with a \u03c0/2 phase difference in time and position. Due to the propagation of the wave, an elliptical movement takes place on the surface. A rotor pressed on it is propelled by this elliptical surface displacement through frictional forces. Theoretically, this micromotor principle presents the following advantages. \u2020 Tel: (33) 76 88 55 30" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000800_941748-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000800_941748-Figure7-1.png", + "caption": "FIGURE 7. A-Frame Optimization Project.", + "texts": [ + " The students are introduced to shape optimization for weight reduction in a simple structural problem such as that shown in Figure 6 [5]. For their final project, the students are given a more open-ended design and optimization project. In the most recent semester, students were asked to design and optimize an A-arm from an automobile suspension system. The objective was to minimize the weight of the arm given the overall dimensions and loading conditions. An picture of one of the submitted solutions is shown in Figure 7. OBSTACLES One of the major problems associated with using advanced software in course work is the additional time involved with software training. Companies typically spend large amounts of time and money for such training, and it is difficult to squeeze enough instruction into the cumculum to enable students to effectively make use of the software. A related problem is that of making software documentation available to the students. The user manuals for commercial CAE programs are typically quite vast, and it is difficult to assemble the critical information and make it readily available to the students" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000011_095765005x31108-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000011_095765005x31108-Figure4-1.png", + "caption": "Fig. 4 Experimental test rig: (a) schematic diagram of performance test set-up and (b) view of test section", + "texts": [ + " Power and Energy the impeller; (b) static pressure distribution (blade loading) around the impeller blade surface; (c) relative velocity vector distributions in the impeller; (d) relative velocity vectors at the impeller leading edge; (e) relative velocity vectors at the impeller trailing edge; (f) absolute velocity vector distributions in the volute casing; and (g) overall static pressure distribution in a centrifugal blood pump Proc. IMechE Vol. 219 Part A: J. Power and Energy JPE126 # IMechE 2005 Figure 3(g) illustrates the overall static pressure distribution in a centrifugal blood pump. The available kinetic energy at impeller exit has been uniformly recovered in the volute casing. 3 IN VITRO HYDRAULIC PERFORMANCE ANALYSIS A schematic diagram and photograph of the experimental set-up is presented in Fig. 4. Basic in vitro experimental measurements including pump static pressure differential, discharge flow rate, impeller rotational speed (r/min), and shaft torque have been taken using water in the present study. The wall static pressure rise across the Fig. 3 Continued JPE126 # IMechE 2005 Proc. IMechE Vol. 219 Part A: J. Power and Energy inlet duct and the volute discharge was measured with a differential-pressure transducer. The pump flow rate was measured with a positive displacement flowmeter of the mock circulatory system" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002102_iecon43393.2020.9255013-Figure13-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002102_iecon43393.2020.9255013-Figure13-1.png", + "caption": "Fig. 13. (a) Mismatch between 180\u00b0 PMs; (b) Spiral angle \u03b1; (c) Single piece 45\u00b0 helical magnet; (d) Single-helix MLS with 45\u00b0 helical PMs.", + "texts": [ + " 12(d) depicts the schematic drawing of single-helix MLS with 45\u00b0 helical PMs. The mismatch between PMs is zero because the sectional drawing of each PM is a rectangle, rather than parallelogram shown in [15]-[16]. In [16], another shape of permanent magnet with parallelogram section drawing is utilized but the mismatch between PMs appeared. The above-mentioned design method is suitable for both the rotor magnets and translator magnets, which is easy to implement especially in mass production. Fig. 13 depicts the assembly of PMs for building the single-helix translator, which is also suitable for the rotor. The iron yokes can be manufactured integrally with the iron core, then PM blocks are inserted into the spaces and fixed by strong adhesives. 2811 Authorized licensed use limited to: Sungkyunkwan University. Downloaded on December 18,2020 at 16:25:50 UTC from IEEE Xplore. Restrictions apply. V. CONCLUSIONS In this paper, a novel single-helix MLS is proposed to provide an additional choice for the application of MLS in WEC system" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000223_1.4030653-Figure18-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000223_1.4030653-Figure18-1.png", + "caption": "Fig. 18 Fixed-guided bending mode test (a) setup\u2014flexion (right)/extension (left) and (b) resultant\u2014moment-assist at knee joint", + "texts": [ + " Figure 17(a) reflects this significant stiffness-transition that can protect knee against hyperextension. Figure 17(b) shows the positive assistivetorque with increasing joint-angle that aids the flexion motion of saw-bones model operated with/without the knee-brace (now solely operating in beam-bending mode A). We have attempted to carefully eliminate sources of stick-slip friction in the capture of the torque-assist/deflection data. The stiffness at fixed-guided bending mode is also tested as shown in Fig. 18. Red arrow (Fig. 18(a)) depicts the direction of resistive stiffness and femur part is pushed to the direction of hyperextension motion, which can cause serious knee injury. Kutzner et al. [1] had established that during activities of daily living, the peak moments as percentage of body weight can range between 0.44% (extension moment) and 3.16% (flexion moment). For an 80 kg subject this can translate to moments ranging from 3.4 Nm to 24 Nm. Building on this, we had examined a design specification of 10% reduction of knee joint moment ( 0.34 Nm to 2.4 Nm) for our brace. Figure 18(b) shows estimated and measured moment at knee joint. The difference between estimated and measured value is mainly caused by accumulated tolerance of prototype, resolution of optical tracking system, friction and deformation of 3D printed parts. We also note that we have only tested the initial prototype for a small range up to 0.6 Nm (due to force/torque transducer resolution and strength limitations of 3D-printed prototypes). Journal of Mechanisms and Robotics NOVEMBER 2015, Vol. 7 / 041024-9 Downloaded From: http://mechanismsrobotics" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000711_pierl09120105-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000711_pierl09120105-Figure8-1.png", + "caption": "Figure 8. Measured farfield radiation pattern of the inkjet printed monopole antenna with different conductor thickness in (a) XY plane, (b) XZ plane, and (c) YZ plane.", + "texts": [ + " The peak total efficiency clearly increases with the thickness of the printed layer (Figure 6). After five printings producing silver conductor thickness of 5.5\u00b5m the total efficiency is very close to the efficiency of the reference antenna. The gain value on the other hand with the printed silver layer of thickness of 1.5\u00b5m showed minimum gain of 0.38 dB. Thickness equal to 5.5\u00b5m showed the maximum gain of 1.96 dB where the maximum gain for the reference antenna was 2.2 dB. Additionally, when the radiation patterns (Figure 8) measured using Satimo showed actually no effect of the conductor thickness on the omni directional performance of the antenna, the final conclusion is that the L-shaped monopole antenna presented here can be fabricated by the printed electronics methods proposed without loosing its performance. The inkjet printed L-shaped monopole antennas with different conductor thicknesses were designed, manufactured and measured. The results show that the printed electronics methods using nano silver inks with high enough conductivity can be used to produce antennas with very competitive performance" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000715_062035-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000715_062035-Figure1-1.png", + "caption": "Figure 1. ANSYS (Meshing)", + "texts": [ + " This is done so as to provide ample number of iterations and at the same time, to provide a 5 IOP Conf. Series: Materials Science and Engineering 263 (2017) 062035 doi:10.1088/1757-899X/263/6/062035 congruence point in a relatively quicker time [10-12]. Also, the radiosity controls that measure the relation between the radiation convergence and the surface parameters has been kept on with a convergence value of 0.0001 W/m2. Smooth contours have been set to provide details about the variations in the input and output results. The meshed model of the designed structure is shown Figure 1. The solver used here was of gauss-seidel iterative type. These are the FEM equations needed and used by the solver while analyzing the data given and these are the main equations used for determining the results. The resources for this data were taken from the help section in the project window main screen. Biot-Savart Law for finding the magnetic vector potential Gauss Law for determining flux density Faraday\u2019s law for calculating the electric field intensity. Ampere\u2019s Law for current density (not compulsory)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003028_asemd49065.2020.9276314-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003028_asemd49065.2020.9276314-Figure2-1.png", + "caption": "Figure 2. No-load magnetic dense cloud for rotor and stator.", + "texts": [ + " And permanent magnet exciting circuit includes permanent magnet, rotor claw, air-gap, stator tooth, stator yoke, stator tooth, air gap, rotor claw and PM. This structure eliminates spatial interaction between electric load and magnetic load caused by slotting rotor. III. RESULTS OF 3D FEM In this paper, electromagnetic performance of a 1.5kW hybrid exciting claw-pole machine for vehicle is simulated by 3-D FEM. No-load magnetic field and output characteristics in power generation and electric state are calculated respectively. Main machine design parameters are shown in Table I. In Fig. 2 (a) and (b), magnetic density cloud of rotor claw and stator core under no-load are shown. It can be seen from that neither of them has magnetic density saturation. Output current and voltage waveforms after rectifying are shown in Fig. 3 (a) and (b). The effective values of output voltage and current are 13.6V and 100A, respectively. Phase current of machine under load operation is given in Fig. 4 (a). In design of 1.5kW hybrid exciting claw-pole machine, the This work was supported by Program for LNIRT (LT2017003 LZGD2017042)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003464_ijvd.2017.082579-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003464_ijvd.2017.082579-Figure1-1.png", + "caption": "Figure 1 Welded beam structure", + "texts": [ + " The Nelder\u2013Mead simplex search method is proposed by Nelder and Mead (1965), which is a local search method designed for unconstrained optimisation without using gradient information. The operations of this method rescale the simplex based on the local behaviour of the function by using four basic procedures: reflection, expansion, contraction and shrinkage. Through these procedures, the simplex can successively improve itself and zero in on the optimum. A welded beam design optimisation problem, which is often used as a benchmark for testing different optimisation methods, is used to illustrate the implementation procedure of the HCSSNM for solving optimisation problems. Figure 1 shows design variables and structure of the welded beam. The objective is to find the minimum fabricating cost of the welded beam subject to constraints on shear stress (t), bending stress (s), buckling load (Pc) and end deflection (d). The beam has a length of 14 in. and P = 6000 lb force is applied at the end of the beam (Siddall, 1972; Rastgell and Philips, 1976; Coello and Montes, 2002). The design variables are thickness of the weld h(x1), length of the weld l(x2), depth of the beam t(x3), and width of the beam b(x4)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000567_tmag.2017.2703844-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000567_tmag.2017.2703844-Figure6-1.png", + "caption": "Fig. 6. PMLSs with different number of pole pairs p. (a) p=1. (b) p=4. (c) p=12. (d) p=22.", + "texts": [ + " Table \u2160 Parameters Common to All Designs Quantity Value Magnet thickness on screw and nut (mm) 3 Magnet coercivity (kA/m) 838 Magnet romance flux density (T) 1.1 Airgap length (mm) 1 Iron core thickness (mm) 3 Axial length of nut (mm) 40 A. Effect of the Number of Pole Pairs The effect of the number of pole pairs on thrust force was analyzed using 3-D FE model in [7], and the conclusion that the thrust force remains essentially constant as the number of pole pairs increases was drawn. According to the conclusion in [7], 2-D axis-symmetric FE model is applicable no matter how many the pole pairs are. Fig. 6 presents the 3-D FE models with the numbers of pole pairs being the only difference, the axial width of the four models is equal to 5 mm so that the same 2-D FE model shown in Fig. 4 is shared. The nut travels axially by a distance of 5 mm (the magnet axial width) while the screw is stationary. The simulation results of models in Fig. 5 and Fig. 6 are shown in Fig. 7. It can be seen that the characteristic of thrust force varies with the number of pole pairs, not in agreement with the conclusion in [7]. It\u2019s worth noticing that the max lead angle of PMLS in [7] is 14.15 deg, however, the lead angle greater than 14.15 deg is considered in this part. So the conclusion in [7] is inappropriate when lead angle increases up to a certain value. The flux leakage between adjacent poles is accounted for the variation of thrust force because the flux leakage becomes larger when pole-pair count increases" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002475_iet-map.2015.0732-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002475_iet-map.2015.0732-Figure2-1.png", + "caption": "Fig. 2 Configurations of the ten-layer FR-4 PCB", + "texts": [ + " The current form distinguishes the letter by including details of design limitation, analysis in the antenna\u2019s unique radiation pattern, parametric studies, and measurement data of radiation pattern from a millimetre-wave chamber. In the following, Section 2 describes the proposed ZOR antenna layouts and parametric studies conducted by full-wave simulations. In Section 3, measurement results are provided and compared with the simulation results to secure the validity of the design. We first describe the antenna design restrictions originating from PCB fabrication rules. Fig. 2a shows the stack-up diagram of a commercial ten-layer FR-4 PCB used in mobile communication handsets. The designed antenna aims at fabricating on the top three layers of the given PCB to avoid interruption by the thru via located below the fourth layer. Indeed, the height between the antenna ground and the aperture is limited to 180 \u03bcm (i.e. 0.05l at 80 GHz). Furthermore, the antenna geometry is required to comply with PCB fabrication rules from the manufacturer. As depicted in Fig. 2b, the line width of the copper should be more than 100 \u03bcm, 1 a Top view b Side view and the gap between the copper should be more than 60 \u03bcm for reliable fabrication. These values consider \u00b110% tolerance during the fabrication process. The stack via has the specific diameters of 100 and 250 \u03bcm for the via post and land, respectively. For PCB-embedded antennas for low-frequency applications, via lands are often not considered in the antenna analysis, as they are much smaller than the operation wavelength" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001594_9783527646982.ch6-Figure6.8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001594_9783527646982.ch6-Figure6.8-1.png", + "caption": "Figure 6.8 (a) Schematic illustration of the horseshoe - patterned metal conductors encapsulated in an elastomeric substrate; (b) fi nite element models for fi ne pitch (top) and coarse pitch (down).", + "texts": [ + " It is important to highlight that this comparative study was done for a constant substrate thickness. However, in practice, it is possible to fabricate and handle thinner substrates with stiffer material properties such as polyurethane and, therefore, the induced plastic strain can be reduced. 6.2.3 Induced Mechanical Interaction on Multitracks The effect of the pitch on the mechanical behavior of the parallel aligned stretchable interconnects is investigated through numerical modeling [4, 5] . Figure 6.8 a illustrates the horseshoe - patterned metal conductors encapsulated in an elastomeric substrate. The metal conductors are completely (above and below) encapsulated and they reside in the center of the polymer substrate along the thickness direction. For visual clarifi cation, only three parallel - aligned interconnects are shown in the fi gure. The angle ( \u03b8 ) of each meander of the patterned metal conductor is 30 \u00b0 . The width ( w Cu ), thickness ( t Cu ), and radius ( r Cu ) of the metal track are 100 \u03bc m, 18 \u03bc m, and 750 \u03bc m, respectively. The substrate is a block with W sub = 20 mm wide and T sub = 1 mm thick. The length of the substrate depends on the number of repeating meander units. A uniaxial elongation \u201c u \u201d is applied to the substrate at the one end and the other end is assumed as a symmetrical plane, which corresponds to the experimental conditions. Figure 6.8 a (top and bottom) show the fi nite element meshes with fi ne and coarse pitch in the relaxed state (i.e., nonstretched). The shown part corresponds to the middle block of the real horseshoe - patterned stretchable sample in the experiment. A displacement ( u ) of up to 50% elongation is applied at one end surface in the numerical models, and the simulated maximum equivalent plastic strain in the metal is used for further analysis. Figure 6.9 shows the maximum plastic strain at 50% elongation as a function of line - to - line pitch" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001347_ijabe.v10i6.3142-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001347_ijabe.v10i6.3142-Figure3-1.png", + "caption": "Figure 3 Position of seed clearing device", + "texts": [ + " The seed filling room is filled with seeds and then seeds accumulate on the surface of sucking plate. Seeds are sucked on the holes of sucking plate under the effect of vacuum of sucking chamber shell. Because of the low degree of sphericity and surface smoothness of rice seeds, it is difficult to precisely control the number of sucked seeds. These reasons will lead to multiple seeds sucked in a hole. In order to further improve the phenomenon, the seed cleaning device is installed in the seed metering device. The seed cleaning device is shown in Figure 3. The two clearing fingers are perpendicular to surface of sucking plate, and length of each clearing finger is 40 mm. There is a gap of 0.1 mm between the clearing fingers and the sucking plate. This gap can prevent mechanical interference when the seed plate rotates. The gap is much smaller than the size of the seed, and seeds will not leak out. The clearing fingers are located on the outer side of the 2 suction holes, 4 mm away from the center of the hole. If the distance between hole center and clearing fingers is too close, it will cause excessive clearing" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001087_j.scriptamat.2006.03.027-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001087_j.scriptamat.2006.03.027-Figure1-1.png", + "caption": "Fig. 1. Principle of dynamic digital speckle patterns interferometry and data processing method.", + "texts": [ + " Half of these samples were annealed at 723 K for 4 h and then cooled naturally to room temperature. The rest were solution treated at 773 K for 3 h and then quenched in water. The solution-treated samples were divided into two groups with one group tested within 5 min in tension and the other tested after different natural aging durations, 3, 4.6, 6.5, 7, 10 and 24 h. In our tests, the DSPI technique, a real-time two-dimensional observation method, was used to measure the geometric and kinetic aspects of the PLC deformation bands. Fig. 1 shows the principle of DSPI and the data processing method. A vibration-resistant testing machine was specially designed for optical interferometry observations during tensile tests. A sample was clamped at two ends with the lower end fixed and the upper one stretched along the Xdirection at a constant speed. The range of the applied strain rates was from 10 5 to 5 \u00b7 10 3 s 1. A charge-coupled device (CCD) camera was placed in front to collect the interference speckle patterns formed on the specimen surface" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002388_2015-01-0090-Figure19-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002388_2015-01-0090-Figure19-1.png", + "caption": "Figure 19. Strain distribution (Von-Mises) in case of a pair of rear wheels at", + "texts": [ + " case of Rear-Left wheel on ramp (Lateral direction) case of Rear-Left wheel on ramp (Longitudinal direction) The purpose of this test focused on the bending behavior of the side rails. A pair of wheels was lifted simultaneously by driving up either front wheels or rear wheels to ramps aligned parallel to each other. The vertical distances between wheel centers and level road were 280 mm for both cases. Figure 18 shows the dynamic responses obtained from the test while ramping up the rear wheels. A contour plot of equivalent strains on the truck frame is illustrated in Figure 19. The lateral and longitudinal strains for front wheels ramping and rear wheel ramping obtained from the simulations and experiments are shown in Figure 20, Figure 21, Figure 22, Figure 23. In both cases, experimental results had the same trend except the locations 1 and 2. For the simulation results, strain magnitudes were slightly higher in case of rear wheels on the ramps compared to the other case. It is possible that the shear effect of shell element formulation is neglected in the simulation as a result of transferring shear effect to bending effect instead" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000781_j.jmatprotec.2016.09.016-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000781_j.jmatprotec.2016.09.016-Figure2-1.png", + "caption": "Fig. 2. The locations of all test samples.", + "texts": [ + " The balance of alloy powders was composed of Mn owder with a constant concentration of 2.0 wt.% and iron powder. FAB-SAW experiments were conducted using a Lincoln DEALARC\u00ae DC-1500 welding machine. The welding parameters ere set as follows: a welding current of 1050 A, a welding voltge of 34 V, a wire feed speed of 170 cm/min, a welding speed of 0 cm/min, a wire extension of 35 mm and a welding heat input of 07.1 kJ/cm. Samples used for metallurgical and mechanical test were xtracted and machined from each weld sample using an electric ischarge machine (EDM), as shown in Fig. 2. The sample for chemcal composition analysis was extracted from the center region of eld metal. Two samples for metallurgical testing including one ample for optical microstructure (OM) and another sample for ransmission electron microscopy (TEM) were obtained. Two lonitudinal samples for tensile test and three samples for Charpy V-notch impact test were prepared. The locations and sizes of samples for mechanical properties test are shown in Fig. 3. The chemical compositions of the weld metals were investigated by using an X-ray fluorescence spectrometer, a CS600 carbon-sulfur analyzer for C and S, a TCH600 oxygen-nitrogen analyzer for O content and an ICP-MS X Series II for Ce content" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002367_iros.2009.5354127-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002367_iros.2009.5354127-Figure4-1.png", + "caption": "Fig. 4 When the Ym axis motor is rotated", + "texts": [ + " In this case, the output axis extending from the anterior end of the Xm axis motor passes through the slider and is joined to the rotating ball. According to the structure stated above, there is no jiggling in the Xm axis motor between the slider and the hollow part of the motor bearing, so the rotation is regulated around the axial direction of this part itself (see Fig. 3). Moreover, due to the slider there can be no reciprocation along the motor bearing, so among the other 2 orthogonal axes the Xm axis motor can turn around the Y0 axis (see Fig.4). With this type of motor bearing, the ends are fixed to the rotation axes, and there is axial support in such a way that rotation is possible at the support-end located on the Z axis of the clamp. There are two bearings installed internally in the support-end, and by aligning the two bearings in the axial direction of the rotation axis, even when an external force is applied to tilt the rotation axis the two bearings act together to support the rotation axis and prevent it from tilting. According to the structure stated above, among the other 2 orthogonal axes, the motor bearing, i" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002605_973233-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002605_973233-Figure7-1.png", + "caption": "Fig. 7. Force Equilibrium", + "texts": [ + " To facilitate calculations, insertion is modeled as movement of the clip from 0 = ei to 0 = 0 in arbitrary increments, for each of which the forces required to maintain equilibrium are calculated. The maximum deflection, occurring when the valve body moves down and touches clip at point B (see Fig. 6), is and W rc = r4 + - 2 (1 5) where w is the width of the clip. The clip deflection at any angle 8 can be calculated as where The contact force p now can be obtained from Eq. (8) as To compute the insertion force all the forces applied to the valve body are plotted in Fig. 7, where F,, Nand p are respectively the insertion force, normal force and coefficient of friction. The equilibrium of those forces at vertical and horizontal coordinates leads to and where Bi is the insertion angle as shown in Fig. 7. By substituting Eq. (20) into Eq. (19), the insertion force can be solved as and The maximum insertion force is determined by comparing F, at each increment of 0.1' in angle 8, which is the angle between points A and B (see Fig. 6). The maximum removal force can be obtained in a similar manner. To avoid confusion the insertion and removal forces used in the following mean the maximum insertion and removal forces, respectively. OPTIMIZATION PROBLEM Outimization Parameters - All of the 14 geometrical parameters shown in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002095_cobep.2013.6785220-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002095_cobep.2013.6785220-Figure4-1.png", + "caption": "Fig. 4. Integrating surface in each stator tooth.", + "texts": [ + " Considering the initial position for the rotor as 0=r\u03b8 and by setting 1A as the base field current while all stator windings current is zero (no-load condition), the magneto-static FE solution of flux density has to be sampled in each stator tooth. The current base of 1A was chosen just for convenience. Actually, any other current base value ( )basei could be assumed, however for arbitrary excitation currents, the ratio between both \u239f\u239f \u23a0 \u239e \u239c\u239c \u239d \u239b base f i i has to be considered. Integrating the flux density B over the area S, the flux is obtained: \u222b\u222b=\u03a6 S j dSB. (7) where j is the stator tooth index and S is the surface which is concentric to the rotor surface and covers each tooth span. Fig. 4 shows the integrating surface S. Assuming that all flux lines quantities pass through stator tooth (no flux leakage), the equation (7) was used to calculate the flux for each j tooth, being j=1,2,\u2026,36 for the target machine. Herein, the flux was captured in all stator tooth for each rotor position until complete one pole pair, and storage as ),( rR j \u03b8\u03a6 . In this case (4 pole machine), one pole pair represents a quarter of machine, which means that the flux density has to be sampled for 90 mechanical degrees in rotor positions" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000782_ssp.198.301-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000782_ssp.198.301-Figure2-1.png", + "caption": "Fig. 2. Draft of the measurement stand", + "texts": [], + "surrounding_texts": [ + "The draft of the impact cylinder is presented in Fig. 1. The movement of the impact piston takes places when the product of P1 and S1 is higher than P2 and S2. After a slight shift of the piston, the area affected by pressure P1 significantly grows. This effect leads to an abrupt acceleration of the movement of the piston caused by the activity of compressed air accumulated in chamber 1. In order for the air flow obstruction to be avoided, the opening in chamber 2 needs to be of a big enough size, which in practice is realised with the use of a quick release valve. The volume of chamber 1, that plays a role of a pressure accumulator, is of basic importance for the level of energy obtained." + ] + }, + { + "image_filename": "designv6_24_0001626_tia.2017.2766585-Figure10-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001626_tia.2017.2766585-Figure10-1.png", + "caption": "Fig. 10. FEM of (a) original IM and (b) optimized IM.", + "texts": [ + "57% torques ripples are due to the supply voltage time harmonics, only. Hence, the complete design methodology is highly effective in reducing the output torque ripple with minor efficiency improvement. The targeted torque ripples are considered first time primarily due to the interaction of time harmonics and space harmonics of slot effect. The performance of optimized design of IM has been validated and compared with the performance of original IM using FEM. The FEM have been developed using the parameters of original IM and optimized IM and shown in Fig. 10. Quarter symmetry has been considered while developing FEM of original IM. But the number of slots per phase per pole is not an integer value in the optimized model, so complete cross section is required for the FEM analysis as shown in Fig. 10(b). The performance of optimized IM has been compared with the original IM using FEM, in the terms of torque, torque ripple, and stator currents as shown in Figs. 11 and 12. Fig. 11(a) shows the rated torque response and ripples in the torque are as high as up to 40% as shown in Fig. 11(b). Stator currents are not purely sinusoidal due to the presence of space harmonics can be observed in Fig. 11(c). Fig. 12(a) shows the torque response, and in Fig. 12(b), ripples in the torque are reduced to 10% in optimized design as compared to 40% in original design of IMs shown in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002448_2011-01-0862-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002448_2011-01-0862-Figure8-1.png", + "caption": "Figure 8. Final concept design using wedge combustion chamber and rocker actuated valves with inlet entering via the cylinder head fire face.", + "texts": [ + " The resulting valvetrain layout is shown in Figure 7b, where it can be seen that an overall height saving of \u223c20 mm was achieved over the direct acting layout with graded buckets. To keep the intake assembly within the target package space, it was decided to locate the inlet plenum alongside the cylinders (beneath the cylinder head overhang). It was a logical step to create an inlet gas path that entered the cylinder head on the fire face rather than a machined face on the side (a manifold face on the side of the cylinder head and intake pipe bend radii would significantly increase the width of the RE engine), as shown in Figure 8. This layout also gave some additional benefits for the cylinder head; a minimal requirement for machining on the inlet side of the cylinder head and no requirement for a side core on the casting. The cylinder head casting had just three external cores, plus coolant jacket, two inlet ports and a conjoined exhaust port. The design of the cylinder head would also be suitable for gravity die casting if larger production numbers were needed. The camshaft was supported by three bearing caps, to allow the use of a fully isolated plastic cover for NVH benefits" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001472_9781118354179.auto089-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001472_9781118354179.auto089-Figure5-1.png", + "caption": "Figure 5. Diaphragm spring clutch. (Reproduced from Nunney, 1998. \u00a9 Elsevier/Butterworth Heinemann.)", + "texts": [ + " In addition to such a design, normal force can be provided in a number of ways: using diaphragm spring, large central coil spring, centrifugal force, magnetic force, hydraulic pressure, and others. Some of the most commonly used and interesting solutions will be presented. Encyclopedia of Automotive Engineering, Online \u00a9 2014 John Wiley & Sons, Ltd. This article is \u00a9 2014 John Wiley & Sons, Ltd. DOI: 10.1002/9781118354179.auto089 Also published in the Encyclopedia of Automotive Engineering (print edition) ISBN: 978-0-470-97402-5 Figure 5 (Nunney, 1998) shows the most common design used today on road vehicles, a diaphragm spring clutch assembly. Nominally, all clutch assembly components shown in Figure 4 are present; however, the spring is of diaphragm type, not requiring separate disengagement levers [marked (5) in Figure 4]. For this role, the diaphragm spring has slots toward the inner diameter, which take on the role of the levers, enabling axial movement (disengagement) of the pressure plate. Fulcrum rings (Figure 5) provide \u201cpivots\u201d for relative movement (flexing) of the spring in relation to the clutch cover, minimizing spring contact stresses, and ensuring high mechanical efficiency. The rivets connect the spring, via the wire rings to the cover. For torque transfer, separate flexible straps are positioned in the circumferential direction and attached to the pressure plate and clutch cover. This clutch design is the most compact, lightest, and has other advantages over other types. The crucial component is the spring itself, which is highly loaded and requires special materials and manufacturing methods", + " With the introduction of engine start-stop technology and mild hybrid vehicles it has become viable to offer an \u201coff the shelf\u201d solution (by ZF) integrating the clutch and electric motor, as shown in Figure 12. Such solution can work well even with manual transmission, with the clutch and motor functions complementing each other and reducing vehicle fuel consumption and emissions. This is the most commonly used design and the presentation here will be concentrating on the typical assembly and component designs and materials. Figure 5 shows a cross section of the typical diaphragm clutch design, with all components of the assembly. Figure 13a show a photograph (from ZF Sachs) from the \u201cengine side,\u201d including the friction disc. Figure 13b shows \u201cgearbox side\u201d view. Figure 14 (from ZF Sachs) shows the power flow\u2014when the clutch is engaged and disruption to power flow when the clutch is disengaged. For the disengaged clutch, it can be clearly seen that all components of the clutch assembly rotate with the engine flywheel, with the exception of the friction disc. The main clutch components shown in Figure 5 are for the \u201cpush-type\u201d clutch. This is the most common type, but it Encyclopedia of Automotive Engineering, Online \u00a9 2014 John Wiley & Sons, Ltd. This article is \u00a9 2014 John Wiley & Sons, Ltd. DOI: 10.1002/9781118354179.auto089 Also published in the Encyclopedia of Automotive Engineering (print edition) ISBN: 978-0-470-97402-5 is important to mention that \u201cpull-type\u201d can also be found on a number of vehicles. The principle of operation for the two types is schematically shown in Figure 15 (for more information, see Shaver, 1997), with springs in the preinstallation shape indicated on (a)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001195_iecon.2015.7392254-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001195_iecon.2015.7392254-Figure8-1.png", + "caption": "Fig. 8. Isometric and side view of the DC cable reel and the power electronics included in a rectangular envelope of the size of the AC reel", + "texts": [ + " Due to the higher voltage, the cable armor becomes thicker and the relation between copper and armor is shifted towards more armor. This is reflected by the different reduction of mass and cable diameter. By further increasing the voltage, at one optimal point the cable diameter will be at its minimum after which it will become larger again. To conclude the comparison, the storage volume and weight of both solutions will be examined. To do so, a rectangular envelope with the dimensions of the AC reel will be the reference. Fig. 8 shows how the DC system could be fitted into this envelope, with the cable reel in one corner and power electronics at the side of it. With this configuration, the total storage volume can be reduced by about one fifth, while a volume of over 350 l is still available for the power electronics. Assuming a needed volume of 18 l for the transformers and semiconductors of all six converters, and furthermore assuming these components contribute at least 25% to the total volume of the converter, the available volume is sufficient for the power electronics" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003470_2005-01-2295-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003470_2005-01-2295-Figure1-1.png", + "caption": "Fig. 1: Simple Gear Dynamics Model", + "texts": [ + " Fte = an effective transmission error force that also accounts for mesh stiffness variations (acts in direction of the plane of action and also in axial direction) Fsh = the shuttling force that is calculated from static analysis of the mesh force (acts in the direction of the plane of action) Ff = friction force computed by static analysis at the mesh (acts at a right angle to the plane of action) Fal = force due to the entrapment of air and lubricants (direction not clearly defined) Because our goal is to develop an excitation measure that may be applied by gear designer, the above equation will be posed as a static excitation in this paper. An alternative would be to perform a dynamic analysis of the gears and their supporting shafting in order to calculate a dynamic bearing force, in which case one would need to use the mesh force as the excitation [12,13]. Because we are calculating a static excitation, approximations will be needed for Fi, Fte and Fal, since they will be affected by dynamics. Approximate Forces The simplest reasonable model for performing gear dynamics modeling is shown in Fig. 1. The transmitted forces lie on the line of action, the line that is tangent to the respective base circles of the two meshing gears. The transmission error is modeled as a displacement input where, ba xxTE (1.2) The impact force that is due to corner contact is approximated to be along the line of action, and a second impact force due to the viscous sliding impact is taken at right angles to the line of action, as is the friction force. In our subsequent analyses, we are going to assume that these forces are of such short duration that their mesh frequency components are negligible", + " We are also going to neglect the entrapment force, since it is felt that it becomes important only for extremely high-speed gearing. We essentially have a pretty good feel for the evaluation of the shuttling force and the friction force (other than the approximation of the coefficient of friction), but in terms of obtaining an equivalent transmission error force, one has to look very closely at the dynamics of the situation. Fig. 2 shows the respective transmission error and friction excitation frequency [14] for a model similar to that of Fig. 1. The line-of-action (LOA) transmissibility indicates bearing forces that result from transmission error excitations, and the off-line-of-action (OLOA) transmissibility indicates bearing forces due to friction that occurs at right angles to the line of action as indicated in Fig. 1. We see here that at very low speeds, the excitation due to transmission error is very small and as speed increases, the effect of the force also increases as we traverse through the first natural frequency of the system. Above that natural frequency, the force levels off some, before it excites the next natural frequency of the system. Fig. 2: Hochmann\u2019s Simplified Model Results In order to put a rational value into the static excitation model, we are going to use the relationship: TEkF mavete (1.3) where kmave = the average mesh stiffness This approximation would provide a valid mesh force if we are operating above the lowest torsional natural frequency of the simple system of Fig. 1. Another way of looking at this force is that it is the bearing force that would result if the inertias were very large so that they remain vibrationally stationary when running at speed. For helical gears, this force has two components, one along the line of action and a second axial force that is evaluated by multiplying the line of action force by the tangent of the helix angle, as shown below: )tan(TEkF maveTEA (1.4) where = helix angle The key to performing an appropriate analysis of the mesh excitation is the evaluation of the distribution of forces along the contact lines" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001472_9781118354179.auto089-Figure29-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001472_9781118354179.auto089-Figure29-1.png", + "caption": "Figure 29. Automated hydraulic clutch control. (Reproduced with permission from LuK GmBH & Co. KG.)", + "texts": [ + " In addition to the above-mentioned modern pneumatic clutch actuation system (Figure 28), which can be automated, two other systems shown earlier, centrifugal clutch (Figure 6) and electromagnetic clutch (Figure 7), can also be relatively easily automated. Both of these are older designs not currently used. Instead, sophisticated hydraulic and electric systems have been developed, as shown in Figures 29 and 30. Obviously, such systems require outside power to operate, that is, hydraulic pressure for the (LuK) system shown in Figure 29 and electric current for the (ZF) system shown in Figure 30. Fenton (1996) gives more detailed insight into the evolution of automated clutch control systems. Figure 31 shows a clutch schematic and basic dimensions. The outer diameter of the friction disc is OD, inner ID, and total spring force, normal on friction surfaces, is FN. Engine angular velocity is \u03c9E and clutch friction disc velocity is \u03c9C. It should be noted that engine velocity \u03c9E is identical to the flywheel velocity and all clutch components fastened to the flywheel also have the same velocity" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001650_pcicon.2015.7435110-Figure14-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001650_pcicon.2015.7435110-Figure14-1.png", + "caption": "Fig. 14 Vibration measurement locations", + "texts": [ + " After all possibilities have been investigated or ruled out, a more focused in depth trouble shooting procedure must be conducted. B. In-depth Trouble Shooting Process This process may require sophisticated vibration analysis equipment and expertise to acquire the necessary information. Too little information, incorrect or unclear information can unnecessarily increase the time of the process and sometimes lead to wrong conclusions. Vibration levels should be measured in three directions (horizontal, vertical, axial) on each housing of the machine (shown in Fig. 14) and in two locations on the shaft if so equipped (90 degrees apart). It may be necessary to add shaft measurement probes if the resolution is not identified with housing probes only. The key information to be captured is as follows: 1) Vibration Levels: overall, unfiltered, and at critical discrete frequencies These measurements should be taken under the conditions listed below. It may be valuable to observe how vibration levels change with motor temperature which can be accomplished through transient loading and offloading conditions" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002381_j.triboint.2014.04.001-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002381_j.triboint.2014.04.001-Figure1-1.png", + "caption": "Fig. 1. Main landing gear (MLG) [1].", + "texts": [ + " In civil aviation, aircraft are required to operate on a variety of runways. As a result, investigations into aircraft maneuvering on rough runways are necessary. High loads on the shock absorber bearings and high sliding speeds induced by rough runways lead to excessive heat generation at the slider bearings, eventually causing structural damage. The root cause of the reported overheating issues has been postulated to take place at the lower bearing\u2013piston sliding interface of the main landing gear (MLG) (see Fig. 1). Since the derivation of the Reynolds equation for very thin fluid films more than a century ago, elasto-hydrodynamic (EHD) studies have focused on the isothermal performance in order to design high efficiency bearings. For more than 30 years, lubrication studies have been extended to include temperature effects [2]. Many of the thermo-elasto-hydrodynamic (TEHD) studies were steady-state and focused on the performance of the bearings [3\u20138]. Recent literature on the performance of slider bearings is very scarce [9]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003869_851385-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003869_851385-Figure1-1.png", + "caption": "Figure 1. Initial ET TPS Material Design", + "texts": [ + " 2 engines during launch and the ascent phase. It also serves as the structural backbone for the space shuttle system accommodating the complex load paths. INITIAL ET TPS CONCEPT The current External Tank (ET) design is much the same as originally conceived in 1972. Significant changes in the Thermal Protection System's (TPS) materials and processes have been made, driven by both changes in requirements and material/processes evolution. The ET design with the initial choice of TPS materials as it existed in 1973 is shown in Figure 1. The materials selected consisted of BX-250 urethane spray-on foam insulation (SOFI) on most of the LH2 tank and Superlight Ablator, SLA-561, on high-heating-rate areas of the L02 tank, nose cap, the intertank and the LH2 aft dome. An ablator is a material specifically designed to withstand high temperatures through controlled sacrificial degradation (energy absorption)--normally used as heat shields for re-entry of orbiting spacecraft. Ablators can withstand significantly higher heating rates than foams; however, the SLA-56l density is 16 pcf, much higher than the BX-250 density of 2 pcf" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000283_mop.29797-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000283_mop.29797-Figure1-1.png", + "caption": "Figure 1 (a) Construction of arms of Fibonacci spiral. (b) Structure of proposed Fibonacci spiral antenna (w 5 400 lm, ts5 1.5 mm, tp5 0.0175 mm. (c) Standard logarithmic spiral antenna. [Color figure can be viewed in the online issue, which is available at wileyonlinelibrary.com]", + "texts": [ + " Further we report the fractalization of the discontinuous geometry of FSA using Koch curve. Reference 5 reports Koch curve embedded over dipole antenna which transforms it into multiband antenna. It is found that the fractalized FSA resonates at nearly the same frequencies as simple FSA without Koch curve. However, the size of antenna has reduced to nearly two third of its original dimension though it still radiates unidirectional beam efficiently with an average directivity of 5.8 dBi. 2.1. Logarithmic Spiral Antenna A standard logarithmic spiral [Figure 1(c)] is designed on the basis of equations given below[6]: r15r0eah (1) r25r0ea\u00f0h2h0\u00de (2) where r1 is inner radius, r2 is outer radius, r0 (5 0.825 mm) is initial radius, h the incremental angle, a (5 0.306349) the progression factor, and h0 (758) the phase shift of the spiral. The progression factor (a) and initial radius (r0) of LSA have been chosen such that its performance could be compared with FSA. DOI 10.1002/mop MICROWAVE AND OPTICAL TECHNOLOGY LETTERS / Vol. 58, No. 6, June 2016 1315 2.2. Fibonacci Spiral Antenna The Fibonacci spiral as shown in Figure 1(a) is made up of a series of quarter-circular arcs whose radii are successively increasing Fibonacci numbers (1, 1, 2, 3, 5, 8, 13, 21, 34) which is governed by recursive formula: Xn5 Xn21 1 Xn22 (3) where X051 and n 5 1, 2, 3, . . .. 8. To construct a spiral arm of FSA [Figure 1(b)], eight quarter circular arcs have been taken whose radius varies from 1 to 34 mm as mentioned earlier. 2.3. Fractalized Koch-Fibonacci Spiral Antenna Koch curve is a simple fractal geometry that starts with a straight line of length x0 (sectorial length of arc having radius 34 in present case). The first limiting Koch curve is derived by splitting this straight line into three equal segments and then 1316 MICROWAVE AND OPTICAL TECHNOLOGY LETTERS / Vol. 58, No. 6, June 2016 DOI 10.1002/mop substituting two line segments of same length as the middle one with indentation angle of 608 from initial line as indicated in Figure 2(a)", + " The antenna geometries have been simulated and fabricated over lossy FR-4 substrate with dielectric constant er 5 4.33, loss tangent tan d 5 0.0012, and thickness, ts 5 1.5 mm. The simulation results are obtained using Finite Integration Technique (FIT) based CST software. The testing of the prototype antenna was done on E5071C ENA series network analyzer for 0.3 to 20 GHz. Unbalanced feeding technique has been used for excitation of the antennas [7]. One of the spiral arms is fed with a coaxial cable of impedance 50X and the other is shorted to ground with a parasitic element as shown in Figure 1(c). It provides quasi anti-phase excitation. Figure 3 shows the magnitude of reflection coefficients for the antennas in the frequency range 4\u201310 GHz. This frequency band is widely used in space research radio navigation, satellite downlink for broadcast, etc. The multiband behavior is evidently because of conducting ground plane just beneath the antennas. Since the effective arm track length of Koch-FSA as given by 1318 MICROWAVE AND OPTICAL TECHNOLOGY LETTERS / Vol. 58, No. 6, June 2016 DOI 10" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003937_978-1-4471-5104-3_5-Figure5.13-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003937_978-1-4471-5104-3_5-Figure5.13-1.png", + "caption": "Fig. 5.13 Typical magnetization curve of ferromagnetic materials", + "texts": [ + " The remanence is the maximum flux density that the magnet can produce by itself. On the other hand, if magnetic field intensity has increased in the opposite polarity, the flux density will eventually become zero. The magnetic field intensity at this point is called the coercivity, denoted Hc. The absolute value of the product of the magnetic flux density, Br and the magnetic field intensity, Hc at each point along the magnetization curve in the second quadrant region is called the energy product, or the density of magnetic energy as it is illustrated in Fig. 5.13. The higher the density of magnetic energy is, the smaller the machine dimensions and core losses are. The operating point can be determined by the intersection between the permeance line and the hysteresis curve in the second quadrant region. The permeance line is determined by the characteristics of the machine structure: air gap length, magnetic path length, and number of coil turns. The large Br produces a large flux in the machine, while the large Hc means that a very large current would be required for demagnetizing the poles [13]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001066_detc2006-99207-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001066_detc2006-99207-Figure3-1.png", + "caption": "Figure 3. Construction for force transmission in the symmetric four-bar (transmission angle shown as dotted lines).", + "texts": [ + " While this is a good initial guess for relative location of the hip, it may not result in optimal force transmission over the whole range of pedal motion, so a parallel shift or offset of this line can be introduced as an additional free variable (shown as d in Figure 2). For any instantaneous position of the mechanism, the Kennedy-Aronhold theorem [13] dictates that the instant center I13 which determines the center of curvature of the coupler curve can be located by finding the intersection of the (extension) line of link 2 and the (extension) line of link 4, as shown in Figure 3. When the crank and rocker links are parallel, the instant center lies at infinity, and the coupler curve has zero curvature at this point. 2 Copyright \u00a9 2006 by ASME ms of Use: http://www.asme.org/about-asme/terms-of-use Downlo As shown in Figure 3, the input force from the hip and the output torque at the crank can be related as follows. The angle HPI13 determines the portion of the input force which is converted to output torque; the transmission ratio (instantaneous mechanical efficiency) is 13sinTR HPI= \u2220 (5) For a unit-length crank, the output torque fluctuates dependent on the changing location of the coupler point and instant center I13 as follows. 13 2 13 13 23 sinN I PT F HPI I I = \u2220 (6) where FN is the directed force applied from the hip H to the pedal P, and I13P and I13I23 are line-segment distances between the respective points" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002584_1999-01-0782-Figure9-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002584_1999-01-0782-Figure9-1.png", + "caption": "Figure 9. FEA analysis of first generation prototype.", + "texts": [ + " For preliminary analytical work however, this modeling scheme seemed reasonable. The prototype wheel design was for a 6X15 5 passenger car wheel and for analysis, a 4,500 N nominal wheel load was assumed. The SFI and SAE specifications require that for testing purposes the amount of bending moment applied to the wheel hub be determined by the formula: (3) For modeling, the wheel rim was constrained to prevent translation in any direction. A load was then applied to the end of the attached shaft and the part was meshed and analyzed. Results of the analysis are shown in Fig. 9. Peak stresses are well within the range for predicting reasonable fatigue life. Axle deflection values based on the FEA analysis are typically less than that for a conventional stamped steel dish wheel of equivalent size. Typically by a factor of 1.5 to 2. Stresses resulting from torque, radial and dynamic forces were examined and were so small that for first order examinations like these studies, it seemed best to ignore them. As a practical matter, the analytical aspects of this wheel development program are concerned with strength and durability of wheels as influenced by cornering forces only" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002355_aims.2015.43-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002355_aims.2015.43-Figure3-1.png", + "caption": "Figure 3. Dimension of aluminium casing of MOPICs trainer.", + "texts": [ + " DESIGN STAGE The PICs provide analogue and digital series of input and output that can be used to control the MOPICs trainer. Fig. 2 shows the block diagram of MOPICs trainer. It consists of PIC Module, Input Module, Output Module and DC Power Supply. IV. FABRICATION AND DEVELOPMENT OF MOPIC TRAINER The fabrications of MOPICs trainer are set to four different segments which are the casing that is made of aluminum steel with 35o slanting panel, PIC module, communication (via USB cable from PC to MOPICs) and power supply unit respectively. Fig. 3 shows the side and top view with dimensions (in inch) of compact MOPICs trainer. Furthermore, the aluminum steel as the front panel are divided into three main segments which are PIC Programming segment, Input segment and Output segment respectively as in Fig. 4. Meanwhile, on top of the panel, as for safety precaution, transparent plate is installed as to avoid the leaners accidentally touch any rotating components such as the servomotor or 12VDC motor during operation. The PIC Programming Section is basically consists of RUN/PROG switch and Reset Button" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000234_icelmach.2016.7732762-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000234_icelmach.2016.7732762-Figure7-1.png", + "caption": "Fig. 7. No load flux lines and flux density of mcC: PM magnetization direction as in Fig. 6", + "texts": [ + " On one hand, good performance of the 4- pole machine are achieved when the flux barrier angle is at about 45 degrees with respect the PM magnetization axis. On the other hand, to achieve good performance in the 2-pole machine, the flux barrier angle must be increased with respect the PM magnetization axis. The selected flux barrier angle is a compromise to achieve satisfactory performance both in the 2- and 4-pole machine. McC has the advantage that it contains a lower PM volume (about 35%) with respect mcA and mcB. Fig. 7 shows the flux lines distribution and the correspondent flux density at no load for mcC. Its performance are shown in Table VI. Similarly to mcB, the estimated efficiency exceed the IE4 efficiency class limit both in the 2- and 4-pole machine. This is an interesting result, considering the lower PM volume of the induced pole rotor structure. However, it should be mentioned that the volume of mcC is 32% higher than the mcA one and 20% higher than the mcB one. It is worth noticing that the 4-pole machine exhibits a high percentage of iron losses, even if the no-load flux density in the stator parts are rather low" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003030_s11263-009-0223-3-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003030_s11263-009-0223-3-Figure1-1.png", + "caption": "Fig. 1 A cross-shaped camera configuration. The reference and the supporting cameras are labeled ref, left, right, top and bottom respectively. An example of occluder and occluded pixels are shown in black and white respectively", + "texts": [ + "eywords Stereo \u00b7 Occlusion \u00b7 Multi-camera \u00b7 Dynamic programming The goal of binocular stereo is to reconstruct the 3D structure of a scene from two views. Occlusion occurs when part M.-A. Drouin ( ) \u00b7 M. Trudeau \u00b7 S. Roy D\u00e9partement d\u2019informatique et recherche op\u00e9rationelle, Universit\u00e9 de Montr\u00e9al, Montr\u00e9al, Canada e-mail: marc-antoine.drouin@nrc-cnrc.gc.ca Present address: M.-A. Drouin Institute of Information Technology of the National Research Council Canada, Ottawa, Canada of a scene is visible in the reference but not in the supporting camera (Fig. 1). The difficulty of detecting occlusion comes from the fact that it is induced by the 3D structure of the scene, which is unknown until the correspondence is established. When extending binocular stereo to multiple cameras, the amount of occlusion increases since each pixel of the reference camera can be hidden in more than one supporting camera. This is particularly true when going from a single to a multiple-baseline configuration. We propose a novel algorithm that improves the localization of disparity discontinuities of disparity maps obtained by multi-baseline stereo", + " Furthermore, in some high level tasks such as navigation or recognition, a coarse resolution may be sufficient for most parts of a disparity map. However, for a specific situation or region of the disparity map a more accurate border localization may be required. Our framework is perfectly adapted to support such high level tasks that may selectively require improvement of the disparity map. In this paper, we used a configuration of five rectified images, equally spaced and arranged in a cross (Fig. 1). The disparity map is reconstructed from the point of view of the center camera which we call the reference. This differs from multi-view stereo which aims at reconstructing a full 3D model. A survey paper by Seitz et al. (2006) compares various multi-view algorithms. The rest of this paper is divided as follows: in Sect. 2, previous work is presented; Sect. 3 describes our algorithm; visibility is discussed in Sect. 4; experimental results are presented in Sect. 5. In Egnal and Wildes (2002), five basic strategies to overcome occlusion for two cameras are presented: left-right checking, bimodality test, goodness jumps constraint, duality of depth discontinuity and occlusion, and uniqueness constraint" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003460_esars.2015.7101415-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003460_esars.2015.7101415-Figure1-1.png", + "caption": "Fig. 1. Geometry of a five-phase Halbach array permanent magnet synchronous motor", + "texts": [ + " The iron losses obtained by pure finite element, classical two points method and proposed six points method are then compared. A 3-D finite element thermal analysis is finally performed and temperature distribution in different parts of the motor are determined. To prevent demagnetization due to high 978-1-4799-7400-9/15/$31.00 \u00a92015 IEEE temperature rise in permanent magnets, a water cooling system is designed and its effect on the temperature distribution of the motor is investigated. II. MODEL OF THE MACHINE The machine geometry is shown in Fig. 1. The stator has five phase ten slot configuration with concentrated winding. The 4-pole rotor consists of 16 pieces of NdFeB magnets. In order to implement the Halbach array, the magnets are magnetized such that they have 22.5 degree phase shift with respect to each other. Table I shows the main design specifications of the machine under study. A. Iron losses The real flux density in different parts of the machine is non-sinusoidal and therefore extra iron losses are produced by flux density harmonics", + " As it is seen the flux density in the tooth is uniform unless in the tooth shoe. Three points are chosen in the tooth. Stator yoke is also divided into two parts where the part above the tooth has usually lower flux density in comparison with the part between two teeth. Finally one point is selected in the middle of magnet. As the rotor yoke iron loss is negligible no point is selected in this region. In this section the proposed method is applied to the Halbach array permanent magnet machine shown in Fig. 1. The flux desnity of points one to six and the harmonic contents of them are illustrated in Figures 3 to 9. It is seen in Fig. 3, that the flux density in point one has tangential and radial components which casues highest iron loss density in this region. As it is seen in Fig. 4, point two has similar condition but the amplitude of flux density in both direction is less than point one. The flux density at point three and four are depicted in Figs. 5 and 6. It is shown that the flux density in tangentil direction is negligible with respect to the radial direction" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003263_1.4031902-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003263_1.4031902-Figure1-1.png", + "caption": "Fig. 1 Sheave grove geometry [6]", + "texts": [ + " The sheaves of a 2 in. wire (50.8 mm) drill line are generally subjected to a maximum wire load of approximately 69 tons for 16 parts and 78.5 for 14 part systems. Therefore, the load on wire (F) of 90 tons is considered as the design wire load in this study to take account of the uncertainties behind load modeling. The in-plane distributed load on the sheave per unit length, q is q \u00bc 2F Dt (1) where 2F is the total vertical load acting on the sheave, and Dt is the inner diameter of the sheave as shown in Fig. 1. The in-plane distributed load on the sheave per unit length (q) can be rewritten as q \u00bc \u00f0b=2 b=2 Prgroove cos hdh (2) 1Corresponding author. Contributed by the Ocean, Offshore, and Arctic Engineering Division of ASME for publication in the JOURNAL OF OFFSHORE MECHANICS AND ARCTIC ENGINEERING. Manuscript received January 2, 2015; final manuscript received October 22, 2015; published online November 13, 2015. Assoc. Editor: Myung Hyun Kim. Journal of Offshore Mechanics and Arctic Engineering FEBRUARY 2016, Vol. 138 / 014501-1 Copyright VC 2016 by ASME Downloaded From: http://offshoremechanics.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/journals/jmoeex/934798/ on 02/21/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use where P is the pressure acting on the contact surface of the groove, rgroove is the radius of the groove, h is the angle to the considered point on the sheave measured from the vertical axis, and b is the contact angle of the wire on the groove as shown in Fig. 1, i.e., of 150 deg [6]. From Eqs. (1) and (2), the pressure acting on the contact surface of the sheave groove due to the wire load is derived as P \u00bc F Dtrgroove sin b 2 (3) Hence, the design external pressure (P) of the 72 in. sheave (i.e., sheave outer diameter (D) is 1829 mm and sheave inner diameter (Dt) is 1684 mm) due to design wire load of (F) of 90 tons can be calculated as 20.1 MPa. 2.2 Design Side Load. In addition to the pressure normal to the groove, there can also be a possibility of acting an out-ofplane load due to uncertainty of angled nature of line of action of wire load (F)", + " The S\u2013N curve C in air is recommended to use cast design geometries in order to allow for weld repairs after possible casting defects and possible fatigue cracks after some service life [1,5,9]. Seven different design geometries of wire sheaves were considered for this study. The considered design geometries are distinguished as double web, double web with holes, straight web, straight web with holes, thin web with stiffeners, thin web with stiffeners and holes, and web with decreasing thickness as shown in Fig. 2. The geometry of the wire grove (i.e., shown in Fig. 1) is same for all the design geometries. The wire sheave curved in a plane can be considered as curved thin-walled beams. Therefore, it is comfortable to use the explicit formulas proposed in literature [11,12] for evaluating stresses and critical buckling loads. However, out of the seven proposed sheave geometries, five does not have constant cross section throughout their curvature as they consist of discontinuities, such as holes and stiffeners as shown in Fig. 2. Since this is one of the major assumptions behind the above-proposed formulas, this may be a hindrance for applying these explicit formulas for majority of the proposed sheave geometries" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000284_s10846-005-0932-y-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000284_s10846-005-0932-y-Figure1-1.png", + "caption": "Figure 1. The PSVS: (a) mirrors position, (b) PSV apparatus mounted on the end effector of a PUMA 761 robotic manipulator.", + "texts": [ + " Refraction phenomena do not appear to first three mirrors because the first surface of them is used (first surface mirrors). To determine the relative location of mirrors in PSVS, a right-hand orthogonal coordinate system is defined. Z-axis of this system coicides with the optical axis of the real camera and the origin of it is the optical center O of the camera. X-axis, vertical to Z-axis, is parallel with the direction of columns increment in the image plane and Y -axis is vertical to the plane XZ. Mirrors of PSVS are vertically located to XZ-plane and form 45\u25e6 angle with Z-axis (Figure 1(a)). It is considered that initially no refraction phenomena exist due to mirror (1) (i.e. by using a Pellicle beam-splitter). Then two virtual cameras are created with their optical axes parallel to the optical axis of the real camera. These cameras are symmetrically located to Z-axis. They have the same geometric properties and parameters (the same of the real camera). Consequently, these virtual cameras constitute an ideal stereovision system with two cameras. This vision system, as it presented here, receives in a single shot one complex image", + " If the intensity of each pixel of an image captured from the left and from the right view are IL(i, j) and IR(i, j), respectively, the intensity of each pixel of the complex image is given as: IC(i, j) = kIL(i, j) + (1 \u2212 k)IR(i, j), (1) where i, j , are indices of the current row and column of a pixel in the image, k is a parameter (k = 0.5 with the beam-splitter used) declaring the reduction of the intensity of pixels of each view because of the beam-splitter. It is obvious that the intensity in a complex image never exceeds its maximum value (for gray-scale images IC(i, j) 255). The baseline of this stereo system is b and it is the distance between the two virtual parallel optical axes (Figure 1(a)). Problems with correct luminosity are reduced from the system by using a regulated lighting system on the apparatus (Figure 1(b)). A small laser module is incorporated to the apparatus. The red spot laser beam is used to periodically check the alignment of mirrors. The front view of the PSVS is properly formulated to accept \u201cspatial\u201d or color filters. In PSVS, to avoid probable shades of parts of complex images because of the improper size or of the location of mirrors and the appearance of ghost images, the calculation of mirrors dimensions is required. The equation providing these dimensions is the following: AC = AB + BC \u21d2 AC = ( OB \u00b7 tan a \u00b7 cos \u03c91 tan(\u03c91 \u2212 a) + OB \u00b7 tan a \u00b7 sin \u03c91 ) + OB \u00b7 tan a sin \u03c91 + cos \u03c91 \u00b7 tan a ", + " Then AJ = AL \u00b7 cos \u03c91 = ( OA \u00b7 tan a \u00b7 cos \u03c91 tan(\u03c91 \u2212 a) + OA \u00b7 tan a \u00b7 sin \u03c91 ) cos \u03c91, (22) BK = BM \u00b7 sin \u03c91 = (OA + ABmin) tan a1 sin \u03c91 + cos \u03c91 \u00b7 tan a1 \u00b7 sin \u03c91, (23) JR = LS \u00b7 cos a = d cos \u03c91 \u00b7 cos a. (24) Substituting Equations (22)\u2013(24) to (21) after some calculations, it results: ABmin = sin \u03c91 + cos \u03c91 \u00b7 tan a sin \u03c91 + cos \u03c91 \u00b7 tan a \u2212 tan a \u00b7 sin \u03c91 \u00b7 [ OA \u00b7 tan a \u00b7 cos2 \u03c91 tan(\u03c91 \u2212 a) + OA \u00b7 tan a \u00b7 sin \u03c91 \u00b7 cos \u03c91 + OA \u00b7 tan a \u00b7 sin \u03c91 sin \u03c91 + cos \u03c91 \u00b7 tan a + d cos \u03c91 \u00b7 cos a ] . (25) For angle \u03c91 = 45\u25e6, (25) is simplified to ABmin = OA \u00b7 2 tan a (1 \u2212 tan a) + \u221a 2d cos a \u00b7 (1 + tan a). (26) To calculate the Minimum Distance of Common View, Figure 1(a) is used. The distance of interest is OB. The point B represents the first common point, which is created from the two virtual cameras, where no refraction phenomena are taken into consideration. From the right triangle (O1AB), it results: tan a = AB O1A \u21d2 O1A = AB tan a \u21d2 OB + b 2 = b/2 tan a \u21d2 OB = b 2 tan a \u2212 b 2 . (27) In this part, the influence of refraction phenomena due to mirror (1) of PSVS is examined. It is desirable, the left and right view of a scene captured by means of PSVS to coincide and to have exactly the same magnification", + " Then the result is a ray radiated from the optical center O to follow slightly different in length paths in two different directions. The magnification of the two virtual cameras is exactly the same and the displacement of the second virtual axis with respect to the optical axis of the real camera is b/2 + m + l. In this case the construction and the calibration procedure requires the careful placement of mirrors (3) and (4). The corresponding points are always found in the same scan line (epipolar line). For the alignment of mirrors of PSVS (Figure 1(a)) an apparatus was used, which consists of a two meters tube with square cross-section (8 \u00d7 8 cm) and proper constructed holders to hold on the tube, during the alignment, the pseudo-stereo apparatus, the targets and a laser source (Figure 5(a)). More details can be found to the paper of Pachidis and Lygouras [16]. Two targets are used in the apparatus. The first target is a glass made surface with parallel horizontal and vertical lines in distances of 1 mm and the second one, the same as previously but with substrate a 3 mm thick piece of aluminum" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003255_14484846.2004.11464471-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003255_14484846.2004.11464471-Figure1-1.png", + "caption": "Figure 1: Ten-stage turbo-molecular pump with O-rings in order to improve the dynamic behaviour", + "texts": [ + "11464471 To link to this article: https://doi.org/10.1080/14484846.2004.11464471 Published online: 22 Sep 2015. Submit your article to this journal Article views: 15 View related articles 91 \u00a9 Institution of Engineers, Australia 2004 Australian Journal of Mechanical Engineering, Vol 1, No.2 conference paper * Paper presented at the Rotordynamics conference 2002 Simple O-rings are particularly suited to provide necessary damping for rotors with roller bearings. This was proven in many cases, for example in turbo pumps, Fig. 1. The behaviour of journal and magnetic bearings as well as of squeeze film dampers is widely covered by previous research and presented in numerous publications around the globe for many years.1,2 However there is a surprising lack of literature about elastomer dampers apart from a series of NASA reports between 1972 and 1980.3,4 In view of the fact that the dynamic properties of elastomers can only be derived from expensive tests, trial and error seems to be the prevalent method in using O-rings as dampers" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000854_6.1995-1779-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000854_6.1995-1779-Figure6-1.png", + "caption": "Figure 6. Real Afterbody Surface Grid", + "texts": [ + " The CFD model included nearly all aspects of the wind tunnel model, including the wing tip mounting system, inlet fairing, centerbody missile fairing, a missile fin cut-out in the wing root trailing edge, horizontal and vertical tails, flowing nozzle, aft facing nozzle base area, dummy sting, and the arresting hook fairing. The geometric compromises made were generally minor. The most significant were: control surface trailing edges were sharpened, the nozzle base was modeled as a flat surface, the sting cavity was modeled as an aft facing step, and the gap between the horizontal tail and nozzle (which allows the nozzle flaps to move in response to throttle setting) was filled. A representative view of the real afterbody surface grid is shown in Figure 6. Note that only 10,000 of 140,000 surface points are shown for clarity. The inlet fairing and missile fairing can be seen in this figure. Closeup views of the real and distorted afterbody grids are given in Figure 7, again with only a sparse surface grid shown. The complete grids for both the real and distorted afterbodies contained approximately five million points divided among sixty zones. The flow solver code employed for this study was the three-dimensional NASTD (NAvier-Stokes Time Dependent) code (Reference 2), a proprietary code developed at MDA" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001638_tim.2015.2477160-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001638_tim.2015.2477160-Figure3-1.png", + "caption": "Fig. 3. System of coordinates between a robot and a camera (right view).", + "texts": [], + "surrounding_texts": [ + "The general formulation to obtain stereo vision requires at least two images of the same object acquired in different positions, by the same camera that moved or by two cameras [24]. The proposed approach uses only a single camera installed on a robot. From the camera, a system of coordinates between the robot and the camera needs to be established [25]. As shown in Figs. 3 and 4, the robot coordinate is given by r xyz and the camera coordinate system is given by Oc XY Z , where z and x axes are parallel to z and x axes of Oc XY Z . The parameter a is the displacement between the robot coordinate and camera coordinate along the z-axis and h is the displacement between the robot coordinate system and camera coordinate system along the x-axis. The proposed approach uses targets in the scene for measuring velocity. In order to establish a geometry between the camera and the target, two feature points, labeled l1 and l2, in the target are selected. The coordinates l1(z1, y1) and l2(z2, y2) must be known (in our approach, they are extracted using the vision-based measurement (VBM) algorithm proposed in [23]). Then, a line segment l is created between these two points with l0 as its midpoint. (The segment is shown to be displaced in Fig. 4 for clarity.) Moreover, b is a displacement between l0 and the coordinate target system along z-axis and h is the height of the camera from the ground. Finally, an angle \u03b8l between the line l and the camera system of coordinates Oc XY Z (Fig. 4) is calculated as Z1 \u2212 Z2 = \u2212l(sin \u03b8l); Y1 \u2212 Y2 = l(cos \u03b8l). Considering (1), these points (l1 and l2) can be modeled in the image plane \u03c0 as l1 : Z1 f = \u2212h x \u2032 1 , Z1 f = Y1 y \u2032 1 l2 : Z2 f = \u2212h x \u2032 2 , Z2 f = Y2 y \u2032 2 . Thus, with respect to the robot pose, the target pose can be expressed in an image as T = [ Z1 Y1 Z2 Y2 sin \u03b8l cos \u03b8l ]T (4) or T = \u2212h [ f x \u2032 1 y \u2032 1 x \u2032 1 f x \u2032 2 y \u2032 2 x \u2032 2 f x \u2032 2l \u2212 f x \u2032 1l y \u2032 1 x \u2032 1l \u2212 y \u2032 2 x \u2032 2l ]T . (5) Converting the image coordinate into pixel coordinate using (2) and (3), it results in x \u2032 1 = \u2212dxm1, x \u2032 2 = \u2212dxm2 y \u2032 1 = dyn1, y \u2032 2 = dyn2 where m1 = (m1 \u2212 m0), m2 = (m2 \u2212 m0), n1 = (n1 \u2212 n0), and n2 = (n2 \u2212 n0) for ease of notation. Replacing x \u2032 1, x \u2032 2, y \u2032 1, and y \u2032 2 in (5), it results in T = h \u23a1 \u23a2 \u23a2\u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2\u23a2 \u23a2 \u23a2 \u23a2 \u23a2\u23a2 \u23a2 \u23a2 \u23a2 \u23a2\u23a2 \u23a3 f dxm1 dyn1 dxm1 f dxm2 dyn2 dxm2 \u2212 f ldx ( 1 m1 \u2212 1 m2 ) dy ldx ( n1 m1 \u2212 n2 m2 ) \u23a4 \u23a5 \u23a5\u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5\u23a5 \u23a5 \u23a5 \u23a5 \u23a5\u23a5 \u23a5 \u23a5 \u23a5 \u23a5\u23a5 \u23a6 . (6) As shown in Fig. 4, l0 is the center of the line l and the distance from l0 to the target center is given by b, thus the target center can be described in r xyz as (note that the coordinate system r xyz is in motion) zl = a + b cos \u03b8l + z1 + z2 2 (7) yl = b sin \u03b8l + y1 + y2 2 . (8) Finally, using the proposed approach, the target center pose can be expressed in image coordinates as \u23a1 \u23a2 \u23a2 \u23a3 zl yl sin \u03b8l cos \u03b8l \u23a4 \u23a5 \u23a5 \u23a6 = h \u23a1 \u23a2 \u23a2\u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2\u23a2 \u23a2 \u23a2 \u23a3 a + b dy ldx ( n1 m1 \u2212 n2 m2 ) + f 2dx ( 1 m1 + 1 m2 ) \u2212b f ldx ( 1 m1 \u2212 1 m2 ) + dy 2dx ( n1 m1 + n2 m2 ) \u2212 f ldx ( 1 m1 \u2212 1 m2 ) dy ldx ( n1 m1 \u2212 n2 m2 ) \u23a4 \u23a5 \u23a5\u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5\u23a5 \u23a5 \u23a5 \u23a6 . (9) Therefore, as shown in (9), the target center pose can be described in terms of images using the relationship between the camera installed in the robot and the target. However, (9) shows the target center pose described from image coordinates of the camera. In order to estimate the velocity, it is necessary to introduce the target center pose seen by the camera on the real space (i.e., w-space). To obtain the transformation between the image coordinates in w-space, first, a set of nonlinear transformations is introduced as w1 = ( 1 m1 \u2212 1 m2 ) (10) w2 = ( n1 m1 \u2212 n2 m2 ) (11) w3 = ( 1 m1 + 1 m2 ) (12) w4 = ( n1 m1 + n2 m2 ) (13) which transforms the terms (m1, n1, m2, n2) from image space to w-space with new terms (w1, w2, w3, w4). The second step is to add two dimensionless scalars s1 = h l f dx , s2 = h l dy dx . Finally, the target can be expressed in w-space as \u23a1 \u23a2 \u23a3 zl yl sin \u03b8l cos \u03b8l \u23a4 \u23a5 \u23a6 = \u23a1 \u23a2 \u23a2 \u23a2 \u23a3 ha + bs2w2 + l 2 s1w3 l 2 s2w4 \u2212 bs1w1 \u2212s1w1 s2w2 \u23a4 \u23a5 \u23a5 \u23a5 \u23a6 = \u23a1 \u23a2 \u23a3 ha 0 0 0 \u23a4 \u23a5 \u23a6 + A \u23a1 \u23a2 \u23a3 w1 w2 w3 w4 \u23a4 \u23a5 \u23a6 (14) where A = \u23a1 \u23a2 \u23a2 \u23a2 \u23a3 0 bs2 ls1 2 0 \u2212bs1 0 0 ls2 2\u2212s1 0 0 0 0 s2 0 0 \u23a4 \u23a5 \u23a5 \u23a5 \u23a6 (15) and a constraint is established in the last two terms of (14) as (s1 2w1 2 + s2 2w2 2 = 1). As a result, the w-space is defined in a 3-D space where w \u2208 R 3. Applying the time derivate in (14), the relative velocity between the robot and the target is obtained as \u23a1 \u23a2 \u23a3 z\u0307l y\u0307l \u03b8\u0307l \u23a4 \u23a5 \u23a6 = \u23a1 \u23a2 \u23a2 \u23a2 \u23a3 0 bs2 ls1 2 0 \u2212bs1 0 0 ls2 2\u2212s1s2w2 s1s2w1 0 0 \u23a4 \u23a5 \u23a5 \u23a5 \u23a6 \u23a1 \u23a2 \u23a2 \u23a2 \u23a3 w\u03071 w\u03072 w\u03073 w\u03074 \u23a4 \u23a5 \u23a5 \u23a5 \u23a6 . (16)" + ] + }, + { + "image_filename": "designv6_24_0000083_ir-03-2018-0042-Figure32-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000083_ir-03-2018-0042-Figure32-1.png", + "caption": "Figure 32 The wearable vehicle while walking by using ADAMS-MATLAB co-simulation", + "texts": [ + " The dashed line represents the front edge of the wearable vehicle support polygon. The dashed line represents the backward edge. The red line represents the center of the support polygon which is to be tracked by the ZMP, and the blue line is the controlled ZMP location. The obtained result shows that the ZMP is always within the support polygon edges, and hence the system stability while walking is ensured. This directly affects the comfort level and ergonomics for the wearers while carrying their luggage. The simulation results of a walking sample are obtained as seen in Figure 32. The reaction force between the feet and the ground strongly affects the behavior of the exoskeleton and human model. To solve this problem multiple modifications were applied to the wearable vehicle feet. For more realistic contact between the wearable vehicle and the ground, a proposed modification is to build the wearable vehicle feet from two parts and connect them with 6 axis springs (three linear springs, and three rotational ones) to simulate the flexibility of real human feet. The simulation of thismodification is shown in Figure 33" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001237_apmc.1997.659348-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001237_apmc.1997.659348-Figure1-1.png", + "caption": "Fig . 1 T h e gyoinetry o f a c i r cu la r ly polar ized n i ic ros t r ip a n t e n n a wi th a near ly s q u a r e p a t c h .", + "texts": [], + "surrounding_texts": [ + "glohil posilioning satellite system (GPS) application. The resonant frequencies, axial ratios, and iatlia~ion patlerns are nicasured. Here, the use of nearly square patch is proven that a successful ciri ular polarized antenna is developed.\n' I . INTRODUCTION 'I'his study is to design and fabricate the right-hand circularly polarized (RHCP) [ 1-61 microstrip antenna applied to GPS. the microstrip patch will be printed on Duroid substrate and the central operating lrequency is at 1575.4 MHz. The nearly square patch will bc' studied. '1.0 design the aspect ratio of tlie iic;arly sqii;ire p;ilcli, we will apply tlie full-wave ;il)pr(xich to analyc tlie incrostrip antenna. And the\noperating handwidIh and radiation pattern will also be calculated. 13asetl on the theoretical prediction tlie iiiicmitrip antenna will be fabricated and measured. I II tli is i n\\,cst igat ioii the circu lady polarized radiation i s obtained when tlie aspect ratio of the rectangular patch is about I . O l - I . O 3 , depending on the substrate ~liickness and perniittiviity. I n this case the antenna\ncan excite two orthogonal of the nearly square patch.\n'l'his kind o f design is the simplest form to generated\ncircular polarization and is very suitable for the GPS RHCP microstrip antenna design.\n2. THEORY", + "We use R03003 substrate which is a ceramic filled PTFE composite with substrate relative dielectric constant 2.98, loss tangent 0.0048. From simulation results, we design nearly square patch\n,I I\nn\ne , = ~ L % C + L G Y A ) ' (1-1 where y7A and <71 are tlie Fourier transform of 2nx and ,?,,, , and Gll , GAl , G, and G, are the Green's\nl'unctions [ I ] oi the antenna structure shown on biglire I . I n order to produce CP radiation at tlie I 11 ii I I I- hcai I 1 c l I rec t I ( ) n, tlie fo I I ow I ng condition i n 11 st be i a t i s l i cd a ( 0 = 0 . I e. ./'E,, LIS,,, =-too , ( 3 1\n5 5 . 0 5 ~ 5 4 . 3 5 ~ 0.762 min fed at 0.33 A B . Figure 2 shown tlie measured input impedance on a Smith chart with 2:l VSWR bandwidth is 22 MHz and Figure 3 shown the measured and calculated input impedance versus frequency. Figure 4 shown the measured axial ratio, the central frequency 1575.4 MHz is about 0.7 dB, and 3 dB CP bandwidth is about 6 MHz. Figure 5 and 6 shown the half power radiated beamwidtli with one-wavelength ground plane is about 70\", it is predicted that larger beamwidth can be reached with larger ground plane. Also tlie frequency drift is 1 MHz over 20-60\u00b0C . It is proven that a successful antenna is developed.\nF i g . 2 Measured i n p u t impedance on a Smi th cha r t f o r Wx = 55.05 inin, Wy = 54.35 inin, tan8 = 0.0048, E, = 2.98, d = 0 . 7 6 2 min 7 AC = 0 . 3 3 E .\n3. RESULTS AND CONCLUSIONS\n23 8", + "aJ W C\n-90\" I I\n-40 1 \" ' \" \" \" ' \" \" ' 1 \" ' 1 I 5 2 I 5 4 I 5 6 I 5 8 1 6 162\nFrequency (GHz) I xo\"\nFig . 5 Measured yz -p lane rad ia t ion pa t te rn fo r an tenna on a one - wave I en g t h g ro u n d p I an e .\nI XI)\"\nFig . 6 Measured xz -p lane rad ia t ion pa t te rn fo r a n t e n n a on a one-wavelength g round p lane ." + ] + }, + { + "image_filename": "designv6_24_0000188_s1526-6125(05)70089-x-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000188_s1526-6125(05)70089-x-Figure1-1.png", + "caption": "Figure 1 Setup of Negative Dieless Incremental Forming Process (Strano 2003)", + "texts": [ + " The validity of the proposed formability space has been tested by means of a statistical tool called binary logistic regression. Keywords: Sheet Metal, Dieless Robotic Incremental Forming, Formability, Binary Logistic Regression Two main variants of the dieless incremental forming process are known: the so-called \u201cnegative form- ing\u201d process and the \u201cpositive forming\u201d process. In negative incremental forming, a ball punch follows a trajectory on a sheet metal blank, according to a programmed tool path. The sheet is clamped at its periphery by bolts on a support frame (Strano 2003). In Figure 1a, the experimental setup used in this study is shown; in Figure 1b, the FEM simulation of a negative incremental forming operation with a spherical punch is shown. In positive forming, the central part of the workpiece is supported by a fixed counterpunch (or mandrel) and the tool-workpiece interface is located on the convex side of the shape to be formed. The sheet can be either fixed at the periphery (Park and Kim 2003) by a blankholder or it can be free. The latter case can be considered a direct evolution of the conventional shear spinning process (Wong, Dean, and Lin 2003)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002408_tap.2012.2220326-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002408_tap.2012.2220326-Figure6-1.png", + "caption": "Fig. 6. Long copper cylinder enclosed by the designed cloak.", + "texts": [ + " Let us first determine the total number . The periodicity around the cylinder circumference is , that is, (8) Substitute (8) into (7), we get (9) By taking all the given values of the MP structure parameters, we obtain the total number , where [.] denotes the integer part operator. Now the micro-strip line length can be easily determined by (8), that is, mm. Now that we have determined all the required structural parameters, the cloak can be implemented easily by 9 interconnecting unit cells to wrap the copper cylinder, as shown in Fig. 6. To verify the cloak, numerical simulations were carried out. Three steps were taken to determine the cloaking effect. First, an infinitely long copper cylinder with and without the cloak were simulated under exactly the same simulation setup (the same boundary conditions and the same background) to obtain the transmission spectra. By comparing the reflection and transmission with and without the cloak, we can determine at which frequencies cloaking effect may be achieved. Second, by monitoring animated fields (both phase and amplitude) at frequencies determined in the first step, we can further determine whether the cloaking effect indeed takes place at those frequencies", + " The above three steps corroborate one another and thus can convincingly verify the designed cloak. The simulations were carried out using the time domain solver in CST Microwave Studio. To mimic an infinitely long copper cylinder, the two boundaries perpendicular to axis are set to be PEC (perfect electric conductor) boundaries while the two lateral boundaries perpendicular to axis is Open boundaries with a reference plane 50 mm away from the cloak. TE-polarized (the electric vector is in parallel with the cylinder axis) waves are incident along direction, as shown in Fig. 6. Two wave-ports are set on the two open boundaries along the direction with a reference plane 200 mm away from the cloak. Fig. 7(a) and (b) shows, respectively, the reflection and transmission of the copper cylinder mm with and without the cloak under the same simulation setup. Compared with the case without the cloak, the transmission is greatly enhanced while the reflection reduced around 3.63 GHz in the case with the cloak. This gives a clear indication that the cloaking effect is very likely to take place around 3", + " In contrast, for a bare copper cylinder without the cloak, strong reflection occurs. An obvious shadowed region appears and the phase-fronts are distorted in the back region. By virtue of the cloak, EM energy flows around the copper cylinder, just as water flows around stones, as shown in Fig. 8(c). Now it can be definitely determined that quite excellent cloaking effect is achieved at 3.63 GHz.What should bementioned here is that the fields and power flows were obtained by simulating the realistic cloak structure in Fig. 6, not by assigning material parameter profiles as in many previous works [4]\u2013[8]. To further appraise the cloaking effect at 3.63 GHz, we computed the bi-static scattering widths of an infinitely long copper cylinder with and without the cloak, as shown in Fig. 9. With the help of the cloak, the scattering widths in all directions can be reduced significantly by more than 15 dB on average. This convincingly confirms the excellent cloaking effect at 3.63 GHz. Note the cloak is quite thin, only less than 1/40 the operating wavelength and 1/30 the cloaking diameter" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001472_9781118354179.auto089-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001472_9781118354179.auto089-Figure3-1.png", + "caption": "Figure 3. Clutch integral to the gearbox assembly. (Reproduced from Nunney, 1998. \u00a9 Elsevier/Butterworth Heinemann.)", + "texts": [ + "g., wheel locking when braking). Main transmission clutch is typically attached to the engine flywheel, as indicated in Figure 2 (Reimpell, Stoll, and Betzler, 2001), which represents by far the most common installation. However, other installations are possible, when the engine and gearbox are split, usually to achieve better weight distribution between vehicle axles. In such cases, the clutch can be still attached to the engine flywheel or designed integral to the gearbox, the latter shown in Figure 3 (Nunney, 1998). Having in mind previous consideration, the most commonly used type of the main transmission clutch will be considered, having the following characteristics: \u2022 Friction type: Torque transfer is based on dry friction principle. Wet clutches and hydrodynamic couplings are different designs studied separately; \u2022 \u201cDisc type\u201d: Friction surfaces are \u201cflat rings.\u201d Cylindrical (drum type) and conical friction surfaces are nowadays very rare. \u2022 \u201cNormally\u201d engaged type: When there is no outside intervention, the clutch will be fully engaged, transmitting the torque" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002378_tasc.2010.2092745-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002378_tasc.2010.2092745-Figure3-1.png", + "caption": "Fig. 3. Cross-sectional drawing of rotor in magnetization.", + "texts": [ + " Pulsed field magnetization will be used for magnetization of rotor. The stator is on the bottom of the rig. The rotor is in the middle of 2 magnetizing coils for PFM processing. After PFM, it will move rotor down into stator for motor operation. The entire rig will be immerged into liquid nitrogen (77 K) during the experiment. As shown in Fig. 2, a pair of magnetizing copper coils is mounted on the motor rig. In the middle of coils, it is the rotor with 15 columns of surface mounted YBCO bulks. The crosssectional drawing of rotor in magnetization is shown in Fig. 3. It will get 2 same poles on rotor in each magnetization processing. The rotor can be rotated by a scaled hand wheel and also can be locked/unlocked by a screwed clamp installed on the shaft. During PFM processing, the rotor will be locked on to prevent it from rotating due to pulsed magnetic force. After the first PFM processing, the rotor will be unlocked and rotated in 90 degrees to obtain the other 2 poles magnetized. A hall probe which locates in the center of the magnetizing coil is pasted in the gap (3 mm) between rotor and a coil", + " 4 magnetizing coils which are successively applied with a pulsed current each pair are used in simulation. Fig. 6 shows simulation results. It only applies one pulse in simulation. From streamlines distribution as shown in Fig. 6, it indicates trapped magnetic field on 4 poles by superconducting bulks distributes asymmetrically. This is because To investigate the distribution of rotor trapped field, it had totally recorded 36 positions by rotating rotor in 10 degrees for each. The rotation angle is as shown in Fig. 3. Fig. 7 shows the applied field and trapped field in original (0 degree) position. The peak of applied field (measured without YBCO bulks) is 850 mT and pulse width is 0.75 s approximately. The trapped field shown in Fig. 7(b) was measured after 10 s. The trapped field increases gradually as number of pulses increasing, and the magnitude of increase is decreasing. To obtain maximum trapped field, the pulse is generated repeatedly until a fixed trapped field acquired. The trapped field raises from initial value of 195 mT to 375 mT at the 19th pulse" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000061_s12541-014-0447-1-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000061_s12541-014-0447-1-Figure8-1.png", + "caption": "Fig. 8 Result of the analysis for stress concentration factor", + "texts": [ + " (4) The gradient of the test result and the S-N curve of the virgin material differed significantly. This result did not consider the stress concentration by the notch. Therefore, the notch effect was considered as follows. Generally, the stress concentration factor Kt is defined as the ratio of maximum stress (\u03c3max) and nominal stress (\u03c3nom) as in Eq. (5). (5) The stress concentration factor was obtained by FEA by applying the notch effect to the caulked part and the value obtained was 3.06, as shown in Fig. 8. The fatigue notch factor Kf was derived as follows; (6) In Eq. (6), a is a material constant, with a value of 0.016, since r was 1.6 mm in the FEA. A value for Eq. (7) can be obtained from the experimental curve by Heywood.7 The S-N curve was derived by Eqs. (8) and (9). (7) (8) (9) Sn 1000 0.75 Su SRS\u2013( )\u22c5= Sn 1000000 0.5 Su SRS\u2013( )CLCDCS\u22c5= Sa Sn ---- Sm Su -----+ 1= Kt \u03c3max \u03c3nom ----------= Kf 1 Kt 1\u2013 1 a r --+ --------------+= K\u2032f 1\u2013 Kf 1\u2013 ------------- 0.1\u2248 Sn 1000 0.75 Su SRS\u2013( )\u22c5 K\u2032f ----------------------------------= Sn 1000000 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001568_s13246-016-0502-6-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001568_s13246-016-0502-6-Figure2-1.png", + "caption": "Fig. 2 A view of the devised jack for hip and knee joint movement", + "texts": [ + " CATIA (an acronym of Computer Aided Three-dimensional Interactive Application) is a multi-platform computer-aided design (CAD)/computer-aided manufacturing (CAM)/computer aided engineering (CAE). CAE completed with ABAQUSE software. All parts of exoskeleton were drawn in and assembled in CATIA (v5r20) software. 1 3 After applying the forces on exoskeleton model, joints torque was taken from software. Figure\u00a01 shows a view of the designed exoskeleton in CATIA Software and its walking model in MatLab. This exoskeleton had two screw jacks for the hip and knee joint movement. The thigh (32\u00a0 cm) and knee (20\u00a0 cm) jacks were located on the back and front side respectively, (Fig.\u00a02). Due to this fact that the generated torque span in the hip and knee joints can be both positive or negative, two-side jacks were used. The designed exoskeleton had four degree of freedom; two degree of freedom for hip and knee (extension\u2013lexion) and two 1 3 degree of freedom for ankle (extension\u2013lexion and abduction\u2013adduction). The structure of the designed exoskeleton was made of aluminum with the mass of 5\u00a0kg. This frame had a male and female connection to adjust the length of thigh and calf in a way that female aluminum proile was designed with dimensions of 35 \u00d7 5 \u00d7 3 cm3 for the femur and 25 \u00d7 5 \u00d7 3 cm3 for the tibia" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003952_978-3-642-33457-3_7-Figure7.6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003952_978-3-642-33457-3_7-Figure7.6-1.png", + "caption": "Fig. 7.6 Setup with the aerostatic bearings\u2014exploded view [10]", + "texts": [], + "surrounding_texts": [ + "At the time of the motor design many of the application requirements were still missing. Namely, the project group at TU Delft lacked practical experience with micromilling and information on torque requirements for high-speed micromilling was rather scarce. Additionally, the work on the motor design started long before models presented in this thesis were developed. Therefore, most of the actual requirements were imposed by the designers themselves and not directly by the prospective application. However, the goal of the design was to offer a good proof of concept rather than a definite solution for micromilling spindles. Taking into consideration available data and technological limits in an academic environment, the designers looked for sensible and adequate requirements for the new spindle and the spindle motor in particular. They are presented as follows: 1. The motor needs to fit both proposed bearings and its stator should have a good thermal contact with the (bearings\u2019) housing. 2. It should be possible for both setups to be manufactured using technology of a standard/university workshop and of-the-shelf components. In other words, miniaturization of the rotor is limited by technological capabilities in the university environment. This is particularly important in the example of AMB whose extreme miniaturization would require utilization of advanced processing techniques and/or components (miniature sensors, coils, etc.). 3. In light of the previous requirement, it was concluded that a rotational speed of 200.000 rpm of a disc-shaped rotor in proposed bearings would be technically achievable (see also Sect. 7.3.1). Therefore, that speed was set as the speed requirement for the motor. 4. No data on required torque for micromilling were available. However, available estimations of cutting forces for milling with sub-millimeter tools suggested rather small load forces, in order of a few newtons and lower [1, 11]. It was, thus, expected that the load torque would be significantly smaller then the drag resulting from windage and eddy-currents. Taking into account worst predictions of airfriction loss (Fig. 3.21 in Sect. 3.6.3) it was concluded that 200 W of power would be certainly sufficient for operation at the required maximum speed. 5. For a high-frequency-operating machine which would be completely enclosed in magnetic bearings, minimization of frequency-dependant losses was extremely important. Besides, due to inability to reliably model air-friction loss and losses in the permanent magnet (Sects. 3.6.3 and 3.6.4), thermal model was not developed, thus, mitigation of losses was inevitable for a safe design. Therefore, minimum loss (= maximum efficiency) was taken as a decisive criterion for both component choice and electromagnetic design. 6. As pointed out in Sect. 3.5, negative motor stiffness, which results from the unbalanced magnetic force, must be, at least, an order of magnitude lower than the stiffness of the radial bearings and the unbalanced force must be lower than the bearing force capacity. The stiffness limit was critical for the case of AMB (estimated in order of 105 N/m) while the force limit was critical for the aerostatic bearings. 7. Structural robustness of the rotor was an equally important requirement for the design. Proper retaining of the magnet was crucial, particularly for a high-speed rotor with a high diameter to length ratio. All these requirements affected the design of the motor whose conceptual design is depicted in drawings in Fig. 7.7 The motor concept is explained in the rest of this section (Fig. 7.8). A laminated, slotless stator core has protrusions corresponding to the axial direction for good thermal contact with the housing. Advantages of slotless machines for very-high-speed operation were discussed in Chap. 2. Exclusion of stator teeth removes slotting-effect harmonics from the PM field while, at the same time, reduces impact of armature-field harmonics in the PM rotor. As a whole, a slotless motor is prone to be more efficient and less susceptible to rotor overheating than its slotted counterpart. Conductors are wound toroidally over the core, thus, dispensing with, for this case, unavoidably long end windings. The windings are non-overlapping, i.e. each phase winding is uniformly wound over two 60\u25e6-sections of the stator circumference. A plastic-bonded magnet of the injection molded type is applied onto the incised part of the rotor disc (see Fig. 7.10). An incision is previously made in the shaft for a better fit of the magnet. The injection molded magnet contains very small magnet particles that are blended with a plastic binder\u2014PPS. After applying this mixture onto the shaft at a very high temperature, the magnet will apply a stress on the shaft during the cooling in the mould. Such a magnet is very resistive to eddy-current losses. At the same time, low remanent flux density of the magnet, as a result of the plastic binder overtaking a great portion of the magnet volume, is quite adequate for a very-high-speed machine (see Sect. 2.4). The magnet is diametrically magnetized providing a perfectly sinusoidal back emf. Finally, in order to sustain very strong centrifugal force at high speeds and ensure transfer of torque in the rotor throughout the whole range of speeds, the magnet needs to be contained in a non-magnetic enclosure/sleeve. A non-conductive sleeve has been conceived as a combination of glass and carbon fiber: details on the sleeve design are presented in Sect. 7.6. In the rest of the section the materials used for the motor parts\u2014stator core, conductors, permanent magnet and sleeve\u2014are discussed." + ] + }, + { + "image_filename": "designv6_24_0001924_sp.j.2096-5796.2019.0016-Figure10-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001924_sp.j.2096-5796.2019.0016-Figure10-1.png", + "caption": "Figure 10 (a) Six-dimensional F / T sensor with a novel structure; (b) Circular spoked structure as an elastic spherical sliding[32].", + "texts": [ + " In recent years, many researchers presented some novel mechanical structures of elastic beam on the basis of the classical structures to improve the performance of multi-dimensional force sensor. A large-range six-axis force sensor made by machining dumbbell grooves in cross beams is developed by Changchun Institute of Optics, which has high sensitivity while ensuring a large measurement range[31]. Schematic diagram of the elastic beams of the sensor is depicted in Figure 9. Mastinu et al. designed a six-axis F /T sensor with a novel mechanical structure, seen in Figure 10a. The sensing element of this sensor is a quasi-statically determined three spoke structure constrained by virtue of elastic sliding spherical joints, seen in Figure 10b, which is designed to avoid friction[32]. Experimental results demonstrated that the sensor have a good performance in linearity, crosstalk and dynamic behavior. Liang et al. developed a novel miniature fourdimensional force sensor with elastic elements consisted of circular diaphragm and cantilever beam, seen in Figure 11[33]. The sensor comprises of a sensor tip, an upper cover, a base frame and elastic elements. Hu et al.presented a novel elastic element design method of a six-dimensional F/T sensor with a floating beam that changing the floating beam to H-beam to increase the stiffness of the sensor and improve the dynamic performance[34]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001801_j.triboint.2016.11.027-Figure16-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001801_j.triboint.2016.11.027-Figure16-1.png", + "caption": "Fig. 16. Static load-deflection test results for NSFB with 47 springs when the rotor operates at 0 rpm, 15.6 krpm, and 24 krpm.", + "texts": [ + " The expression of Ai is given as A A A A A= ( + + )/i 1 2 3 4 (A6) where A G E \u03c0 a p \u03b8 \u03c0 a p \u03c0 a p \u03c0 a \u03b8 \u03c0 a p \u03b8 \u03c0 a p \u03b8 \u03c0 a p \u03b8 \u03c0 a p \u03b8 \u03c0 a p \u03b8 \u03c0 a p \u03b8 \u03c0 a p \u03b8 \u03c0 a p \u03b8 \u03c0 a p \u03b8 \u03b8 \u03c0 a p \u03b8 \u03b8 = / (5184 + 1296 sin( ) + 972 + 4536 \u2212 5184 cos(2 ) + 4464 sin( ) + 1728 sin( ) \u2212 972 cos(2 ) \u2212 4536 cos(2 ) \u2212 144 sin(3 ) + 324 sin(2 ) + 162 sin(2 ) \u2212 64 sin(3 ) \u2212 324 sin(2 ) \u2212 648 sin(2 )) J I s s s s s s s s s s s s s 1 7 6 6 1 3 2 4 5 4 2 7 6 1 2 2 4 1 4 4 2 1 3 2 4 1 5 4 2 1 2 2 4 1 6 4 2 1 4 2 4 1 4 4 2 1 3 2 4 1 1 5 4 2 1 1 (A7) A E G \u03c0 a p \u03c0 a p \u03b8 \u03c0 a p \u03b8 \u03c0 a p \u03b8 \u03c0 a p \u03b8 \u03c0 a p \u03b8 \u03b8= / (\u2212648 + 4608 sin( ) + 648 cos(2 ) + 512 sin(3 ) \u2212 324 sin(2 ) + 648 sin(2 ))I J s s s s s s2 5 4 2 4 4 2 1 5 4 2 1 4 4 2 1 6 4 2 1 5 4 2 1 1 (A8) A \u03c0 a \u03c0 a p \u03c0 a p \u03c0 a \u03b8 \u03c0 a p \u03b8 \u03c0 a p \u03b8 \u03c0 a p \u03b8 \u03c0 a p \u03b8 \u03c0 a p \u03b8 \u03c0 a p \u03b8 \u03c0 a p \u03b8 \u03c0 a p \u03b8 \u03b8 = 5184 \u2212 324 + 1296 \u2212 5184 cos(2 ) + 5904 sin( ) + 14400 sin( ) + 324 cos(2 ) \u2212 1296 cos(2 ) + 144 sin(3 ) \u2212 162 sin(2 ) \u2212 448 sin(3 ) + 324 sin(2 ) s s s s s s s s s s 3 7 6 3 2 4 5 4 2 7 6 1 2 2 4 1 4 4 2 1 3 2 4 1 5 4 2 1 2 2 4 1 4 2 4 1 4 4 2 1 3 2 4 1 1 (A9) 4 2 2 2 (1/2) 2 2 2 2 2 (A10) where \u03b81 is the location of friction force acting on the left side of the spring. See Appendix Fig. 16. See Appendix Table 4. Fig. 16 shows the load-deflection test results for the NSFB with 47 springs when the rotor operates at 0 krpm, 15.6 krpm, and 24 krpm. The linear stiffness, displacement amplitude, area of hysteresis loop, and loss factor are extracted from the load-deflection curves and are listed in Table 4. The linear stiffness at the operating speed of 0 rpm is the largest (0.2856 MN/m), whereas the value at 15.6 krpm is the smallest (0.2554 MN/m). The linear stiffness increases from 0.2554 MN/m to 0.2597 MN/m as the speed increases from 15" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001824_022050-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001824_022050-Figure4-1.png", + "caption": "Figure 4. Equivalent stress. Figure 5. Total deformation.", + "texts": [ + " Figure 3 shows all the force acting on the upright, simultaneously, so that the upright can be tested for the worst case scenario. . During braking, throttling and steering simultaneously the upright experiences maximum forces, so the upright is designed for a condition in which the car corners while braking and also encounters a bump. So all the before mentioned forces act simultaneously on the upright. The maximum stress and total deformation were obtained as 57.95N and 0.05mm respectively. Figure 4 and figure 5 depicts the stress flow and deformation flow path respectively. The specialized 3D printer called Mark forged mark one was used, it uses two nozzle deposition techniques to deposit the thermoplastic and the additives on the base layer by layer leading to final dimensions of the product. Both the materials, thermoplastic and additives, are fed through separate 3rd International Conference on Advances in Mechanical Engineering (ICAME 2020) IOP Conf. Series: Materials Science and Engineering 912 (2020) 022050 IOP Publishing doi:10" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001360_aero.2017.7943762-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001360_aero.2017.7943762-Figure6-1.png", + "caption": "Figure 6. Camera FOV for Galaxy 28", + "texts": [ + " As with any data link, it was vital to carefully consider the required SNR and the available downlink SNR (C/No) at the earth station during the camera mission. Conventional link budgets were used to analyze these. The required SNR was measured prior to the launch by combining noise with the simulated downlink at the receiver to determine the link SNR threshold. Later missions measured BER curves in conjunction with this test. During mission planning, it was necessary to consider what will be visible within the camera Field Of View (FOV), as well as where the sun will be for specific video capture events. Figure 6 shows a model used to consider what would be captured by a given camera. The camera FOV is the cone with the point at the center top of the figure, looking down towards the 2 main reflectors. Image resolution is approximately 720 x 480 pixels, after being digitized from the NTSC color camera. Individual video frames were compressed, encoded and transmitted with no differential encoding (no delta frames). This eliminated error propagation in video (i.e., loss of one frame did not impact adjacent low error frames)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001388_s1644-9665(12)60163-0-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001388_s1644-9665(12)60163-0-Figure6-1.png", + "caption": "Fig. 6. Scheme of sheet hydroforming (a) and examples of products (b) [13]", + "texts": [ + " Hydroforming of sheet material (SHF) is up to now mainly used for small batch production due to a comparatively high cycle time. Furthermore, sheet hydroforming requires higher clamping forces than tube hydroforming, causing more cost-intensive presses. However, advances in process and press technology, e.g. [3\u20136], increasingly contribute to a wider industrial application of sheet hydroforming, in particular for the flexible manufacture of small batch sizes. One of the first examples for an industrial application is the sheet hydroforming of roofs for luxury class cars, Figure 5. Figure 6 shows the scheme of a sheet hydroforming. When the punch pushes the sheet metal into the die cavity, within which oil or other liquids are contained, pressure pi that can press the sheet metal tightly onto the punch will be generated. At the same time, the liquid in the die cavity will flow out between the upper surface of the die and the sheet metal, what results in reduction of frictional forces. By this process, the limit drawing ratio value of sheet metal can be increased. The liquid can be used as a punch, a draw die or an assisting way to improve sheet formability. Actually, almost all of the materials used in conventional stamping can be used in sheet hydroforming. Depending on the different means, the liquid pressure in the die cavity is from around 30 to 150 MPa, but the usage of 200 MPa has also been reported [13, 25]. Roof for luxury class car (Figure 5), deep and partially conical cup (Figure 7) or cups with complicated geometries of bottoms (Figure 6) are examples of parts made by sheet hydroforming. It is well known that formability of the lightweight materials usually increases at elevated temperature levels [27\u201328].Warm forming technology with selective heating enables manufacturing of lightweight parts by utilizing the increased formability at elevated temperature [28]. However, it is quite difficult to determine optimal temperature distributions in tooling [29]. Warm hydroforming technology for lightweight materials is currently being developed to achieve reduced number of manufacturing steps and part consolidation" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002516_adfm.201807082-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002516_adfm.201807082-Figure3-1.png", + "caption": "Figure 3. Illustration of the self-assembly shaping (SAS) process. a) Stretch-Body with two predesigned contact points was joined together. Contact points are adhered well due to postprinting UV irradiation. Sintered-Body retains the self-assembled design. b) Two surfaces of Stretch-Body predesigned contact points before joining. c) The interface of the two contact points after assembly. d) Dense Sintered-Body obtained after sintering.", + "texts": [ + " KGaA, Weinheim possible to fabricate large mesh samples with similar fine features of minimum layer height (see Figure S12, Supporting Information). Besides its stretchability, a Stretch-Body is also able to undergo a self-assembly process (as shown in Figure 1c previously) by binding to itself due to its sticking properties. This is a unique feature that differentiates itself from the Flex-Body; the StretchBody is capable of adhering to itself (as shown in Figure 1c previously). The adhesiveness of the Stretch-Body is advantageous in realizing the SAS process.[24] The typical SAS process is elaborated further in Figure 3. As shown in Figure 3a, a Stretch-Body with mesh morphology was twisted 360\u00b0 and subsequently connected at the two ends with a predesigned contact (see also Movie S5a,b, Supporting Information). The stretchability, as well as the adhesiveness of the StretchBody (see Figure S13a,b, Supporting Information), can be mainly attributed to the presence of the 2-HEA monomer. To better understand the SAS process, a simple model is proposed in Figure 3b\u2013d. Figure 3b depicts the two predesigned contact area prior to joining. After the printing process, a complete 100% acrylate double bond conversion (DBC) is usually not achievable, and the resultant photocured sample still retains its tackiness. The tackiness of the sample can be attributed to the following factors: (i) the low DBC after robocasting and UV exposure, (ii) the presence of oxygen-containing functional group in the plasticizer (e.g., ethylene glycol or diethylene glycol) and (iii) the presence of dangling oxygen-containing functional group (hydroxyls ROH, ether ROR\u2032 as well as ester RC(O)OR\u2032) in the cured backbone. When the two predesigned contacts were brought together, as depicted in Figure 3c, the presence of the oxygen-containing functional groups such as hydroxyls-, ether-, and ester-containing functional groups from 2-HEA and EG/DEG plasticizer allows hydrogen bonding interactions between hydroxyls, hydroxyls with ether linkages or hydroxyls with the ester linkages. Such hydrogen bonding interactions provide immediate bonding of the Stretch-Body after contact for further crosslinking during after-printing UV irradiation, improving the overall DBC. The resultant structure is stable which allow the shape and the joint retention thereafter, even after sintering as depicted in Figure 3d (see also Figure S14 and Movie S6, Supporting Information). In this manner, the SAS process potentially allows the fabrication of geometrically complex structure by a modularassembly through the aforementioned postprinting processing. While cMAS and SAS have demonstrated tremendous potential to fabricate geometrically complex structures, cMAS and SAS rely on either the incorporation of ceramic weight or the self-assembly process; a process which has a low degree of Adv. Funct. Mater. 2019, 1807082 1807082 (7 of 12) \u00a9 2019 WILEY-VCH Verlag GmbH & Co" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001943_0890-6955(95)00104-2-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001943_0890-6955(95)00104-2-Figure5-1.png", + "caption": "Fig. 5. Rotary angle head attachment.", + "texts": [ + " Moreover, post-multiplying \u00b0A4_ by ~A,, the configuration of the cutting tool with respect to the spindle, one obtains the ability function of this 4-axis vertical machine tool. NC Data Generation for 4-axis Machine Tools 347 Note that since equation (14) can never equal the desired curer location matrix, equation (13), it is impossible to machine this screw without modification of this vertical machine tool. Also since element (3,3) of equation (13) is constant, one possible modification to this machine tool is the installation of a rotary angle head attachment (Fig. 5), which would allow for the cutting tool to be positioned in the horizontal plane and then rotated an offset angle \"y with respect to its rotation axis. In other words, with the attachment the position and orientation of the cutting tool can be described by: 4_A, = Rot(y.90o)_Rot(x,3,)T_rans(0,0,d) = C3, -S~/ -dS_3, ] = - - 1 0 0 - 0 0 0 (15) where _d is the length of the cutting tool, measured from the origin of frame (xyz)t to the cutter's end. Neglecting the orientation of the xt and y, axes, the curer location matrix with respect to frame (xyz)o is given by: Ii ? C~/$_e i ? - C1tCO t \u00b0A,=\u00b0A, -~A_, - - - ? - S ' y 0 0 (a, + b3 + _dC v)SO,-(b, + g,)CO~ - - ( a 4 + b_ 3 .4- dC3')COl-(b4 + al)SO_l bt + b2-dS',/ 06) 348 Psang Dain Lin and Ming Far Lee By equating the elements of equations (13) and (16) the desired twisted angle y of the attachment and the four link variables of this machine tool (__O1, ___2, b2, b3 which are listed in the Appendix A by equations (A8)-(A12)) can be determined. Referring to Fig. 5 again, when the tip of the cutter coincides with the origin of workpiece frame (xYZ)o the position and orientation of tool frame (xyz)t with respect to the workpiece frame is given by: ii0 1 ~ ? - C z \u00b0A-' - ? -S'y (17) 0 0 1 By equating the corresponding elements of equations (16) and (17), one obtains the following expressions of origin of workpiece frame in terms of joint variables of machine tool: _01o=0 (18) _b2o = - b l + dS'y (19) bso = -a_4-_dC7 (20) b4o = - a l (21) Referring to Fig. 4, the desired NC position commands of A, X, Y, and Z axes, expressed with respect to screw frame, are now given by the corresponding differences of equations (A9)-(AI2) and (18)-(21): A = __01-_01o (22) X = -(_bz-b2o) (23) NC Data Generation for 4-axis Machine Tools 349 y = - ( b 3 - b 3 o ) (24) Z = -(b4-b_4o) (25) The values of Y and Z are constants and are referred to as positional constraints" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000083_ir-03-2018-0042-Figure9-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000083_ir-03-2018-0042-Figure9-1.png", + "caption": "Figure 9 Seating mechanism described in the two modes", + "texts": [ + " This allows for operating in a stationary position as well. Independent suspensions are used to pass obstacles, while the wearable vehicle moves over the obstacle, the right rear shock compresses, while the others adjust for equilibrium, ensuring high stability. The seat axis of rotation is fixed on the trunk and the other side is fixed on the thigh link or in the rear wheels mechanism. This motion is completed automatically by using two securing mechanisms to adapt the two modes with a simple design as shown in Figure 9. For instance, if the wearer switches into walking mode, the rear wheels mechanism is allowed to oscillate forward and backward to stabilize the system.While, if the wearer switches into fast motion mode, the rear wheels mechanism will be fixed in the thigh link of the system. The secure mechanism is actuated using two linear actuators which facilitate the transition between two positions without effort on the wearer\u2019s part. It is worth to mention that the wearer can unlock the secure mechanism manually in case of electrical power lack, meaning that the system will not lock the person\u2019s movement" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001982_s0268-0033(02)00043-8-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001982_s0268-0033(02)00043-8-Figure1-1.png", + "caption": "Fig. 1. Cylindrical handle used in the experiment: (1) force sensor, (2) button, (3) cylindrical part made of nylon, (4) groove for wiring, (5) aluminum tube, (6) nut and (7) rubber disk.", + "texts": [ + " The effects of all the other variables on the mean values were minimized by randomizing the order of the bouts of the experiment with respect to the independent variables. Ten healthy volunteers (five men and five women) participated in the experiment. All the subjects were right-handed. The forces acting on the fingertips in power grip were measured by force sensors mounted on cylindrical handles. The three handles of the experiment, with diameters 4, 5 and 6 cm, and length about 10 cm, consisted of six separate parts allowing relative rotation about a common axis (Fig. 1). The positions of the parts were secured with a bolt running along the axis. Radial holes of 1 cm depth and 1.2 cm diameter were drilled on four of the parts. Load cells (Kyowa LM-5KA, Tokyo, Japan) and nylon buttons were mounted onto these holes, so that their upper surfaces were slightly above the level of the handle surface. The wiring of the cells was placed on grooves milled on the parts to allow a natural grip. Load cells were connected through an amplifier to a PC computer. Measuring hardware and software (Dasylab 5, Datalog GmbH, M\u20aconchengladbach, Germany) were used to generate feedback to the subject, to handle the different stages of the measurement, and to store the data at a frequency of 1000 Hz for later processing" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001826_870305-Figure20-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001826_870305-Figure20-1.png", + "caption": "Fig. 20", + "texts": [], + "surrounding_texts": [ + "10\no Heater and air conditiorer\n870305\nHaterials:\nTrends:\n- Fousing parts: PP TF - Blower wheel: PON - Thin-layer technique for - Flaps hard/soft\n- Flilps: - Controls:\nhousing parts\nNetal/PU foam ABS", + "870305\n11\nHaterials:\nTrends:\n- Housing parts: PP TF - Blower wheel: POM - Thin-layer technique for - Flaps hard/soft\n- Flaps: - Controls:\nhousing parts\nHetal/PU foam ABS", + "12\n870305\nA wide range of materials and manufacturing processes is available for this field of application. Apart from the flat, difficult to mold, hard-fiber boards, use is made of thermoplastics, such as PP and ABS, as substrates. There is also the large group of moldable substrate materials (see Table 4). A detectable trend is that there is a widening out of the materials and processes with which it is possible to obtain maximum variability and integration; this means\nfreedom of design degree of molding hard/soft combination variation possibilities for the decorative material." + ] + }, + { + "image_filename": "designv6_24_0001721_ip-h-2.1990.0051-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001721_ip-h-2.1990.0051-Figure3-1.png", + "caption": "Fig. 3 Current basis functions a Shaded regions show form of the 6 functions used for the simple tripole b Tripole loop: variable I (eqns. 2 and 3) measured clockwise from point 0 along mean circumferential path indicated by arrows", + "texts": [ + " The integral equation, obtained by using electromagnetic boundary conditions to relate these fields to the currents induced in the conducting regions of the arrays, was reduced to a set of linear equations by expanding the currents as a set of basis functions. The equations were solved in turn for the coefficients of these functions by matrix inversion. Six sinusoidal basis functions were used for the simple tripole. Three had half periodicities equal to the length of each arm and each of the remaining three had twice this periodicity and extended over two arms. Their distribution over the tripoles is sketched in Fig. 3a. The current was constant across the width, W , of the conductors. In the case of the tripole loop, the expansion of the induced currents required, including an important constant term, six cosinusoidal functions of the form ~ , , / ( 2 / WL) cos (2mrr l l~) m = 0-5 (2) IEE PROCEEDINGS, Vol. 137, Pt. H , No. 5 , OCTOBER 1990 and five sinusoidal terms C,4(2 /WL) sin (2n7d/L) where n = 1-5 (3) where L was the total length of the loop and I was a parameter measuring he distance around-the loop from an origin taken at the point 0 in Fig. 3b; the mid end of the y-directed arm. With these basis functions, the computed values of C , and C, converged when 441 or more Floquet modes were used to expand the local fields. 2.3 Measurements To assess the reliability of the computed results a series of arrays with salient values of the lattice parameters have been constructed. The experimental arrays of tripoles measured 20 cm square and were illuminated by collimated beams produced by a lens 60 cm in diameter. A few measurements made at 80\" incidence required much larger surfaces with dimensions of 120 cm x 60 cm" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000188_s1526-6125(05)70089-x-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000188_s1526-6125(05)70089-x-Figure2-1.png", + "caption": "Figure 2 Schematic Axial Cross Section of Shear Forming Process", + "texts": [ + " In Figure 1a, the experimental setup used in this study is shown; in Figure 1b, the FEM simulation of a negative incremental forming operation with a spherical punch is shown. In positive forming, the central part of the workpiece is supported by a fixed counterpunch (or mandrel) and the tool-workpiece interface is located on the convex side of the shape to be formed. The sheet can be either fixed at the periphery (Park and Kim 2003) by a blankholder or it can be free. The latter case can be considered a direct evolution of the conventional shear spinning process (Wong, Dean, and Lin 2003). Some important variables of the negative forming process are shown in Figure 2, where an axial cross section of a formed part is schematically drawn. The vertical and horizontal axes, z\u0302 and \u03c1\u0302 , respectively, are shown, along with the vertical part slope, , the initial and instantaneous thickness, t0 and t, respectively, and the longitudinal and thickness directions, l\u0302 and t\u0302 , respectively. If the part is axisymmetric, the horizontal position of a point on the blank inner surface is equal to the Journal of Manufacturing Processes Vol. 7/No. 2 2005 zontal radius of curvature of the part at that point", + " In simple shear conditions, a horizontal and planar sheet metal blank with initial thickness t0 is deformed to a final conical geometry with vertical slope and final thickness equal to: 0 sin( )t t= \u22c5 \u03b1 (1) Equation (1) is called the sin law of thickness (Wong, Dean, and Lin 2003). The true strain in the thickness direction t\u0302 is therefore: ( ) 0 ln ln sint t t \u23a1 \u23a4\u03b5 = = \u03b1\u23a3 \u23a6 (2) The planar strains are l, which is measured in the longitudinal direction, l\u0302 , and , which is measured in the tangential (hoop) direction, not represented in Figure 2. The hoop strain, , is the minor planar strain, and unless the deformation is locally perturbed by a small horizontal curvature radius or by relevant friction forces, its theoretical value may be assumed to be = 0. Therefore, the longitudinal strain, l, is the major planar strain and is equal to l = \u2013 t. The formability of sheet metals when plastically deformed by dieless incremental forming has often been investigated by researchers (e.g., Shim and Park 2001). Although incremental forming is a relatively new process, some common knowledge is already available", + " In positive incremental forming without a blankholder, the process is not only limited by the possible occurrence of cracks but by the possible occurrence of wrinkling in the flange. The long time required for forming even small-sized parts is one of the main disadvantages of the process. Therefore, a robust knowledge of the forming limits, in respect to fz, would allow the user to set the feed rate at the maximum possible value. The effect of the part curvature radius, (measured onto a horizontal plane, Figure 2), has been Journal of Manufacturing Processes Vol. 7/No. 2 2005 seldom investigated in explicit terms. However, it is known that in negative forming the shear strain is no longer prevailing at sharp corners (that is, for very small values of ), where the deformation is tensile on both principal axes and fracture more likely appears (Kim and Yang 2000; Filice, Fratini, and Pantano 2001). When the curvature radius, , is large compared to the punch radius, r, it does not have a direct effect on formability, but rather it has a more complex influence on the process mechanics (Strano, Ruggiero, and Carrino 2004)", + " Therefore, any previously known FLD model cannot be used for prediction of fracture in incremental forming. \u2022 In incremental forming, the major and minor planar strains, and l, respectively, at failure are measured for building the FLDs. The underlying assumption is that the planar strains are principal (Bunge et al. 2000). Unfortunately, this is not true if the sheet deformation is by simple shear. In this case, the actual principal axes do not rotate during the forming process and are oriented at 45 degrees with the vertical and horizontal directions (Segal 2002). The principal axes are shown in Figure 2, labeled as 1\u0302 and 2\u0302 . \u2022 The FLDs can be used as an engineering design tool, provided that the deformation of the final sheet metal product can be accurately predicted, e.g., by an FEM analysis. However, the finite element simulation of incremental forming processes is a very lengthy task (Muresan et al. 2005). A tool for failure prediction, based on technological variables (such as the feed rate) rather than strains, would be a very useful alternative. \u2022 The FLDs exhibit a strain path dependency" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003121_12.826223-Figure13-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003121_12.826223-Figure13-1.png", + "caption": "Figure 13: Circumferential E-Beam Weld of the Tube to the Pads", + "texts": [ + " The weld fixture with the strut tube and pads is then placed into an electronic beam welding vacuum chamber. The door is closed and the titanium strut tube is e-beam spot welded to the pads at both ends. The spot weld is strong enough to Proc. of SPIE Vol. 7439 74391B-7 Downloaded From: http://proceedings.spiedigitallibrary.org/ on 06/18/2016 Terms of Use: http://spiedigitallibrary.org/ss/TermsOfUse.aspx hold the parts together without the fixture. The fixture is removed and the spot welded strut tube assemblies are loaded back into the e-beam weld chamber for the final weld (see figure 13). The final e-beam weld goes completely around the circumference of the tube and fuses the two similar materials together. No filler is used so the e-beam weld is the same strength as the bulk Titanium material. The joint between the tube and the pad is designed specifically to accommodate the e-beam weld (see figure 14). The depth of the weld, intensity of the e-beam and the speed of rotation is determined by the required penetration, cross sectional area, and material being welded. Prior to beginning welding on the strut tube assembly, a number of sample pieces (with similar features) are welded, sectioned, and examined to confirm that the e-beam setup will complete the weld properly (see figure 15)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000489_tgrs.2021.3051727-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000489_tgrs.2021.3051727-Figure6-1.png", + "caption": "Fig. 6. It is the LVLH rotating coordinate system of primary satellite. In other words, the LVLH is centered on the primary satellite.", + "texts": [ + " More than a decade ago, scholars invented several classical satellite formations for different M-SAR missions, including Dual-Helix [4], [5], Pendulum [1], [6], and Cartwheel formations [7]\u2013[10]. In Cartwheel formation, a reference satellite flies on a circular orbital plane with several auxiliary satellites rotating around it in the same plane [7]\u2013[10]. This type of formation provides a varying along-track baseline. Unfortunately, the horizontal baseline of Cartwheel formation is almost zero, i.e., the component of baselines along the Z -axis of the local-vertical, local-horizontal (LVLH) coordinate system (as detailed in Fig. 6) is zero. Worse still, the coupling between the cross- and along-track baselines of Cartwheel is serious because its cross-track baseline is mainly derived from the vertical baseline separation x(t), and one of its inherent property, a direct consequence of the Clohessy\u2013Wiltshire (CW) equation detailed in Section III-A, is the close coupling between the vertical separation x(t) and along-track displacement y(t) [6]. Therefore, to eliminate the coupling and to reduce the difficulty of satellite formation optimization, the Pendulum formation was brought forward", + " The C-W equation is given as [1] \u23a7\u23aa\u23a8 \u23aa\u23a9 x\u0308 \u2212 2w y\u0307 \u2212 3w2x = fx y\u0308 + 2wx\u0307 = fy z\u0308 + w2z = fz (7) where (x, y, z), (x\u0307, y\u0307, z\u0307) and (x\u0308, y\u0308, z\u0308) represent the relative position, relative velocity, and relative acceleration expressed in the LVLH coordinate frame; ( fx , fy, fz) is the perturbation or active control force in the same coordinate system; and w is the Keplerian mean motion of primary satellite. The LVLH frame is a noninertial frame that rotates along with the motion of the primary satellite. The X-axis of the LVLH frame points outward along the primary satellite\u2019s radius; the Z -axis is perpendicular to the orbital plane; and the Y -axis completes the right-hand frame. The LVLH frame is shown in Fig. 6. Please note that, in the spaceborne M-SAR system, the primary satellite refers to the satellite that transmits radar signals. Because the solution of C-W equation [1] has a secular term, the relative motion usually diverges with time, unless external control is applied or a boundedness condition is satisfied. The bounded periodic solution of C-W equation can be expressed as (8) [1] \u23a7\u23aa\u23aa\u23aa\u23aa\u23aa\u23a8 \u23aa\u23aa\u23aa\u23aa\u23aa\u23a9 x(t) = A sin( 2\u03c0 T0 t + \u03b1) y(t) = 2A cos( 2\u03c0 T0 t + \u03b1) + y z(t) = B sin( 2\u03c0 T0 t + \u03b2) (8) where A and B are the amplitudes of the motion on the X- and Z -axes, respectively, y is the displacement between the satellite formation center and reference frame, \u03b1 and \u03b2 are the initial phases corresponding to the initial positions of satellites in LVLH, and T0 is the orbit period [1]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002388_2015-01-0090-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002388_2015-01-0090-Figure2-1.png", + "caption": "Figure 2. CAD model of the truck frame", + "texts": [ + " Channel shaped side rails are capable of resisting bending loads, while appropriate cross members may increase torsional stiffness of the total frame [1, 2]. In Hotchkiss suspension, leaf springs are attached to an axle. A front spring end is pivoted to a vehicle structure, while a rear end is connected to a shackle [3]. The ladder frame of the investigated truck was comprised of side rails, cross members, structural reinforcements and suspension brackets. A Computer Aided Design (CAD) model, which represented the actual geometry of the truck frame (Figure 2), was generated in SOLIDWORKS. The truck frame consisted of a main frame and a sub frame, on which an engine and a transmission system would be mounted. The main frame comprised two straight side rails and five cross members. The channel-shaped cross members were located at front and rear ends of both side rails. Two circular closed tubes connected both side rails at locations of rear spring mounting brackets while the other one was located approximately in the middle of the main frame. Structural reinforcements were added between upper and lower flanges of the frame rails and connected to circular tubes" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001667_6.2002-5402-Figure12-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001667_6.2002-5402-Figure12-1.png", + "caption": "Fig. 12: Baseline configuration for the supersonic business jet showing surface densities at the cruise condition and structural stresses at the maneuver condition. The density is normalized by the freestream value and the von Mises stresses are normalized by the material yield stress.", + "texts": [], + "surrounding_texts": [ + "1H. M. Adelman and R. T. Haftka. Sensitivity Analysis of Discrete Structural Systems. AIAA Journal, 24(5):823\u2013832, May 1986. 2M. A. Akgu\u0308n, R. T. Haftka, K. C. Wu, and J. L. Walsh. Sensitivity of Lumped Constraints Using the Adjoint Method. AIAA Paper 99-1314, Apr. 1999. 3N. Alexandrov and J. E. D. Jr. Multilevel Algorithms for Nonlinear Optimization. Technical Report NASA TR-94-53, 1994. 4N. M. Alexandrov and R. M. Lewis. Comparative Properties of Collaborative Optimization and Other Approaches to MDO. In Proceedings of the First ASMO UK / ISSMO Conference on Engineering Design Optimization, July 1999. 5K. G. Bhatia and J. Wertheimer. Aeroelastic Challenges for a High Speed Civil Transport. AIAA Paper 93-1478, Feb. 1993. 6R. Braun and I. Kroo. Development and Application of the Collaborative Optimization Architecture in a Multidisciplinary Design Environment. In Multidisciplinary Design Optimization: State of the Art. SIAM, 1996. 7S. A. Brown. Displacement Extrapolation for CFD+CSM Aeroelastic Analysis. AIAA Paper 97-1090, Jan. 1997. 8E. J. Cramer, J. E. Dennis, P. D. Frank, R. M. Lewis, and G. R. Shubin. Problem Formulation for Multidisciplinary Optimization. SIAM Journal on Optimization, 4:754\u2013776, Nov. 1994. 9A. DeMiguel and W. Murray. An Analysis of Collaborative Optimization Methods. AIAA Paper 2000-4720, 2000. 10P. E. Gill, W. Murray, M. A. Saunders, and M. H. Wright. User\u2019s Guide for NPSOL (version 4.0). A FORTRAN Package Nonlinear Programming. Technical Report SOL86-2, Stanford University, Department of Operations Research, 1986. 11A. A. Giunta. A Novel Sensitivity Analysis Method for High Fidelity Multidisciplinary Optimization of Aero-Structural Systems. AIAA Paper 2000-0683, Jan. 2000. 12M. E. Holden. Aeroelastic Optimization Using the Collocation Method. PhD thesis, Stanford University, Stanford, CA 94305, 1999. 13A. Jameson. Aerodynamic Design via Control Theory. Journal of Scientific Computing, 3:233\u2013260, 1988. 14A. Jameson, L. Martinelli, and N. A. Pierce. Optimum Aerodynamic Design Using the Navier\u2013Stokes Equations. Theoretical and Computational Fluid Dynamics, 10:213\u2013237, 1998. 15S. Kodiyalam and J. Sobieszczanski-Sobieski. Bilevel Integrated System Synthesis with Response Surfaces. AIAA Journal, 38(8):1479\u20131485, aug 2002. 16I. Kroo, R. Tracy, J. Chase, and P. Sturdza. Natural Laminar Flow for Quiet and Efficient Supersonic Aircraft. AIAA Paper 2002-0146, Jan. 2002. 17I. M. Kroo. Decomposition and Collaborative Optimization for Large Scale Aerospace Design. In Multidisciplinary Design Optimization: State of the Art. SIAM, 1996. 18J. R. R. A. Martins, J. J. Alonso, and J. Reuther. AeroStructural Wing Design Optimization Using High-Fidelity Sen- sitivity Analysis. In Proceedings \u2014 CEAS Conference on Multidisciplinary Aircraft Design Optimization, Cologne, Germany, pages 211\u2013226, June 2001. 19J. R. R. A. Martins, J. J. Alonso, and J. J. Reuther. HighFidelity Aero-Structural Design Optimization of a Supersonic Business Jet. AIAA Paper 2002-1483, Apr. 2002. 20J. R. R. A. Martins, I. M. Kroo, and J. J. Alonso. An Automated Method for Sensitivity Analysis Using Complex Variables. AIAA Paper 2000-0689, Jan. 2000. 21J. R. R. A. Martins, P. Sturdza, and J. J. Alonso. The Connection Between the Complex-Step Derivative Approximation and Algorithmic Differentiation. AIAA Paper 2001-0921, Jan. 2001. 22K. Maute, M. Nikbay, and C. Farhat. Coupled Analytical Sensitivity Analysis and Optimization of Three-Dimensional Nonlinear Aeroelastic Systems. AIAA Journal, 39(11):2051\u2013 2061, Nov. 2001. 23J. Reuther, J. J. Alonso, A. Jameson, M. Rimlinger, and D. Saunders. Constrained Multipoint Aerodynamic Shape Optimization Using an Adjoint Formulation and Parallel Computers: Part I. Journal of Aircraft, 36(1):51\u201360, 1999. 24J. Reuther, J. J. Alonso, A. Jameson, M. Rimlinger, and D. Saunders. Constrained Multipoint Aerodynamic Shape Optimization Using an Adjoint Formulation and Parallel Computers: Part II. Journal of Aircraft, 36(1):61\u201374, 1999. 25J. Reuther, J. J. Alonso, J. R. R. A. Martins, and S. C. Smith. A Coupled Aero-Structural Optimization Method for Complete Aircraft Configurations. AIAA Paper 99-0187, 1999. 26J. Reuther, J. J. Alonso, J. C. Vassberg, A. Jameson, and L. Martinelli. An Efficient Multiblock Method for Aerodynamic Analysis and Design on Distributed Memory Systems. AIAA Paper 97-1893, June 1997. 27J. Sobieszczanski-Sobieski. Sensitivity of Complex, Internally Coupled Systems. AIAA Journal, 28(1):153\u2013160, Jan. 1990. 28J. Sobieszczanski-Sobieski and R. T. Haftka. Multidisciplinary Aerospace Design Optimization: Survey of Recent Developments. AIAA Paper 96-0711, Jan. 1996. 13 of 14 American Institute of Aeronautics and Astronautics Paper 2002\u20135402" + ] + }, + { + "image_filename": "designv6_24_0000894_piers.2017.8262402-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000894_piers.2017.8262402-Figure1-1.png", + "caption": "Figure 1. Schematic diagram of the proposed antenna. (a) Top view (feeding structure), (b) back view, (c) 3-D view.", + "texts": [ + " The feeding techniques are oriented in such a manner that they can produce orthogonal mode in the dielectric resonator. Symmetric radiation pattern and low cross-polarization is achieved from the both port. Paper is organized in the following sections: antenna geometry and final result are given in Section 2 and Section 3 respectively. Diversity performance and conclusion are given in Section 4 and Section 5 respectively. Geometry of the feeding structure as well as the complete antenna structure is shown in Fig. 1. The proposed antenna structure consists of a single alumina ceramic material based cylindrical dielectric resonator (Al2O3 (\u03b5Alumina = 9.8, tan \u03b4 = 0.002)) which is placed on a FR4 substrate (\u03b5sub = 4.4, tan \u03b4= 0.02). Two different feeding mechanisms are used to excite the cylindrical dielectric resonator. The dielectric resonator is excited with the help of two different feeding techniques. At port 1, a plus shaped CPW transmission line is used to excite the dielectric resonator. A microstrip line (etched on bottom layer of the substrate) is fed from port 2" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000972_iros.2006.282448-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000972_iros.2006.282448-Figure5-1.png", + "caption": "Fig. 5. Convergence of the limit-cycle from the various initial conditions", + "texts": [ + "4 and the dynamic equation of the closed-loop system is represented by: x(3)(t) + 2\u03b6b\u03c9bx\u0308(t) + \u03c9b 2x\u0307(t) = kp \u03b1 (\u03c9b 2\u03c6(x) + b\u03c6\u0307(x)) \u03c6(x) = { fh(t) (x < 0) \u2212mg (x > 0) (14) where x is the position of the load cell and x(n) is the nth order derivative of x. This system is 3rd order differential equation with a nonlinear element \u03c6(x) being a step function with respect to x. We define the transfer function from the control input u, which means force, to the position x as Txfh(s) = Tvfh(s) 1 s (15) Using the parameters of the left column in Table I, the phase portrait of the limit-cycle is calculated as shown in Fig.5. The horizontal axis and the vertical axis indicate the position and the velocity of the load cell, respectively. The upper half of the plane indicates that the motor is lowering the object, and the lower half indicates that the motor is hoisting it. A negative position (left half plane) indicates that the object is in a position above the ground, while the positive position (right half plane) indicates that the object is placed on the ground and the rope becomes loose. Regardless of the initial condition, the dynamics converge on a fixed locus shown in a thick black line" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003834_ectc.2007.373951-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003834_ectc.2007.373951-Figure1-1.png", + "caption": "Figure 1: (a) A piece of paper is cut into a serpentine. (b) When pulled, the serpentine elongates by twisting out of plane.", + "texts": [ + " The concluding remarks are given in Section 5. 2. Illustration of a principle A helical spring can elongate substantially, even though the material that makes the spring can only sustain a small strain. One could fabricate electronic circuits on a helical platform, but this approach would require microfabrication in three dimensions, a technology that requires substantial development itself. To be compatible with planar microfabrication technology, the platform must be planar. As an illustration of a principle, Fig. 1 shows a piece of paper cut into a serpentine, and pulled at the two ends. While initially planar, the serpentine elongates by twisting out of plane, so that a large elongation induces only small strains. The serpentine illustrates the principle: a film of a stiff material can be made compliant if the film is suitably patterned. For a film on a substrate to elongate substantially by twisting out of plane, two conditions must be met: the substrate must be sufficiently compliant, and the film must be suitably patterned", + " The substrate is meshed with eight-node linear brick elements, with size-matching elements at the film/substrate interface, and coarser elements far away from the interface. We model both the film and the substrate as linear elastic materials, with Young's modulus Efilm = 100 GPa and Esub =1 MPa to 100 MPa, so that Esubl Efilm ranges from 10-5 to 10-3. Poisson's ratio is taken to be 0.3 for both materials. 0.12 - max 0.08 - 0.04 - n nn _ --- hLsub=5 - hsb/L=0.05 Esub I Efilm = 0-5 )max < 3.5% at a relative elongation of 25%. The out-of-plane displacement of the film is antisymmetric with respect to the x2 axis (similar to the elongated paper serpentine in Fig. 1 (b)); when the modulus of the substrate increases, the displacement is gradually confined in the plane, and \u00a3max also increases. For example, when Esubl Efilm 1 cmax = 11.6% at the relative elongation of 25%. For a serpentine with a large width-to-thickness ratio, bending and stretching within the plane leads to a much larger strain than bending and twisting out of the plane. Figure 3 also shows that 8max increases as the substrate becomes thicker. Further calculations show that, however, Emax becomes insensitive to the thickness of the substrate when hsub L exceeds about unity" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000761_j.jmatprotec.2016.09.004-Figure9-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000761_j.jmatprotec.2016.09.004-Figure9-1.png", + "caption": "Fig. 9. Stretching strain distribution at 1.5 s, (a) radial strain (b) tangential strain (c) axial strain (d) effective strain.", + "texts": [ + " 8(c) shows the axial position of 18 points. Fig. 8(d) shows he initial longitudinal single grid, and Fig. 8(e) shows the sinle longitudinal grid after deformation. The parallel element mesh (d = 4 mm). deformation also occurs. The axial flow of the material increases from the inner wall to the outer wall, and the difference increases as rolling length increases. Therefore, axial inhomogeneous deformation is significant in the CWR of hollow parts. Uneven deformation and local deformation are mainly attributed to surface friction. Fig. 9 exhibits a CWR-stretching strain distribution at 1.5 s. The stretching stage was a stable rolling stage in which the strain changed smoothly and the radial metal was compressed. The radial strain was compressive and was maximized in the entrance side as the tensile strain state. The tangential strain was dominated by compressive strain, that is, the deformation zone located near the deformation of the metal pull produced tensile strain. The axial strain was tensile. The deformation zone generated by the squeeze pressure strain of the die wedge indicated that when the inner metal experienced maximum tangential and axial strains, the outer layer gradually decreased in size" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001017_12.2004415-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001017_12.2004415-Figure3-1.png", + "caption": "Figure 3 Layout of collimated laser inputs and reflecting mirrors on the 60 cm x 60 cm upper breadboard. The lasers are all directed downward by the set of right-angle prisms mounted in the center of the breadboard.", + "texts": [ + "aspx The design of the beacon transmit system was undertaken with the objective of assuring that all of the beams overlap at the telescope focus, thus assuring that they overlap in the far field. Each collimated beam is fed into the optical system and aligned by controlling the tip and tilt of the beam collimator mount and a secondary tip-tilt mirror. This allows fine control over both the position from which each beam originates, as well as the angle into which it enters the system. Initial beam alignment is performed on a small breadboard elevated above the main optical table (Figure 3). This places the individual mounts close to the stable breadboard surface, and allows all of the beams to be directed together down to the main table below. It also provides easier top-access to the mounts in the laser enclosure system, and limits interference with the crossing beams on the table below. The hexagonal pattern of the beams is achieved by sending the beams into an assembly holding six right-angle prisms, similar to an approach verified in previous experiments[4]. This assembly is necessary to place the beams close enough together that they can be propagated through the F/76 telescope system without vignetting or diffracting from mirror edges" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002329_icinfa.2015.7279553-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002329_icinfa.2015.7279553-Figure1-1.png", + "caption": "Fig. 1 Hydraulic ejection loader composition", + "texts": [ + " Then the dynamic simulation model was established based on ADAMS and verified by bench experiment in section 3. The consistency of ejection velocity was taken to evaluate the performance stability of the ejection loader in section 4. Simulation results show that the consistency of the ejection velocity gets worse with the decrease of the gas pressure in the accumulator. Parameter analysis shows that velocity consistency can be obviously improved by valve displacement adjustment. Conclusion and suggestions were given in the end. As shown in Fig.1, the hydraulic ejection loader consists of oil source, bladder accumulator, relief valve, proportional valve, ramming cylinder, speed mechanism, rammer body, track and sliding pusher. Bladder accumulator features advantages of quick response and small inertia to meet the demand of big instantaneous delivery during the ejection process. Before ejection starts, motor drives the pump to inject oil into the accumulator until the pressure of the accumulator reaches a certain value. Then the relief valve will unload the pump to save power" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002440_j.sna.2016.01.048-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002440_j.sna.2016.01.048-Figure3-1.png", + "caption": "Fig. 3. Schematic diagram of experimental setup.", + "texts": [ + " Then, the degassed mixture was poured into a prepared mold (outer and inner diameters and thickness: 20 mm, 5 mm, and 500 m, respectively) and cured in the thermal vacuum chamber at 60 \u25e6C for 4 h. Finally, the 500- m-thick PDMS membrane with a Young\u2019s modulus of about 2.63 MPa was obtained by carefully detaching it from the mold by using tweezers. Note that the prototype of the liquid lens does not have any membrane in the lens chamber; so, the liquid interface is open, which is similar to the other liquid lenses [10,40]. s as a i p a p a E R e p 3 t s s t b a p a w t i i h The schematic diagram of the experimental setup is shown n Fig. 3. The setup consists of electrical and optical systems. A ower supply (IT6720, ITECH Electronic Co., Ltd.) is used to operte the electromagnetic system to actuate the tunable liquid lens laced on a three-dimensional (3D) traverse system. Test images re obtained by using a charge-coupled device camera (EO-1312C, dmund Optics) or a high-speed camera (Phantom Miro eX4, Vision esearch Inc.), which is integrated with a zoom lens (VZMTM 450i o, Edmund Optics). The images are recorded on a personal comuter" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000534_aa-06-2019-0104-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000534_aa-06-2019-0104-Figure1-1.png", + "caption": "Figure 1 PCA for partial alignment", + "texts": [ + " (2014) introduce an initial transformation based on principal component analysis (PCA) and conduct an optimization framework of TrICP to complex global transformations. However, this initial process may lead to misalignment when the overlap of two point sets is only partial, i.e. partial alignment, since the distribution of an object is not consistent with that of partial object in general. For example, the barycenter for the hammer is completely different from that for its partial data set as in Figure 1. To settle this problem, we need to introduce feature based alignment to TrICP. It is essential to define a proper feature based criterion for cutting point set into the overlap part and the trimmed part. Curvature and normal direction are two widely used geometric features in point set registration (Chen and Medioni, 1992; Johnson and Hebert, 1997). The distance between two normal directions, i.e. the function of their inner product, is also introduced as a criterion for point set registration" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001764_tia.2010.2057398-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001764_tia.2010.2057398-Figure1-1.png", + "caption": "Fig. 1. Cage model\u2014cut view.", + "texts": [ + " During each start or stall, the centrifugal forces, the uneven temperature along the bar height, and the different radial thermal expansions between the end rings and the core end cause cyclic stresses. Finite-element analysis (FEA) tools available nowadays combined with good modeling techniques can help the engineer to estimate with accuracy the mechanical behavior of the components. However, analytical pre-work is necessary to ensure that the temperature input to the model is accurate. The typical rotor geometry under consideration is presented in Fig. 1. There is no possible test to be performed in the factory to ensure that a motor will meet the life requirements, so the engineer and the customer need to rely on the calculation processes. Some materials that are frequently used in electrical machines do not have their thermal properties easily found in textbooks. The material data used here might be considered as typical, and the values were determined from catalogs and/or test data. The properties consider average operating temperatures. Typical silicon steel has a loss per mass of approximately 5 W/kg when magnetized with 1" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002862_detc2007-34856-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002862_detc2007-34856-Figure5-1.png", + "caption": "Figure 5: Sprocket with a non-integer number of teeth", + "texts": [ + " Because each sprocket has a different pitch diameter, and thus a different number of teeth, each provides a different gear ratio when driven by another sprocket. To allow a bicycle to have infinitely incremented gear ratios, there would need to be an infinite number of sprockets, each with a different pitch diameter, and thus infinitely different numbers of teeth. If a sprocket is created, however, with a pitch diameter whose resulting circumference is not divisible by the pitch of the chain, a non-integer number of teeth would result, as shown by the overlapping teeth in Figure 5, thereby ensuring at least one place on the sprocket where the chain will not mesh properly. Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 02/04/2016 T (diametral pitch = 16) Standard bicycles overcome the non-integer tooth problem by having a finite number of sprockets, both driving and driven, which all have the same diametral pitch. Each of these sprockets has an integer number of teeth, which allows them to mesh properly throughout their complete rotation. The pivot-arm CVT, originally developed by Mortensen [7], and later analyzed and modified by Christensen [8] at Brigham Young University, is an embodiment that employs compliant members to provide a mechanism that will change its active diameter to create a continuous range of mechanical advantage" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003208_j.wear.2018.02.020-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003208_j.wear.2018.02.020-Figure1-1.png", + "caption": "Fig. 1. (a) Schematic illustration of the abrasion-testing equipment, (b) SEM micrograph of the abrasive sand [22].", + "texts": [ + " A deviation was taken in terms of duration of test in order or achieve appreciable wear, later, the weight loss was normalized against the load and distance traversed. Rectangular samples of dimension 25.4\u00d776.2\u00d712.7 mm3 were machined from the initial block using electro-discharge machine (EDM). The broad faces of the samples were ground using 400 \u00b5m grit SiC papers to remove any damage caused by EDM prior to the testing and to achieve the desired surface roughness. The schematic of the equipment and the secondary electron image of the abrasive sand used for the testing are shown schematically in Fig. 1a) and b) respectively [22]. The rotating speed of the wheel and sand flow rate were kept uniform throughout the experiment. The experiment was conducted for ten minutes at the specified parameters, and the specific wear rate was calculated using weight loss after abrasion. The abrasion test was conducted on five samples to confirm and validate the repeatability of the results under the same experimental conditions. The specific wear rate (SWR) was averaged out and reported with statistical error" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001939_2015-01-0088-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001939_2015-01-0088-Figure7-1.png", + "caption": "Figure 7. Gear Shift Fork - Stress Results", + "texts": [ + " Stresses in each leg should be below the endurance strength of the material. As the constraint of bulky design already mentioned with two pads, a third pad has been introduced to share the load. But this is done in a phase wise manner in which the third pad will only come to contact after there is specified deflection in the main two legs. This third pad is located at the center of the fork. If the FOS is less than 1.0, then equal amount of thickness is increased on both the legs until we arrive at or above FOS 1.0 as shown in the Fig. 7. Contact analysis has been carried out to check the contact status using FE Ansys Solver. Surface to surface contact is provided between mating surfaces of forks legs & third pad with synchronizer ring. Contact starts at fork legs that will have a hinge effect before it touches third pad meeting the required deflection. A minimum nominal gap is provided for 3rd pad to share the load when max/abuse load appears on the load shifting Jaw. Gap could be adjusted if the FOS in the legs goes below 1.0 for the Max/Abuse Load as shown in the Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001284_0196-8904(95)00355-x-FigureI-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001284_0196-8904(95)00355-x-FigureI-1.png", + "caption": "Fig. I. Cross-section of an 8/6 slotted SRM with simplified drive circuit for one phase.", + "texts": [], + "surrounding_texts": [ + "Pergamon o196-s9o4(95)oo355-x\nEnergy Convers. Mgmt Vol. 38, N6. I, pp. I-II, 1997 Copyright \u00a9 1996 Elsevier Science Lid\nPrinted in Great Britain. All rights reserved 0196-8904/97 $t5.00 + 0.00\nP E R F O R M A N C E A N A L Y S I S O F PV P U M P I N G S Y S T E M S\nU S I N G S W I T C H E D R E L U C T A N C E M O T O R D R I V E S\nH A M I D M. B. METWALLyI\"[ \u2022 and WAGDY R. ANIS 2 ~Electrical Engineering Department, Faculty of Engineering, Zagazig University, Zagazig and 2Electronics and Communications Department, Faculty of Engineering, Ain Shams University, 1 Sarayat Street,\nAbbasia, Cairo, Egypt\n(Received 25 January 1995; received for publication 6 November 1995)\nAbstract--A PV pumping system using a switched reluctance motor (SRM) is thoroughly investigated in this work. This motor is supplied by a d.c. voltage through a simple switching circuit. This drive circuit is much simpler than the normal d.c./a.c, inverter required to supply the induction motor. The efficiency of this motor is considerably higher than that of the equivalent d.c. or induction motors. In addition, because of the simple construction, the SRM is cheaper than these conventional drives. Because of the above advantages oftbe SRM, the proposed system has higher efficiency and lower cost as compared with other systems. A design example is studied in detail to explore the advantages of PV pumping systems based on this new drive. The study of the performance of the proposed system showed that the operating efficiency of the motor is about 85% during most of its working time. The matching efficiency between the PV array and the proposed system approaches 95%. The major part of the losses takes place in the pump and the riser pipes. This loss represents one third of the total available energy. Copyright \u00a9 1996 Elsevier Science Ltd\nPumping systems Photovoltaic systems PV system simulation Performance of PV pumping systems\nI. INTRODUCTION\nWater pumping is one of the most popular applications of solar energy now-a-days. The pumping system consists of three basic components, a PV array, an electric motor, and a water pump. In addition, in some cases, there are storage batteries and power conditioning equipment. Generally, the electric motor used in these systems is one of two main types, either induction motor or d.c. motor.\nThe squirrel cage induction motor is a well developed motor. It is one of the simplest forms of motors available. Because of its simplicity, it is cheap, robust, and reliable. An electronic inverter is necessary to convert the d.c. power generated by the array to an a.c. power for the motor. For variable speed operation, a variable frequency supply is required. The supply must also be of variable voltage if a constant torque is to be obtained throughout the speed range. The requirements of variable voltage and frequency supply, as well as an a.c. waveform of minimum harmonics, necessitate the use of a complex and, hence, expensive inverter. As the inverter becomes complex, the reliability in service decreases. Hilloowala and Sharaf[l] investigated the design, modeling, and simulation of a PV powered pump irrigation system using a single phase induction motor drive.\nThe d.c. motor, on the other hand, is the most complicated and, hence, expensive motor. It has all the common disadvantages associated with the sliding brush contacts and commutator, such as the need for routine inspection and periodic maintenance. Brush contacts are also unsuitable for hazardous environments, and the associated arcing problem limits motor size and speed range. Direct current motors are not suitable for high speeds because of the wear and tear of the commutator and brushes, which limit the life of the motor. In spite of all these disadvantages, d.c. motors are used extensively in PV pumping systems. This is because they can be coupled directly\ntTo whom all correspondence should be addressed at: Yanbu Industrial College, P.O. Box 30436, Yanbu AI Sinalyah 21477, Kingdom of Saudi Arabia.", + "to the PV array, giving a simple and inexpensive system. The design and performance of directly coupled PV pumping systems are investigated in many publications, some examples are found in Refs [2-10].\nIn order to avoid the complex electronic inverter needed for the induction motor and the brushes and commutator in conventional d.c. motors, brushless d.c. motors are used in pumping systems, but at limited scale. The brushless d.c. motor has a rotating permanent magnet and a stationary armature winding, thus avoiding brushes and commutator. Commutation of electric current in the stationary armature is done by an electronic circuit in accordance with the rotor position. This is why brushless d.c. motors are called electronically commutated d.c. motors. The design of a photovoltaic pumping system based on this drive is reported by Longrigg [11].\nIn this work, a PV pumping system based on a new drive is investigated. The new drive is called a Switched Reluctance Motor. Although this motor is a new entrant to the field of variable speed drives, it has gained acceptance worldwide in many applications. As reported by Jones [12], this motor is used in a wide range of general industrial and domestic applications, such as fans, blowers, conveyors and pumps, and it may earn the title of the motor of the 1990s.\n2. SWITCHED RELUCTANCE MOTOR\nRecently, the interest in using SRMs as variable speed drives has increased remarkably. This is because this motor has many advantages over the other types of motors. It is robust, reliable, cheap, and of higher efficiency. During the last 15 yr or so, many publications concerning the design and performance prediction of this motor appeared. Among those important publications are Lawrenson et al. [13, 14] who investigated the design and performance aspects of the motor. Davis et al. [15] studied the drive circuit design.\nA brief review of the construction, principle of operation and figures of merit of this motor is given here. Figure 1 shows the lamination profile of an 8/6 slotted motor as an example. Also, the elementary drive circuit of one phase is shown in the figure for the purpose of illustration. The construction of the motor is very simple with the following features:\n(1) there are salient poles (teeth) on both stator and rotor, i.e. it is doubly salient (2) concentrated simple windings mounted on the stator poles (3) there are no windings of any type on the rotor (4) the number of poles (teeth) carried on the stator and the rotor are different (5) the windings on diametrically opposed poles are connected in series and/or parallel to form\none phase, and the number of phases and poles are open to a wide variety of choices.\nThe operation of this motor is based on the production of reluctance torque, that is, the torque produced by the forces tending to align the stator and rotor teeth when a particular phase is excited.", + "These forces are independent of the direction of the excitation current. By switching the excitation current between phases, in an appropriate manner, continuous rotation is achieved. Although the motor can operate in the open loop mode, the closed loop mode of operation is usually preferred. This is to avoid the difficulties which could arise during starting or during load disturbances. In the open loop mode, it is not necessary to sense the rotor angle at the instant of switching-on, and the rotor adjusts itself according to the mechanical load imposed on it. In the closed loop mode, the phase switching can be controlled via a simple shaft-mounted position transducer, so that the stator windings are switched on and off in accordance with the rotor position.\nSimplicity of construction makes the motor cheap, robust, and reliable. Unidirectional currents in stator windings allow the use of cheap and simple power converters. In addition, this motor has better thermal characteristics than other types of motors because of the absence of windings on the rotor. The above features produce a motor with a high specific output and high efficiency over a wide range of speeds and output powers.\nThe block diagram of the system is shown in Fig. 2. It consists of a PV array, a storage battery connected to the array through a battery voltage regulator, a switched reluctance motor together with its drive circuit, and a centrifugal pump. In the following subsections, each component is introduced and the mathematical models for the main components are given.\nThe PV array I - V characteristic is given by:\nI = ILoG -- I0[exp((V + IRs) /VT) - 1] - (V + IRs)/Rsh (1)\nwhere I is the PV array current (A), IL0 is the light generated current of the PV array at G = 1 kW/m 2, G is the solar irradiance (kW/m2), I0 is the PV array reverse saturation current (A), V is the PV array operating voltage (V), V T is the thermal voltage of the PV array (V), R s is the series resistance of the PV array (f~) and Rsh is the shunt resistance of the PV array (f]).\nThe function of the battery voltage regulator (BVR) is to protect the battery against overcharging and deep discharging. If the battery is overcharged, the BVR opens switch SI, in Fig. 2, to disconnect the PV array, and the load is supplied from the battery and, hence, it is partially discharged. On the other hand, if the battery is deeply discharged, the BVR opens switch $2, and hence, the load is disconnected to allow the PV array to charge it. Thus, the BVR controls the battery charge and, hence, the battery voltage within the allowed limits. Such controlled performance ensures long life for the battery. The detailed design and performance characteristics of one such battery voltage regulator are given by Anis [16].\nThe function of the battery is to keep the operating voltage of the motor within certain allowed limits. This is necessary in order to operate the motor at high efficiency. Allowing a wide range" + ] + }, + { + "image_filename": "designv6_24_0002711_j.autcon.2020.103529-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002711_j.autcon.2020.103529-Figure8-1.png", + "caption": "Fig. 8. Optimum mesh density in (a) one particular layer of the hollow cylinder, (b) TPMS-Gyroid, (c) TPMS-Primitive block.", + "texts": [ + " (2) and depends on the distance of consecutive points in the ordered dataset of the generated toolpath. During the simulation, each step adds one element set that is grouped following the converted toolpath trajectories from Gcode. The material early-age properties (i.e., strength and stiffness) used in each set of elements during the printing process thereby vary according to time. For 3DCP simulation of the hollow cylinder, linear eight-node continuum solid elements (C3D8R) with an average size of 5 \u00d7 5 \u00d7 5 mm3 are applied for each released layer. This optimal mesh density (Fig. 8(a)) is chosen after performing a convergence study with different refinements of mesh sizes until there was no change at the moment of collapse [30\u201332]. This provides reasonable accuracy whilst being inexpensive in terms of computational cost. Each layer is tied together at the bottom and top surfaces. With regard to the simulation of TPMS V. Nguyen-Van et al. Automation in Construction 123 (2021) 103529 structures, the first-order four-node solid elements (C3D4) are adopted. The optimum element size is selected as 4 mm for simulating the Gyroid (Fig. 8(b)) and Primitive (Fig. 8(c)) block, which was demonstrated in the author\u2019s previous work [17]. It is assumed that there is high friction between the bottom layer in each printed structure and the floor. Then, all boundary conditions are fixed at the layer\u2019s bottom surface. General contact is used for the other layer surfaces. Nonlinear effects due to large displacements and deformations, as well as unsymmetrical matrix storage, are chosen due to fresh-state material behaviours. Additionally, parabolic extrapolation is defined for each increment to make the step increment increase smoothly and overcome some common errors caused by large distortions as the printing material reaches its yielding state" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001317_amm.415.524-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001317_amm.415.524-Figure3-1.png", + "caption": "Fig. 3. Worm rough cutting process", + "texts": [], + "surrounding_texts": [ + "When re-equip gear-hobbing machine, firstly, the rotational speed must be slow down, secondly, corresponding work needed to add to the rotary table. Rotary center axis of worm must vertical with rotary center axis of the rotary table, and through the level line of worm bearing. Level line of worm bearing should be equal to the length of worm and worm gear centre distance. A dial to adjust the peripheral feed movement was added on the grinding head on rotary table of the machine. And then installed on the dividing head a little slash skateboard grinding head to control the axial feed movement of the grinding wheel. Wheel was mounted on the spindle spindle, spindle used special inverter, the maximum rotational speed of spindle is up to 35000r/min. According to processing requirements, worm as a workpiece, the requirements of rotating speed is not high, therefore should let the speed of main axle get down. So, a drive which the velocity ratio is 40 was affixed between the main electrical machinery and engine bed of existing gear-hobbing machine, its original of I level triangle with drive did not variable, then, the roll tooth machine of speed dropped to 36 r/min. Its structure as shown in Fig. 1." + ] + }, + { + "image_filename": "designv6_24_0001395_ijvp.2016.075351-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001395_ijvp.2016.075351-Figure2-1.png", + "caption": "Figure 2 Finite element coordinate system (see online version for colours)", + "texts": [ + " Both the ANCF and the FFR formulation will be used in this investigation. ANCF finite elements will be used to develop the fully nonlinear coupled model, while the FFR formulation will be used to define the equations required for the post-processing stress calculations based on forces obtained using rigid body simulations. For this reason, these two formulations are briefly introduced in this section. The FE/FFR formulation requires the use of four coordinate systems for each finite element, as is shown in Figure 2. The global coordinate system 1 2 3X X X is fixed in time and space. Kinematic constraints which can represent mechanical joints and specified motion trajectories are formulated in this coordinate system using a set of nonlinear algebraic equations that depend on the system generalised coordinates and can also be time-dependent. A body coordinate system 1 2 3 ,i i iX X X called the floating frame of reference (FFR), forms a single set of axes for the entire assembly of elements in the body i and, as such, serves to express the connectivity of all the elements in the body", + " The element coordinate system 1 2 3 ij ij ijX X X for an element j on the deformable body i is rigidly attached to the element. This coordinate system translates and rotates with the element. The final coordinate system, the intermediate element coordinate system 1 2 3 ,ij ij ij i i iX X X has its origin rigidly attached to the body coordinate system and does not follow the deformation of the element. This coordinate system is initially oriented to be parallel to the element coordinate system and is represented by the dotted axes lines shown in Figure 2. It is important to note that the use of the intermediate element coordinate system is necessary in order to obtain an exact modelling of the rigid body inertia in the case of complex structures with discontinuities. This concept is similar to the parallel axis theorem used in rigid body mechanics. Furthermore, since this coordinate system has a constant orientation with respect to the body coordinate system, exact modelling of the rigid body inertia in the body coordinate system can be obtained using a constant transformation" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003168_6.1978-1006-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003168_6.1978-1006-Figure5-1.png", + "caption": "Figure 5 Short Scarf Nozzle Cross Section", + "texts": [ + " In this case, the nozzle was rough scarfed to the final exit plane angle, placed upon the shear spin mandrel and slapped with lead until the part conformed to the mandrel and thus had the appropriate inner mold line contour. Rim Preforms Short scarf nozzle rim preforms are fabricated using a hydroform process. A male punch which bas tbe dimensions of the inner mold line of the rim is used to control the dimensions of the rim preform. The inner section of the rim preform is trimmed away to provide the rim cross section shown in Figure 5. \\J Rim preform dies for tbe long scarf nozzles are considerably more complicated due to the steeper scarf angles involved. Typically, short scarf nozzles have scarf angles varying from 16 to approximately 21\" from a flat plane as compared to the up to 64\" off a flat plane required for long scarf nozzles. The form die shown in Figure 9 clearly indicates the degree of overhang which exists at one end. Forcing the columbium material back around this corner proved to be a significant problem. A normal hydroform operation proved to be inadequate to move the metal to the required degree" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002224_3527608117-Figure4.14-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002224_3527608117-Figure4.14-1.png", + "caption": "Fig. 4.14 Layout of an automated casting cell for production of LOKASIL\u00ae-cylinder crankcases.", + "texts": [ + " Thereby the end of the mold filling is of central importance for the process control concerning preform infiltration. Here the casting piston speed must be reduced so as to keep the increase in pressure during 1054.3 Production of LOKASIL\u00ae Cylinder Crankcases the infiltration within permitted values. This process is so stable that the contact surface inspection by eddy current can be reduced to a sampling inspection. The squeeze casting machine is a component of a fully automated casting cell (Fig. 4.14), in which two robots take over all necessary handling. Whilst robot 2 cleans the die and spreads the die release agent, robot 1 takes three heated preforms (Fig. 4.15) out of the preheating furnace and sets them on tempered removable sleeves. Subsequently, these are positioned in the casting tool. After closing the die the casting process takes place, as the casting piston moves upward and fills the cavity slowly with melt. The preforms are infiltrated with melt and the casting solidifies under high pressure" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001014_ispcc.2012.6224344-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001014_ispcc.2012.6224344-Figure2-1.png", + "caption": "Figure 2. Simulated Metal gate Symmetric DG-MOSFET", + "texts": [ + " An asymmetric DG-MOSFET either has synchronized but different input voltages to both of the identical gates, or has the same input voltage to two gates but gates having different work functions. The name of \u201csymmetric\u201d and \u201casymmetric\u201d essentially depicts presence or absence of symmetry of the electric field inside the channel of the DG-MOSFET [4]. In this work we had designed an n-channel metal gate symmetric DG-MOSFET on Sentaurus TCAD simulator [5-6]. Spacers of Si3N4 is used to reduce fringing field effect. In this simulation Density Gradient quantization model had been used [7-8]. The simulated device structure (figure 2) is a symmetric DG-MOSFET with following parameters: Threshold voltage of the device is an important parameter which decides the device performance. The value of gate to source voltage (Vgs) for which significant amount of mobile electrons accumulates in the channel region so that a conducting channel is formed, called the threshold voltage. From figure 3 and figure 4 it is clear that as we increase the dielectric constant of gate dielectric, linear as well as saturation threshold voltage of the device decreases" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002082_mop.10661-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002082_mop.10661-Figure1-1.png", + "caption": "Figure 1 Schematic drawing for the two-arm, conical log-periodic antenna with rectangular teeth", + "texts": [ + " 1, January 5 2003 pattern, have been extensively studied and documented with design graphs based on empirical results [2, 3] and accurate numerical analysis [4]. The objective of this paper is to present a new broadband log-periodic antenna with high directive gain, good match, and linear polarization: the conical log-periodic antenna. This antenna is derived from the conical spiral antenna introduced by Dyson [3] and the planar log-periodic antennas introduced by DuHamel and Isbell [5]. It is shown schematically in Figure 1. Like all frequencyindependent antennas, the geometry of this antenna is mainly described by angles, and lengths are introduced to specify the smallest and largest dimensions of the (finite) antenna. In section 2, the antenna geometry is presented, including various antenna arm designs, and the modeling of the antenna using the finite-difference time-domain (FDTD) method is briefly described. Results for the impedance, gain, and pattern are then presented in section 3. The two-arm, conical log-periodic antenna, shown in Figure 1, consists of two metallic strips or arms placed on the surface of a cone. A practical approach for constructing this antenna is shown schematically in Figure 2(a). The metallic arms are formed on a MICROWAVE AND OPTICAL TECHNOLOGY LETTERS / Vol. 36, No. 1, January 5 2003 29 planar, flexible circuit board through a wet chemical etch. The circuit board is then wrapped around a conical mandrel, and soldered or taped along its seam. The basic cone, shown in Figure 2(b), is characterized by the half angle of the cone o and the diameters d and D that limit the extent of the antenna at the small and large ends", + " Hence, for practical purposes, the antenna is linearly polarized on axis. Figure 5 shows vertical-plane ( 0\u00b0), far-zone field patterns for the antenna with the sinusoidal teeth at three frequencies within the operational bandwidth: (a) f 1.5 GHz, (b) f 3.0 GHz, and (c) f 4.5 GHz. Patterns are given for the two orthogonal components of the electric field, namely, the component (solid line) and the component (dashed line). For this antenna, the component of the field in the forward direction ( z\u0302, see Fig. 1) is clearly dominant, and the radiation is concentrated near the direction 180\u00b0. Hence, this antenna predominantly radiates linear polarization unidirectionally towards the apex of the cone. The conical, log-periodic antenna was analyzed using the FDTD method. Three different tooth shapes for the antenna were investigated: rectangular, triangular, and sinusoidal. These new antennas were shown to be broadband with uniform input impedance, gain, and pattern. The antennas with the triangular and sinusoidal teeth were seen to clearly outperform the antenna with rectangular teeth", + " They have been obtained by considering proper geometrical modellings of AUT and by exploiting the results [7] concerning nonredundant representations of the EM fields, radiated by sources enclosed in arbitrary convex domains with rotational symmetry and observed on surfaces having the same symmetry. The bi-polar scanning proposed by Rahmat-Samii et al. in [8, 9] represents a convenient alternative to collect NF data over a plane. In such a scanning method the AUT rotates axially, whereas the probe is attached to the end of an arm that rotates around an axis parallel to the AUT one. This allows collection of the NF data on a grid consisting of concentric rings and radial arcs (see Fig. 1). The bi-polar scanning maintains all the advantages of the planepolar one, while providing a simple and cost-effective measurement system. In fact, since the arm is fixed at only one point and the probe is attached at its end, bending is constant, which allows planarity to be maintained. Moreover, rotational movements are preferable to linear ones, since rotating tables are more accurate than linear positioners. Unfortunately, the original approach in [8, 9] does not take advantage of the nonredundant representations of EM fields and, as a consequence, it requires a large amount of useless NF data" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003607_wnwec.2009.5335798-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003607_wnwec.2009.5335798-Figure1-1.png", + "caption": "Figure 1. Four units of DSEG section.", + "texts": [ + " THE FINITE ELEMENT ANALYSIS OF DSEG In this paper simulation model of 24-32-DSEG is built in ANSOFT, and simulation is made. The data of flux density are exported to MATLAB to calculate magnetic field change of DSEG. As in the mode of half wave rectification, DSEG external characteristic is harder [14], the simulation and analysis is based on the mode of half wave rectification. We can get the average flux density waveform of four units which are stator tooth, rotor tooth, stator yoke and rotor yoke. Table I shows Major dimension of the motor.Fig.1 shows the four units of DSEG section, unit 1 is stator yoke, unit 2 is stator tooth, unit 3 is rotor tooth, and unit 4 is rotor yoke. Fig.2 shows the flux density waveform of the 4 units. It is obvious that the flux density of each unit changes periodically, and the change is complicated. The frequency of unit 1 and that of unit 2(stator core) are the same, they are both 266.7Hz, and the frequency of unit 3 and that of unit 4(rotor core) are the same, they are both 33.3Hz which is lower than that of unit 1and unit 2" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003651_kem.579-580.265-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003651_kem.579-580.265-Figure1-1.png", + "caption": "Fig. 1 2D and 3D model view of the transmission(Excluding housing)", + "texts": [ + " Aimed at a certain automobile transmission, this paper mainly completes the entire transmission system modeling in design module of the MASTA software. The sequence of the entire modeling process in order is to establish the model of all gear shafts, and then to determine the space position of these shafts, followed by assembling other parts of gear shaft, such as gears, bearings, synchronizers, etc. According to the detailed parameters of the gear, shaft and bearing and other parts, the transmission simulation model is established in the MASTA software as shown in Fig. 1. Load cases are mainly used to define the input conditions of the proposed model, which can be load cases under actual working conditions or load cases of bench test. In the MASTA software, the working condition usually refers to torque, rotating speed and acting time under the action of a certain power flow, while the load cases are the combinations of different working conditions. Therefore, before the definition of load cases, we need to define the input and output power flows for the proposed model" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000083_ir-03-2018-0042-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000083_ir-03-2018-0042-Figure7-1.png", + "caption": "Figure 7 Front wheel description", + "texts": [ + " The wearable vehicle has only two-force sensors that are installed in the feet module instead of using four sensors. The detailed design is shown in Figure 6. The components of the fast motion mode which allow the system to be used as a vehicle, including the front wheel mechanism, rear wheels mechanism and seatingmechanism. Front wheels are fixed onto the shank of the wearable vehicle. These links can be adjusted by telescopic columns that are fixed in different positions to adapt to the two modes of motion by using a securing mechanism as shown in Figure 7. The castor wheel is fixed by axial shaft and two ball bearings on both sides to decrease friction and wearing. Elastic material is fixed in the top of the chassis of the front wheel to absorb the shock while moving in uneven terrain. All-terrains wearable vehicle B.M. Sayed et al. Industrial Robot: the international journal of robotics research and application Volume 46 \u00b7 Number 6 \u00b7 2019 \u00b7 740\u2013762 The rear wheels are attached to the trunkmechanism by using a movable link which is actuated to assure that the center of gravity of the wearable vehicle stays within the support polygon" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000761_j.jmatprotec.2016.09.004-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000761_j.jmatprotec.2016.09.004-Figure4-1.png", + "caption": "Fig. 4. (a) Geometry mode of forging for DEFORM-3D, (b) forging for hollow valve blocking.", + "texts": [ + " ) The friction coefficient was assumed to be constant throughout the CWR process regardless of whether cooling water was used. The frictional force in the shear friction model is defined by (Eq. (3)): s = m \u00d7 k (3) here fs is the frictional stress, k is the shear yield stress, and m is he friction factor. Table 2 provides a summary of the adopted simulation paramters of the CWR process. The main process parameters for imulation are listed in Table 3. .2. Forging process The geometry model illustrated in Fig. 4(a) consists of a top die, orkpiece, and bottom die. The forging die in CWR has been used in ecks for locating datum. Neck size is unchanged during formation. he hollow valve stem that is connected to the neck parts of the urface is the cone. The angle of the valve cone is 4\u25e6. The forging ie is positioned on the conical surface, thereby reducing the size f the workpiece during axial channeling formation. Moreover, the ngle of the cone ( ) is equal to that of the workpiece, die radius R) is greater than the target product neck radius of 1\u20132 mm. These actors enhance the formation of the workpiece in place. Fig. 4(b) epicts the control valve employed in hollow valve forging. The summary of the adopted simulation parameters of the forgng process is listed in Table 4. 3. Experimental tests on the chain process for producing hollow valves Experimental tests on the CWR process were conducted using the H500 mill and electric tube furnace at the University of Science and Technology Beijing in Beijing, China. This equipment is displayed in Fig. 5(a). Fig. 5(b) shows the forging mill. The forging experiment was conducted at Huaiji Dengyun Auto-Parts Co" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003663_amm.416-417.281-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003663_amm.416-417.281-Figure1-1.png", + "caption": "Fig. 1 Cross-section of (a) 12/8 TLSPM; (b) 15/8 TLSPM Fig. 3 Winding arrangement of 12/2 slot/pole", + "texts": [], + "surrounding_texts": [ + "Three-phase TLSPM is considered in this paper. The motor can be divided into integral and fractional slot motor according to per pole per phrase slot number q, where the q is integer to integral slot, and q is fraction to fractional slot. Integral slot motor regards a pair poles as a cycle, while fractional slot motor is different. Assume that the motor slot and pole pair number have the common factor t, then they can be expressed as Z=Z0t and p=p0t respectively. Per pole per phrase slot number q is expressed as 0 0 22 mp Z mp Z q (2) It is shown in (2)that fractional slot motor puts p0 as a cycle. Supposing p0 pair poles of the motor part as a virtual unit motor for simplifying study, a motor may be composed as t virtual unit motor. Per pole per phase slot number q can be transformed as d c bq (3) where c and d have no common factor. The unit motor number and slot number of TLSPM can be expressed as follows. Unit motor number d p t 2 (4) Unit motor slot number )(30 cbdz (5) Winding factor reflects the extent of phase voltage reduction due to the short pitch and distribution structure, and the derivation of winding factor for fractional slot is as follows. Short pitch ratio mq y (6) Short pitch winding factor ) 2 sin( pK (7) Distribution factor )(2 sin)( ) 2 sin( cbdm cbd mKd (8) Winding factor dpdp KKK (9)" + ] + }, + { + "image_filename": "designv6_24_0002770_aps.2002.1018237-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002770_aps.2002.1018237-Figure2-1.png", + "caption": "Fig. 2 (a) A top view of the 3x3 array at the transverse plane of a hard walled waveguide with the array parameters. (b) Experimental setup for measuring the scattering parameters of the 3x3 system.", + "texts": [ + " This GSM analysis tool was used to better understand the effects of some of the design parameters on the amplifier performance in [ 101. In this section, a spatial power combiner system consisting of a 3x3 array of coax fed patch antennas inside a hard hom will be considered. A comparison between simulation and experimental results for this system will be presented. A single coax fed microstrip patch antenna (as the unit cell) was designed to resonatearound 10 G H z u s i n g a R o g m h i d 4003 substrate. The other array parameters in Fig. 2a are vertical spacing (dy), horizontal spacing (dx) and aperture dimensions (X and Y). The array was placed at the hard hom aperture (with the dimensions 7.3025 cm and 5.398 cm) with a uniform horizontal spacing of 0.32h, a uniform vertical spacing of 0.28h. The hard horn was fabricated using a Rogers Duroid 5880 dielectric and a standard gain hom with a length of 8.6 cm. A near field scan performed at the aperture of the hard horn at a frequency of 9.7 GHz showed a uniform field distribution in the absence of the array. The coax fed patch antenna array was put in front of the hard hom as shown in the experimental setup in Fig. 2b. A network analyzer (HP 85lOC) was used to obtain the scattering parameters. One of the ports was connected to the waveguide that feeds the hard horn antenna. The second port was connected to one of the coaxial connectors that feed the patch antennas while the rest of the coaxial connectors were terminated with 50Q loads. A general waveguide-based spatial power combining system. The same structure was also simulated using the generalized scattering matrix cascading approach. The GSM for the hard horn was obtained through mode matching with a staircase approximation" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001120_6.1997-1198-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001120_6.1997-1198-Figure5-1.png", + "caption": "Figure 5. Wing Twist Activation Concepts", + "texts": [ + " Figure 4 shows calculated (nominal) values of torque at wing mid-span and tip to achieve 2 and 5 degrees of twist for a full-scale aircraft and scaled models. (The torque requirements increase essentially as the fourth power of the geometric scaling factor - the values shown in the figure are slightly different because of differences in the materials used.) While it is feasible to achieve the torque requirements for the models, it is obvious that meeting scaling requirements will be a significant challenge to transition this technology to a full-scale aircraft. This is discussed further in Section 3. 2.2 Design Wing Twist: Several design concepts (Figure 5) were considered for twisting the wing for the wind tunnel models. Initial trade studies indicated that the integrated torque box concept was structurally most efficient. However, on further examination, the design presented severe manufacturing difficulties and appears to be somewhat impractical. Hence the shape memory alloy (SMA) torque tube actuation was chosen and a design with two concentric tubes as shown in Figure 5A was implemented. This technique functioned well in the tunnel, but because the final wind tunnel model was significantly stiffer than the scaled model (primarily due to escalation of wing skin and spar web thickness from the original scaled values to prevent local panel buckling), maximum wing tip twist of only about 1.25 degrees was realized. If the stiffness were scaled exactly, 3 to 5 degrees of twist at the wing tip could easily have been achieved. Further details of the torque tube design are presented in Reference 10 and 11" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000969_1687814015589561-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000969_1687814015589561-Figure1-1.png", + "caption": "Figure 1. Finite element model after meshing (a) is made up of four sections, namely the car body, the brackets, the bumper system and the impactor and (b) consists of three parts, namely the bumper beam, the foam and the fascia.", + "texts": [ + "13 Meaningful or useful results will depend on the type of elements employed, the element quality, the number of elements used, and the simulation of boundary constraints and loading conditions.13 Since the software Altair HyperMesh has great advantages over other software in meshing, it is adopted in this study. At first, the computer-assisted design (CAD) data of the impact model was imported to the coupling environment of Altair HyperMesh and LSDYNA. Second, the geometry cleanup and meshing were done to the model. The finite element model after meshing is shown in Figure 1. Figure 1(a) is the impact finite element model and Figure 1(b) is the finite element model of the bumper system. The bumper beam is attached to two energy absorbing brackets as shown in Figure 1(a), and the brackets are attached to the thin-walled structure which is used to simulate the car body using the rigid steel material. The connector between the car body and the energy absorbing brackets is accomplished by shifting the rear circle grid of the two brackets to the car body. The connector which is also used for the brackets themselves between the brackets and the bumper beam is achieved by setting up rigid joints. Nonlinear explicit impact modeling elements were used in this study. Because the thickness of the bumper beam, brackets, fascia, and car body is much smaller than the other dimensions, the shell element is the best choice for meshing" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003215_6.1988-4153-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003215_6.1988-4153-Figure4-1.png", + "caption": "Fig. 4 Finite-element model degrees of freedom", + "texts": [ + " The DAS is built around a STD bus architecture and employs an 8088 microprocessor to direct its operations. Data and commands are passed betweentheVAXStation I1 and the DAS via a high speed parallel interface. By absorbing the time delays associated with sensing and actuation, the DAS frees the VAXStation I1 computer to give it more time for the computations associated with control algorithms. 111. DYNAMIC MODEL A finite element model was developed for the purpose of obtaining the system modes. Each rib, and the boom, is divided into 10 beam-type elements (see Fig. 4 ) . and the hub is modeled as a very stiff plate. A consistent formulation is used to develop the mass matrix, so that accuracy of the modes would be maximized for this level of discretization. The finite element model takes into account the inertia of the levitator pulleys and the mass of the counterweights, the tension in the boom caused by gravity, and compression in the ribs caused by the coupling wires. The normal modes and their frequencies can be obtained by solving a generalized eigenvalue problem in standard form: where K is the stiffness matrix, M is the mass matrix, and x is the eigenvector with frequency o" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001717_j.biosystemseng.2008.02.010-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001717_j.biosystemseng.2008.02.010-Figure2-1.png", + "caption": "Fig. 2 \u2013 Formation of a jet of air on a flat surface.", + "texts": [ + " The classic turbulent jet theory developed by Abramovich (1963), in conjunction with a text on fluid mechanics by Fay and Sonwalkar (1996), was used to establish how the air-jet forces and velocities are influenced by supply pressure and nozzle diameter. The impingement of a gas jet onto a surface results in a complex flow field. In the simplest case of a jet discharging a fluid with a uniform initial velocity field Uo into a medium moving at zero velocity, the boundary layer thickness in the initial section of the jet is zero (Fig. 2). The short region of the jet in which the centre line velocity remains constant is called the \u2018initial region\u2019 (Fig. 2). As reported by Abramovich (1963), the velocity profile reduces and expands with increasing distance from the jet source. The plane representing the limit of constant axial velocity is called the \u2018transitional crosssection\u2019. Beyond this point in the \u2018main region\u2019 of the jet, the centre line velocity of the jet, Um, gradually reduces as the diameter of the jet continues to expand. However, in the main region of the jet the pressure is almost constant and equal to the pressure in the surrounding space. Following the turning of the jet in the impingement zone, the flow exits parallel to the surface (Fig. 2). The pressure on the impingement surface returns to ambient in the \u2018wall-jet\u2019 region. Calculation of the momentum of an air-jet and hence the force available requires knowledge of the jet\u2019s initial velocity and its mass flow rate. The calculation is not straightforward because even at relatively low differential pressures the compressibility of air results in a supersonic flow. As a result we need to know whether the flow is critical or not. For a perfect gas with constant specific heat the absolute critical pressure Pc at which the supersonic flow occurs can be described by (Fay and Sonwalkar, 1996): Pc \u00bc Ps 2 g\u00fe 1 g=\u00f0g 1\u00de (1) Abramovich (1963) concluded that, for a jet exhausting to the atmosphere under the ISO standard atmospheric conditions of PA \u00bc Pc \u00bc 101", + " It was found that PDA decreased as the nozzle diameter increased (Fig. 16). With a 2.5 mm diameter nozzle a high percentage of the arils was damaged, and the percentage of damaged arils increased with increasing air pressure. This shows that under these conditions the impingement pressure on the surface of fruit was greater than the mechanical strength of the fruit flesh. This could be explained by the fact that the impingement pressure depended on both the impingement force and the impingement surface area (Fig. 2). In the case of the nozzles with a diameter of 2.5 mm, although the impingement force is the low (Fig. 6), the impingement area is also small in comparison with other nozzles (Fig. 2). Hence, a greater pressure may be applied to the surface of fruit. The fact that a higher percentage of damaged arils was obtained for nozzles 2.5 mm in diameter, in comparison with others (Fig. 16), demonstrates this. However, because the impingement area covered a small area of the fruit surface, it did not have the capability of penetrating into the interstices between the arils to extract them as effectively, as observed for the nozzles with diameters of 3 and 3.5 mm. For the nozzles of 4.5 mm diameter, although the normal impingement force was higher than with 3 and 3" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003284_20140824-6-za-1003.01709-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003284_20140824-6-za-1003.01709-Figure3-1.png", + "caption": "Fig. 3. Shematic of the system from a control point of view", + "texts": [ + " Several parameters of the introduced model are determined by physical properties of the EXCOR system. The remaining parameters were determined by experiments and suitable fitting procedures. More details about the modelling process are given in [4]. In this section the developed control system will be presented. At first an analysis from the control point of view is performed. After that an approach for a model based estimation of the control variable is given. Finally the basic control strategy is explained. Fig. 3 shows an overview of the appreciable values of the control system. The available measurement variables of the treated heart assist device are the piston position x and the piston drive pressure p p . Due to reliability there is no further sensor information in the artificial blood pump. There are two control signals in the EXCOR VAD. A motor torque can be applied to the electro-mechanical piston drive. Furthermore a balancing valve can be opened or closed to compensate the enclosed air mass in the pneumatic system" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000725_17513472.2013.765311-Figure18-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000725_17513472.2013.765311-Figure18-1.png", + "caption": "Figure 18. Folding sequence for a type e quadrant.", + "texts": [], + "surrounding_texts": [ + "As noted earlier, there are 136 possible distinct colour schemes for flat units, but every unit can be divided into two halves and there are only 16 distinct half-unit colour schemes. But the problem is even smaller than that: each half-unit consists of two quadrants, and there are only four possible colourings per quadrant. So we can create any four-quadrant unit if we have a folding method that allows the four quadrants of a unit to be folded independently; then only eight distinct quadrant folding methods would be needed. In order to create a modular cube from the unit, one must fold a unit that possesses the desired colour scheme and also has tabs and pockets in the right places \u2013 basically, the same places as the original Simon Cube (or the generic variant). Another important consideration is efficiency of paper utilization: if the unit is too thick and has too many layers, it will be difficult to fold and assemble the units. If we take the square tabs at the ends of the units to be unit squares, then the Simon and generic units are folded from a 4 \u00d7 4 grid of squares. In order to get enough additional paper to create colour patterns, it is necessary to go to a larger grid, but as it turns out, one needs only one additional square of width/height for each quadrant to obtain all the desired colourings. So a 6 \u00d7 6 square gives sufficient paper for folding all of the colour schemes. Every unit starts out the same way, using the folding sequence shown in Figure 12. The procedure for folding any unit specified by its 4- character label is to fold the top and bottom quadrants on the left side, corresponding to the first two letters; then rotate the unit 180\u25e6 and fold the top and bottom quadrants on the (new) left side corresponding to the second two letters. The mapping of which letters go with which quadrants is shown in Figure 13. Folding sequences for upper-left quadrants of type a\u2013d are shown in Figures 14\u201317. (Note that in order to keep the folding sequences tractably short, I have used several terms common in the world of origami to describe combination folds: \u2018reverse fold\u2019, \u2018Elias stretch\u2019 and so forth. See, for example, my book [8] for a detailed explanation of these manoeuvres.) Folding sequences for the lower-left quadrants of types e\u2013h are shown in Figures 18\u201321. Each quadrant is shown folded in isolation here, but of course, you will need to fold four quadrants together in a real unit. Generally, the top and bottom quadrants can be folded nearly independently of one another, but after folding the pair from one side and rotating to fold the other two, some care must be taken when folding the last two quadrants to avoid disturbing the first two folded. All of the quadrant types give rise to the desired colouring and have the tabs and pockets required to assemble the units. However, quadrant g is slightly different from all the rest; its tab is triangular, rather than square, so its joint is not as secure as the others. That is an aesthetic deficiency in this type of unit. That establishes another aesthetic criterion for colourings: all else being equal, it would be desirable to avoid use of quadrant g. Even if a unit set includes quadrant g, there is a way of avoiding it that sometimes works: that is to invert the parity of the paper, i.e. start with the white side up when folding the basic building block. This would have the effect of inverting the colour parity of the exposed regions (as well as changing colours for some hidden regions), in a way that can be summarized as follows: a \u2194 d b \u2194 c e \u2194 h f \u2194 g Inverting the colour of the unit will change any g quadrants to f quadrants. However, it will also perform the reverse as well, so any unit whose colouration is of the form *f\u2013*g would be immune to this dodge. But perhaps we will get lucky. Let us find out." + ] + }, + { + "image_filename": "designv6_24_0002470_tmtt.2005.845707-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002470_tmtt.2005.845707-Figure8-1.png", + "caption": "Fig. 8. Class-C baluns. (a) C1 prototype with two zeros at dc, a zero at infinity, and a nonredundant UE. (b) C2 prototype with a zero at dc, two zeros at infinity, and a nonredundant UE. (c) and (d) Corresponding miniaturized mixed lumped-distributed physical topologies.", + "texts": [], + "surrounding_texts": [ + "A 50 : 100- Class-B balun with a center frequency of 1 GHz and a commensurate frequency of 2 GHz is synthesized here. The objective is to design a balun with a return loss of 15 dB at the single-ended port and a bandwidth ratio of 1.54 : 1 giving lower and upper passband edge frequencies of 0.82 and 1.26 GHz, respectively. The purpose is to implement the balun on an FR4 printed circuit board (PCB) with the following specification: Substrate thickness: 62 mil (1.57 mm) Relative dielectric constant: 4.7 Metal thickness: 2.2 mil (0.05 mm) Minimum spacing: 7 mil (0.17 mm) Loss tangent: 0.016. (19) Now, the characteristic polynomial of the prototype may be constructed for a passband ripple of 0.05 dB using [13] (or [14]) with two zeros at dc and two zeros at infinity leading to1 (20), shown at the bottom of the following page, from which the square of the magnitude of the reflection transfer function is evaluated using (21) 1High precision of the numerical values must be retained throughout the synthesis process. may then be found with the knowledge of (22) leading to (23), shown at the bottom of this page. The input impedance is then evaluated in a 1- system from (24) which is then synthesized using standard element extraction. Upon synthesis, the resulting prototype is scaled up to suit a 50- system, and a 50- UE was then added after the source impedance, resulting in the prototype of Fig. 10(a). At this point, according to (6)\u2013(8), we have the following circuit parameters obtained from Fig. 10(a): (25) (20) (23) and, thus, substituting (25) into (9)\u2013(11) gives (26) Now the parameters in (26) are then substituted in (12)\u2013(13) to give (27) (28) Conversion of the above set of even- and odd-mode impedances into physical dimensions using ADS2 yields width and spacing of values mil 0.75 mm mil 0.08 mm (29) mil 0.96 mm mil 0.06 mm (30) Unfortunately, according to (19), the spacing between the coupled lines violates the minimum spacing restriction imposed by the PCB manufacturer. This is when the circuit transformations become useful. By transforming the prototype of Fig. 10(a) to that of Fig. 10(b) and making use of the stepped impedance transmission-line equivalence, the transformed prototype of Fig. 10(c) results. Performing similar manipulation as above, we get new coupled-line parameters as follows: (31) (32) This leads to new physical dimensions of mil 0.43 mm mil 0.22 mm (33) mil 0.41 mm mil 0.22 mm (34) It is now clear that the above dimensions are realizable; however, it is always desirable to obtain identical coupled-line parameters to construct a physically symmetrical balun structure. This would allow the division of the balanced load resistance by a factor of two to make feasible practical measurements. In terms of circuit synthesis, it is actually possible to obtain a prototype with elements values such that this condition is satisfied. This is done by iterating on the synthesis cycle for different values of passband ripple until the evaluated values of the resulting even- and odd-mode impedances of the pair of coupled lines become identical. Fig. 10(d) shows another synthesized and transformed prototype achieving a passband ripple of 0.0806 dB, i.e., a return loss of approximately 17.37 dB with the same bandwidth ratio of 1.54 : 1. It is now seen from the figure that the impedances of the UE and the input series stepped transmission line are virtually equal (80 79.993 \u2014this is just the necessary required condition. 2Advanced Design System (ADS), Version 2003A, Agilent Technol., Palo Alto, CA, 2004. Now, from Fig. 10(d), we have (35) and substituting these into (9)\u2013(11) yields (36) Evaluating (12)\u2013(13) using (36) gives identical coupled-line parameters of values (37) At this stage, a manual optimization step is believed to improve the return loss level. Doing so results in the circuit of Fig. 10(e) with a return loss of approximately 20 dB. However, the values of the impedances of the UE and input-series stepped transmission line must remain fixed (80 ) to guarantee the condition of physical symmetry. The prototype of Fig. 10(e) results in the physical balun layout of Fig. 10(f) with even- and odd-mode impedances as shown in the same figure. In a similar step as before, ADS was utilized to obtain some physical dimensions that were found to be mil 0.25 mm mil 0.23 mm (38) The shunt open-circuited stub at the balanced output and the balanced load must now be scaled accordingly using (15) and (16) with evaluated as 0.561 to give 270.51 and 221.707 , respectively. The balanced load impedance is then scaled to 100 by a pair of transmission lines of impedance 74.449 each. The remaining stubs are left unscaled according to (17). Due to the fact that we are using a bandpass prototype, it is possible to approximate any or all of its open-circuited stubs by lumped capacitors with very little deterioration of circuit\u2019s passband performance. Now, each open-circuited stub is approximated by a lumped capacitor using the relationship (39) where is the impedance of the open-circuited stub and its corresponding capacitor value. This approximation is valid over the vicinity of the passband of the filter but will affect the upper stopband characteristics. However, in many cases, such a step will lead to better stopband performance suppressing upper undesired passbands. This point will be elaborated upon in Section V. Performing the approximation using (39) gives values of capacitors as shown in Fig. 10(f). A final superficial optimization step is required to adjust the return loss level after approximation to lumped capacitors giving lines that are 15-mil (0.38 mm) wide, 7-mil (0.17 mm) apart, and 950-mil (24.13 mm) long. Also, each output matching line is 49-mil (1.24 mm) wide, and 1636-mil (41.55 mm) long. It is interesting to note that this example shows yet another transformed Class-B balun topology not included in Fig. 9. This topology was also used in [5], apparently without any rationale behind its derivation." + ] + }, + { + "image_filename": "designv6_24_0001943_0890-6955(95)00104-2-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001943_0890-6955(95)00104-2-Figure2-1.png", + "caption": "Fig. 2. Meshing elements 5a and 5b.", + "texts": [ + "a_r (3) The Denavit-Hartenberg notation is an important mathematical tool in the kinematic analysis of open and closed-loop mechanisms [11]. It can also characterize the abilities of multi-axis machine tools and generate the desired cutter locations. The illustrative example shown in Fig. 1 is the variable pitch screw transmission mechanism (VPSTM) used in a shuttleless weaving loom [1]. It consists of a driving slidercrank mechanism and a driven variable pitch screw possessing four cylindrical meshing elements (Fig. 2). Elements 5a and 5b engage one thread on one side of the screw, while elements 5c and 5d (not shown in Fig. 2) engage the other thread on the other side. Here only the thread meshed with elements 5a and 5b is considered, since after this threaded profile is machined, the screw blank can be turned 180 \u00b0 and the same cutting path followed in machining the second thread. Note that in order to maintain a constant thread thickness the whole meshing part (Fig. 2) is twisted an angle ]3. To synthesize the profile for this screw, all links are numbered sequentially starting from the screw marked as \"0\" and ending at the meshing element marked as \"5\". Once the frame (xyz)i (i = 1,2,3,4) has been assigned to link i, the link parameters can be tabulated (Table 1). The relative position and orientation of frame (xyz)4 with respect to frame (xyz)o is given by equations (I) and (2): I CG SOl 0 fSOi 1 o,4, = fi ' - 'A,= Is0, I C 0 I 0 I f c o ~ - ' = ' , (4) NC Data Generation for 4-axis Machine Tools 343 Note the following observations from Figs 1 and 2 : - 0 2 - 03 + 04 = a3C(O2 + 0 3 ) + a: CO2 = 0, b4 = f, and alSO2 + a3S(~t2 + 0 3) = h(O~), simply equation (4)", + " The purpose of rough cutting is to give a basic shape to the workpiece, while leaving a small machining allowance ~ for the finishing operations. The location of cylindrical end mill (diameter is d0 with respect to a conical meshing element at the kth cut is shown in Fig. 3. This cutter location can be expressed as: 346 Psang Dain Lin and Ming Far Lee In order to prevent overcutting, the cutter should be located inside the meshing element, and )t is restricted in the interval of ( - r + d,/2 + E)--So and S=So cosec 8. The probability that the n o d lies between 6 and 8+d8 is given by Q(0) do. By considering the ratio of a spherical belt to that of a hemisphere, it is easily established that Q(e)=sin 8. The probability distribution for S is P'(S)dS which must therefore equal -Q(f3)d6. The result is generalized by setting P(x)=SoIy(S), where x=S/So and dx=dS/So", + " This arises from the fact that a layer of greater spacing S has a greater probability of intersecting the plane of section. Secondly, however, in a bulk material, with a single basic spacing SO, colonies or grains with larger 8 and smaller S have a correspondingly larger Probability of intersection. These two effects exactly cancel out. The authors are indebted to a referee for pointing this out. Some further details and derivation of the general formulae are given in the Appendix. It is important to recognize that the fundamental lamellar repeat distance is that for a combination of one layer of each kind as in Fig. 2(b), and not for either layer separately. Figure 3 is of fundamental importance for evaluating images from laminated structures. Corresponding relationships exist for random sectioning of other basic 142 SUU~~UXS such as distributions of spheres. In practice a frequency distribution of spacings as a fmction of S would enable the value of SO to be indicated accurately providing the sample was large enough and sufliciently representative. If there is a spread of spacings, as for example in a normal distribution, the shape of the basic dismbution is closely represented by the form of the curve or dismbution towards lower values of s. If the spacings are represented by two or more basic dismbutions this will be indicated if the peaks are high enough above the base and enough measurements have been made. In addition to the statistical distribution of spacings which m a y be treated, the variation of the angle 4 in Fig. 2 may be examined. This angle lies in the plane of the surface and is measured relative to some reference direction, and defines the directions of the traces of the lamellae in the surface. Statistical variations in may be associated with the presence of preferred directions adopted during the formation of the structure. They can then be associated with some texture present in the material, either prior to the formation of the lamination (as in a solid state transformation) or simultaneously developed (as in a solidification process involving prefemd directions)", + " Within these three regions smaller units of cells, similar to those shown in Fig. 4, are present and have varying spacings and orientations. Figure 6 shows optical transform patterns from Fig. 5. These are taken (a) with a wider beam diameter covering one of the major boundaries. There is a clear indication of two main groups with different orientations. Secondly, the \u2018diffraction spots\u2019 are quite large because there is a considerable variation in sub-cell spacing and orientatiOn. In cases like this it is quite easy to make reasonable average measurements of D and + (Fig. 2) and to give approximate ranges. Figures 6(b) and qc) show the diffraction patterns from the separated areas, which are indicated in both Figs. 5 and 6 by B and C respectively. The range of spacings and angles is then listed in Table 1. These results record the spacings for the exact magnification of Fig. 5 as well as for the magnification actually employed. It will be seen that there is good agreement with the spacings in Fig. 5. The angles 4 are approximately correct for normals to these planes, although the *action patterns as here illustrated are seen fiom the \u2018other side\u2019 of the object, i.e. looking towards plane G from the left in Fig. 1. This is equivalent to a rotation of the plane about the vertical axis in Fig. 2. The spacings in region A are not greatly different from those in B. In obtaining Figs. 6(b) and 6(c), the beam diameter was reduced the structure round the individual spots is now more affected by the diffraction of the beam on passing through a (circular aperture). It follows that \u2018selected area optical diffraction\u2019 becomes less exact and patterns more confused, the smaller the area irradiated. Satisfactory practical compromises can usually be made. 145 F ig . 6 . O pt ic al d if fr ac ti on p at te rn s r ela te d to F ig ", + " Thirdly, because the s t r u c n w has been heated for some time below the transformation temperature the lamellae in some grains have begun to break up into arrays of globular particlesthe process of spheroidization. Diffraction patterns from such areas may lose their simple character. The more general interpretation or evaluation of optical transforms as a measure of spheroidization is a likely development. K. W. Andrews and S. R. K e r n The pearlite spacings were measured in sufficient grains to enable an estimate to be made of the smallest spacing (i.e. equivalent to SO in Fig. 2). Comparisons were also made between a selection of measurements by optical diffraction, and by direct averaging of the spacing separations on suitable images of colonies as observed on a projection microscope. It was thus possible to make an inadental comparison between measurements by the two methods. These results were supplemented by further measurements of spacings from other samples and both optical and electron micrographs were used. Figure 9 shows the comparison in terms of spacings on the micrographs (in mm units)", + " R. Jolly and P. C. Wall who ammbuted some of the information used in Figs. 9 and 10. This was part of a research project for the third year of their degree. They also thank 154 Mr J. H. W d e a d who has kindly provided valuable information and background on the general approach to the quantitative metallography of particle sizes and dismbutions. A note on the &vatiOn of the forntulae fm Fig. 3 The approach is similar to that of Pellisier et al. (1942) apart from minor changes in notation. In Fig. 2(b) the lamella is the combined double layer (of cementite and ferrite in the case of pearlite) and S is the repeat distance for th is layer in the plane of section. SO is the true thickness. Clearly S= SO cosec 8 The probability that a normal to a single lamella lies at an angle between fl and B+dO is given by - area of zone '( de'area of hemisphere- 277 2n sin 8 d8 =sin 8.d8 ~ _ _ - The original authors did not consider the effect of layers forming aggregates of 155 K. W. Andrews and S. R. Keown different orientation, and we are indebted to a referee for drawing our attention to this" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002426_tvcg.2007.1033-Figure11-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002426_tvcg.2007.1033-Figure11-1.png", + "caption": "Fig. 11. This figure shows the mapping from muscle activation levels to joint torques.", + "texts": [ + " Given anatomical details derived from the biomechanics literature, we can estimate the maximum forces that can be transmitted from muscles through tendons to joints to contact points and, finally, to be applied to the object. To compute the set of wrenches that the hand can apply to a grasped object, we need the following equations. All anatomical parameters for these equations can be found in the Appendix, along with their original references. First, the joint torques generated by the tendons due to muscle activation can be computed as follows (Fig. 11): J \u00bcMPa; \u00f010\u00de where P is diag\u00f0p1;max; p2;max; . . . ; pn;max\u00de, where pi;max is the maximum force that can be generated along tendon i. Matrix M contains joint moment arm information and converts tendon forces to joint torques. Parameter a is an n 1 vector of activation levels, ranging from 0 (inactive) to 1 (at maximum force), for n tendons. Then, when the hand grasps an object, we can also map from the contact forces to the joint torques as follows (Fig. 12): 0J \u00bc JTf; \u00f011\u00de where f is a 3m 1 vector of contact forces with m as the number of contacts, and JT is the contact Jacobian, which maps contact forces to joint torques" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001360_aero.2017.7943762-Figure23-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001360_aero.2017.7943762-Figure23-1.png", + "caption": "Figure 23. Program A Boom View Simulation", + "texts": [ + " Sirius FM-6 Sun Planning Sirius FM-6 used 5 external cameras (Figure 13) to monitor all phases of the main reflector boom and bundle deployment and generate a video [2], see images Figure 19 through Figure 22. Another satellite with a large deployable antenna, Program A, has camera system was similar to Sirius FM-6 except that the more complicated deployment required 8 external cameras (Figure 13) to monitor all phases of deployment. In addition, simulated views through the camera lens were developed to assist in determination of best camera position and alignment, see Figure 23. As of this writing, the mission has not been launched. However, launch is expected early 2017 and the camera system will be used to observe the deployment of the 18-m antenna. To improve capabilities, SSL has developed an enhanced camera system (Figure 24) which is planned for use monitoring auxiliary payload deployments and an Earth view on an upcoming mission. It is planned for monitoring antenna deployments on the upcoming SXM-7/8 missions. There are significant design features added to this system: \u2022 Use of available bandwidth on standard spacecraft housekeeping telemetry link \u2022 Optical fiber SpaceWire links to high definition compact camera heads \u2022 General purpose image processing and storage management (Linux), \u2022 Scripted image capture \u2022 64 GB of internal image storage" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000083_ir-03-2018-0042-Figure23-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000083_ir-03-2018-0042-Figure23-1.png", + "caption": "Figure 23 FEA of the wearable vehicle while operating fast motion mode after optimization technique", + "texts": [], + "surrounding_texts": [ + "The human walking gait data freely provided by the OpenSim model Gait2392 is used. OpenSim software is built for the musculoskeletal simulation of the human during various motion activities (Delp et al., 2007). The OpenSim gait is shown in Figure 24. This gait data describes a full step of a normally walking human starting, and ending with the toe-off phase of the left foot. This obtained motion data is exported into MSC ADAMS model of the human. Revolute joints are added between the human pilot body parts and the normal motion splines are applied to those joints. The CAD model of the wearable vehicle is exported also to MSC ADAMS. The next step is to merge the twomodels (the humanmodel and the wearable vehicle model). Multiple variations have been carried out on the model of the wearable vehicle with the human. After the two models are merged, the human model is interfaced to thewearable vehicle throughMSCADAMS contacts. The load carried by the wearable vehicle in the simulation is 55 Kg. To measure the zero moment point (ZMP) in walking mode, force sensors were attached to the wearable vehicle feet. All-terrains wearable vehicle B.M. Sayed et al. Industrial Robot: the international journal of robotics research and application Volume 46 \u00b7 Number 6 \u00b7 2019 \u00b7 740\u2013762" + ] + }, + { + "image_filename": "designv6_24_0003600_anie.201300371-Figure13-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003600_anie.201300371-Figure13-1.png", + "caption": "Figure 13. A) Representation and real-life image of a cantilever-like two-component LCE producing asymmetric cilia-like motion depending on the light source. Reprinted with permission from Ref. [109]. Copyright 2009, Nature Publishing Group. B) Top: Representation of a motor laminated with a photoresponsive LCE film. Bottom: By simultaneous irradiation with UV and visible light, the plastic motor is propelled by a bendingunbending movement of the LCE film. Reprinted with permission from Ref. [159].", + "texts": [ + " Once the director orientation is rotated with respect to the long axis of the sample, an out-of-plane twisting motion can be observed owing to a shear gradient and contraction along the diagonal of the cantilever.[156] The bending of the photoactive LCEs can further be improved by creating splayed or twisted instead of just planar alignments of the mesogens.[157] The deposition with an ink-jet printer allows different LC materials to be arranged in cantilever-like shape: one sensitive to UV light, the other to visible light. Thus a cilia-like motion can be induced by addressing the different components with their respective wavelength of light (Figure 13A).[109] Instead of covalently linking the azo moiety to the elastomer, Palffy-Muhoray and co-workers created LC elastomer network with an azo-dye simply dispersed in it. When floating on water, the material was found to swim into the darker regions, that is, away from the irradiating light source as a result of exchanging momentum between water and the sample upon its bending motion.[158] A light-driven plastic motor was realized by wrapping a photoactive LCE film around two pulleys and illuminating the film with UV light and visible light from two opposing sides. The resulting contraction on one and expansion on the other side results in a rolling motion of the film, which propels the two pulleys (Figure 13 B).[159] The same concept and material could also be used for mimicking the three-dimensional movements of an inchworm walk and a robotic arm motion.[160] In recent years we could observe significant progress in the fabrication of LCEs, creating elaborate shapes that are capable of performing complex motions. Particularly, the possibility to manufacture micro- and nanosized LCEs allows their integration into lab-on-chip systems. These advances will fuel the transition from fundamental research to competitive commercial applications" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000210_1.2826118-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000210_1.2826118-Figure2-1.png", + "caption": "Fig. 2 Geometric entities Invoived in tlie synthesis of tlie spatiai guidance linkage", + "texts": [ + " From a methodological view point, instead, it provides additional insights because it po tentially enables symbolic computation of every coefficient of the only univariate algebraic equation that condenses the spatial Burmester problem. Moreover, the method is not based on the preemptive evaluation of the number (twenty) of possible solutions but provides such a number as a side resuh. A numerical example is finally reported that shows appli cation of the proposed method to a case study. Compatibility Conditions As outlined in the previous section, the spatial Burmester problem is aimed at synthesizing the linkage of Fig. 1 in such a way that the guided body can be indexed through seven given poses. With reference to Fig. 2, a couple of cartesian reference frames, W,, and W^, with origin at O/, and O^ respectively, are considered. While Wf, is fixed to the base, W^ is attached to the guided body. The jth (j = 1, . . . , 7) of the seven poses that the guided body is required to assume with respect to the base can be expressed by the 3 X 3 orthogonal matrix Rj for coordinate transformation from W^ to W/,, and by the components of vector (O \u2014 O )\u0302 in reference frame Wj^. The generic connecting rod BG (see Fig. 2) complies with the seven prescribed poses if and only if the following com patibility conditions are satisfied (Wampler et al., 1990): (G - B)J(G - B)j = L' 0 = 1, . . . ,7 ) (1) where ^ denotes the transpose operator, and L the unknown length of rod BG. Now the following positions are introduced: ^. = [(o.-o,)l (2) where the subscript g or b outside the square brackets means that the vector components in reference frame W or, respectively, W/, are considered. Consequently, vector (G \u2014 B) can be expressed in refer ence frame Wj, as (G - B)j = Rje -f+sj (3) and Eqs", + "org/terms Numerical Example This section shows application of the proposed method to a case study. In order to allow accurate verification of the reported numerical results, all input data are considered as free from round-off errors, hence exactly specified by the displayed number of digits. Table 1 lists the seven prescribed poses for the guided body in terms of vectors Sj, Uj, and Vj (j = 1, . . . , 7), all of them referred to frame Wf, and expressed in arbitrary length unit. While Sj has an already-known meaning [see Fig. 2 and positions (2)], vectors Uy and \\j indirectly define the 3 x 3 orthogonal matrix Rj representing the orientation of refer ence frame W^ with respect to W,,. Precisely, the x-axis of W\u201e has the same direction and orientation as Uj, while the y-axis of Wg lies in a plane parallel to both u^ and Vy, with a positive component along v,. By taking advantage of the proposed elimination method, the 60 X 60 matrix M is formed as having the leftmost twenty columns linearly dependent on x^. As an alternative to the symbolic computation of the determinant of matrix M \u2014or, which is the same, of the twenty-one coefficients tj (;' = 0, " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003423_ecce.2019.8913146-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003423_ecce.2019.8913146-Figure4-1.png", + "caption": "Figure. 4: Radial flux DFIG", + "texts": [ + " One of the major drawbacks of DFIG topology is that the power conversion circuitry is complex unlike the asynchronous generators and on top of that the slip rings connected to the rotor windings demand regular maintenance which proves to be inconvenient and expensive for ensuring long term operation of the system. Radial flux DFIG topologies are quite common in the literature as already discussed. The prime contribution of this work lies in making the introduction of AF-DFIG topology which to the authors knowledge is yet to be explored in detail. Before diving in the design details this subsection aims at laying out fundamental structures used for the work, for both axial and radial flux topologies. Fig. 4 shows a 2D view of the inner rotor radial flux DFIG topology used in this work, while Fig. 5 displays a single stator single rotor AF-DFIG 3D structure. Envelope dimensions, used for subsequent optimization, of the AF-DFIG are also marked in Fig. 5, where is outer diameter of the machine, is inner diameter of the machine and is the stack length of the machine. In addition to these parameters several other slot dimensions are optimized for the AF-DFIG as discussed in the next section. The fundamental difference between the two topologies lies in the air gap flux direction as clearly suggested by their names" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001924_sp.j.2096-5796.2019.0016-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001924_sp.j.2096-5796.2019.0016-Figure7-1.png", + "caption": "Figure 7 Schematic view of a six-di\u2010 mensional F/T sensor with composite elastic beam[29].", + "texts": [ + " Exact solution of the six-dimensional forces / torques can be obtained by decoupling the six groups of bridges. This kind of force sensor has high rigidity. It was first proposed by Carnegie Melon University, and also studied by Huazhong University of Science and Technology, that is HUST-FS6 6 DOF F/T sensor[28]. The aforementioned structures are cross-beam or vertical-beam structures with simple monolithic elastic elements, whereas interdimensional coupling is serious. In 1982, Schott proposed a six-dimensional F /T sensor with double-ring composite elastic beam structure[29], as depicted in Figure 7. This structure with combination of elements can reduce or eliminate the interdimensional coupling. Aiguo SONG et al: Multi-dimensional force sensor for haptic interaction: a review Figure 8 shows a classic structure of another commonly used 6 DOF force sensor based on Stewart platform[30], a 6 DOF parallel mechanism proposed by Stewart in 1965. The six spherical hinges of the upper or lower platform are distributed around the same circle. They are semi-symmetric, i. e., spherical hinges 1, 3, 5 and 2, 4, 6 distributed at intervals of 120\u00b0, respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000875_tap.2017.2780895-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000875_tap.2017.2780895-Figure2-1.png", + "caption": "Fig. 2. Using perforation technique to alleviate the problems associated with aligning and bonding the individual elements of the DRA. (a) Replacement of the free space between the elements in the radiating layer with lattice of holes in the perforated structure based on the perforation technique. (b) Variation of the effective dielectric constant with different diameters of perforated holes (d) for the fixed coupling gap (g) of 0.25 mm.", + "texts": [ + "0012) with a thickness of 0.508 mm. The aperture-coupled structure provides ample space for the biasing network by separating the radiating elements and feeding lines, which is extremely useful in expanding the subarray to larger arrays. Moreover, we have used the perforation technique to fabricate DRAs from a single dielectric sheet. By this means, we can eliminate the difficulties in aligning and bonding the individual element of the DRAs [21]. The utilized perforation technique can be clarified considering Fig. 2. In Fig. 2(a), we have shown the radiating layer of the proposed ESPAR and its perforated structure. It can be observed that the free space between the elements in the radiating layer is replaced with lattice of holes in the perforated structure. The diameter and spacing of the holes determine the effective dielectric constant of the material surrounding the DRAs, which is defined as [21] \u03b5eff = \u03b5r (( 1 \u2212 \u03c0 2 ( d d + g )2 ) + \u03c0 2 ( d d + g )2 (1) where d is the perforated hole diameter, and g is the edge-toedge distance between the holes, as shown in Fig. 2(b). The effective dielectric constant versus the diameter of the holes is also shown in Fig. 2(b). As can be observed, the dielectric constant of the surrounding areas of the DRAs is reduced from 10.2 to 3.1 by making lattices of holes with a diameter of 2.1 mm and a spacing of 0.25 mm. The first step in implementing the proposed CP-ESPAR is designing the driven element with a good CP performance. There are several techniques reported in [22] for achieving circular polarization. Here, we cut a narrow crossedslot with unequal arms in the ground plane [23] to have circular polarization. It is worth noting that according to [24], the fundamental mode of the proposed DRA is TE11y and the presence of the crossed slot on the ground splits this mode into two near-degenerate orthogonal modes to form CP polarization [23]", + " Once the driven DRA is designed, the second step is to place the parasitic DRAs close to the driven element. To this end, we propose to place the parasitic elements based on the sequential feeding technique, which is capable of improving the CP performance in terms of bandwidth and crosspolarization [20]. In this technique, the antenna elements and their corresponding feeding phases are rotated by a particular angle as presented in [20]. For the rotation, we can adopt one of the approaches presented in [27, Fig. 2]. In this paper, the proposed structure shown in Fig. 1 is used. The parasitic elements are arranged as a 2\u00d72 square lattice with angular orientations of 0\u00b0, 90\u00b0, 180\u00b0, and 270\u00b0 to enhance the coupling between the excited and the parasitic elements. To clarify the remaining steps of the design, a network circuit analysis is performed as shown in Fig. 3. According to Fig. 3(a), the CP-ESPAR can be modeled as a five-port network, where its center element (port 1) is connected to the source, and the parasitic antennas (ports 2\u20135) are terminated with reactive loads" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002642_j.mechmachtheory.2017.02.003-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002642_j.mechmachtheory.2017.02.003-Figure5-1.png", + "caption": "Fig. 5. Detail of the actuation and transmission system.", + "texts": [ + " In particular, the most relevant result is that the planet carrier revolution starts automatically when the robot hurts an obstacle and the contact friction is able to stop the wheels rotation. For wheelchair applications, the same locomotion unit structure has been used but a different actuation system has been implemented [29] . Due to safety issues, all the degrees of freedom must be controlled and thus two different motors are used to control independently the wheels rotation and the planet carrier revolution. The adopted architecture is represented in Fig. 5 . Both planet carriers are connected to the same motor (Mp) in order to have a synchronous rotation while two different motors (Ms) are used to control the solar gear rotation of each locomotion unit. The wheelchair behavior is affected by the relative positions and connections between the functional elements previously introduced. Indeed, starting from the same functional elements and keeping in mind the considerations done at the end of the introduction, several structures can be designed. The innovation presented in this paper is the introduction of the cam mechanism between the wheelchair frame and the seat in order to filter the oscillation introduced by the locomotion unit motion" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003845_8.467637-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003845_8.467637-Figure2-1.png", + "caption": "Fig. 2. Nonoffset main reflector geometries (y = 0 plane) (a) scanning configuration using the lower section of the tilted parabola, (b) using the upper parabola section.", + "texts": [ + " Lorenzo are with the Department f, vertex at the Origin Of the uys system, and eastem University, Boston, MA 02115 USA. of Tecnologias de las Comunicaciones, Universidad de Vigo, Spain. IEEE Log Number 9414058. 0018-926)095$04.00 0 1995 IEEE GARCIA-PIN0 et al.: SHAPED DUAL-REFLECTOR ANTENNA 1023 parabola is i. Two different configurations of the central section of the reflector can be obtained by choosing the upper (x > 0) or lower (x < 0) parts of the parabola and joining the selected portion with its specular image about to the Zaxis. Fig. 2(a) shows the reflector central section geometry when the lower part of the parabola is chosen. This is the configuration adopted in [ 121-[ 141 where effectively more than half of the main reflector is efficiently illuminated by the feed, increasing losses due to phase aberrations but also increasing the percentage of illuminated aperture, yielding a net improvement of the antenna efficiency until a balance between these effects is achieved. The geometry adopted in this paper is the one represented in Fig. 2(b) due to the subreflector location considerations discussed in Section 111. In both structures, a feed located at F (F\u2019) can be used to produce a beam in the d (2\u2019) direction by illuminating the appropriate surface section. As the normal to the surface at 0 is coincident with the axis of symmetry 2, first derivative continuity of the reflector surface at the central point is ensured. To construct an offset reflector surface, the two parabolas forming the reflector central section are placed in the y = yoff plane as described in [14]", + " W E TILTING FLAT SUBREFLECTOR When hyperboloidal or ellipsoidal subreflectors are used with a stationary feed, optimum scanning conditions are achieved when the subreflector is rotated about its feed-focal point because this rotation moves the virtual point source feed along the focal arc (which is close to the optimum scan-gain contour described in [4]) and scans the beam. The feed boresight direction must be repointed, however, during scanning. In comparison, a flat subreflector has a simpler geometry and can be rotated without requiring feed repointing. The optimum planar subreflector location for any scan angle is the one that images the virtual source in the focal arc to the stationary feed point. The geometry of Fig. 2(b) is adopted in this paper because a smaller subreflector can be used to intercept the rays propagating from the feed to the main reflector. This subreflector can be tilted to redirect rays for the various scanning directions. Fig. 5(a) shows an example of the extreme scanning operation of a non-offset antenna configuration = 0). The subreflector plane is perpendicular to the line connecting the feed point source A and the focal point F at the midpoint of that line. The limits of the subreflector are determined by the efficiently illuminated reflector portion (between E and I)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001801_j.triboint.2016.11.027-Figure10-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001801_j.triboint.2016.11.027-Figure10-1.png", + "caption": "Fig. 10. (a) Schematic diagram, (b) picture, and (c) front view of the test rig to measure lift-off speed, drag torque, and temperature of NSFB.", + "texts": [ + " However, the direct damping coefficients of the NSFB with axial preload are much larger than those of the NSFB without axial preload, which was shown in Fig. 9(b). The results confirm the contributions of the axial preload to the improvement of the dynamic damping coefficients. Similar experimental results regarding the effect of axial preload on the equivalent viscous damping coefficient and loss factor of the underlying elastic structure were also presented in Ref. [27]. Bearing test is a critical step before it was applied to support the rotor. Fig. 10 shows a schematic view of a test rig to determine the feasibility of the NSFB and measure the lift-off speed, drag torque, and temperature of the NSFB. As shown in Fig. 10(a), a rigid rotor is supported on a holder that consists of two high-speed ball bearings. The ball bearings are lubricated by a set of air/oil lubricating system. The test NSFB floats on one end of the rigid rotor and the other end of the rotor was connected to a high speed motor through flexible coupling. The operating speed of the motor can be controlled by a converter. The maximum operating speed of the high speed motor and the holder can reach 40 krpm. Both the motor and the ball bearing holder are cooled by a water cooling system. The parameters of the test NSFB are listed in Table 1 and the nominal radial clearance is designed to be 170 \u00b5m. Fig. 10(b) shows a picture of the entire test rig and Fig. 10(c) shows the front of the test rig. A rod is fastened to the test NSFB to transfer the drag torque of the bearing to a load cell, which can also prevent the bearing sleeve from rotating. A lateral load was applied on the NSFB by using a steel wire and a pulley. The rotational speed was recorded by a photoelectric tachometric transducer mounted on a magnetic base. Moreover, the foil temperature of the test NSFB was measured by a thermoelectric couple that is located in the gap between the adjacent springs and fixed onto the back surface of the top foil", + " The reason is that the gas film is built between the journal and bearing surface, thus few heat is generated and the bearing achieves heat balance. The total temperature rise is approximately 1.5 \u00b0C, which indicates that the test NSFB has a superior start-stop characteristic. Moreover, the foil temperature stabilizing at 21.5 \u00b0C confirms the feasibility of the test NSFB. Fig. 13 shows a schematic diagram of the bearing static load performance measurement test rig. The test rig is same as the test rig in Fig. 10 except the load device. The midplane of the tested NSFB is connected to a load device, which consists of a micrometer head, two Rotor Load cell Eddy current sensorMicrometer head rods, and a spring, through a strain gauge type load cell. The connection prevents the bearing from rotating caused by the drag torque between the bearing and the rotor. The spring between the NSFB and the micrometer head is used to transform the load and confirm that the applied load is centred at the bearing. The load cell is installed between the test NSFB and the spring to exclude the influence of the spring on the stiffness of the test NSFB" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001472_9781118354179.auto089-Figure17-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001472_9781118354179.auto089-Figure17-1.png", + "caption": "Figure 17. Clutch assembly with an automatic internal wear adjustment mechanism. (a) The principle of wear adjustment mechanism. (b) Gearbox and engine side views of the clutch assembly. (Reproduced with permission from ZF Friedrichshafen AG.)", + "texts": [ + " The latter can lead to the fingers touching the bearing at all times, causing excessive bearing wear and reduction in clutch torque capacity. To prevent this, most clutch control Encyclopedia of Automotive Engineering, Online \u00a9 2014 John Wiley & Sons, Ltd. This article is \u00a9 2014 John Wiley & Sons, Ltd. DOI: 10.1002/9781118354179.auto089 Also published in the Encyclopedia of Automotive Engineering (print edition) ISBN: 978-0-470-97402-5 mechanisms have manual or automatic bearing clearance adjustment. There is however another solution\u2014Figure 17 shows a clutch assembly design (XTEND, from ZF Sachs), with an internal adjustment mechanism, ensuring constant normal force and spring fingers/thrust bearing clearance. As a result, such a design should provide constant maximum torque and constant clutch pedal travel in all conditions, for the entire life of the assembly. The diaphragm spring is the most commonly used spring type in today\u2019s road vehicles, and is also known as belleville, disc, or conical spring. The spring (any spring type) is a crucial part of the clutch, very important in providing the torque and complex to manufacture and install" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003630_6.1973-885-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003630_6.1973-885-Figure5-1.png", + "caption": "Fig. 5 Rudder surfaces", + "texts": [], + "surrounding_texts": [ + "STOL RIDE CONTROL FEASIBILITY STUDY\nG. 0. Thompson,* D. L. Eslinger,t C. K. Gordon,** R. 0. Dodson,tt The Boeing Company, Wichita, Kansas\nSummary\nA study to determine the feasibility of developing a ride smoothing control system for a passenger STOL aircraft has been performed for NASA Langley Research Center by Boeing-Wichita, with de Havilland Aircraft of Canada, Limited, as the principal subcontractor. The de Havilland DHC-6 Twin Otter was selected for the feasibility study, since it is the only STOL aircraft presently certificated and in use by a number of United States air carriers.\nThe study indicated that a ride control system that significantly reduces vertical and lateral accelerations can be practically implemented on the Twin Otter with minimum airplane performance degradation. The system uses symmetrical ailerons, elevator, rudder and spoiler control surfaces with accelerometers and rate gyros for motions sensors.\nI. lntroduction\nA ride smoothing system may be required for passenger acceptance of future low to moderate wing loading STOL vehicles because of the flight dynamic characteristics and operational environment. Consequently, development of ride smoothing system technology is an important element of NASA's advanced STOL transport technology activities.\nConcern over the unwanted response of aircraft to atmospheric turbulence dates back to the initial NACA report by Hunsaker and Wilson in 191 5.(' ). However, indepth research regarding the use of active control systems for ride smoothing did not begin until the late 1940's and early 1950's. The work at NACA Langley Aeronautical Laboratory by Phillipsand raft(^) discusses some of this early research. lntroduction of the jet transport, with its higher wing loading and cruise altitude, led to an improved ride without an active ride control system and thus greatly reduced the immediate need for ride smoothing research. However, recent feeder airline considerations of STOL aircraft, which f ly low with low to moderate wing loading, have renewed interest in ride smoothing efforts.\nThe Series 300 Twin Otter, Figure 1, has a maximum gross weight of 12,500 pounds and carries 20 passengers and a crew of\nFig. 1 de Havilland DHC-6 Twin Otter\ntwo. A wing loading of approximately 30 lb/ft2 makes it susceptible to ride quality degradation in severe turbulence. The Twin Otter is powered by two Pratt and Whitney PT6A-27 turbbprop engines. Maximum cruising speed is 180 KTAS and maximum operating altitude i s 10,000 feet because of an unpressurized cabin. Maximum range is 755 nautical miles; however, the average flight time is less than 45 minutes for normal commuter service in the United States. Approximately 90 of the 366 Twin Otters in service operate in the United States under FAA Part 135 rules. A three-view illustration of the Twin Otter is shown in Figure 2.\nI I. Design Criteria\nA survey was conducted of existing data on operational profiles of the DHC-6 Twin Otter. Based on this survey, typical climb, cruise and landing approach conditions were selected for design flight conditions, as tabulated in Table 1\nTable 1 Study flight conditions\nCondition\nAirspeed (KIAS)\nAltitude (ft)\nGross weight (Ib)\nFlap position (deg)\nCG location (% MAC)\nGlide slope (deg)\nClimb Cruise Landing Approaci-\nThis study was conducted for the NASA Langley Research Center, Contract NASl-11683, under the direction of D. W. Conner. 'Chief of Flight Controls Staff, Member AlAA **Specialist Engineer, Flight Controls Staff tSenior Specialist Engineer, Flight Controls Staff ttspecialist Engineer, Aerodynamics Staff, Member AlAA\nD ow\nnl oa\nde d\nby U\nN IV\nE R\nSI T\nY O\nF N\nE W\nS O\nU T\nH W\nA L\nE S\n(U N\nSW )\non O\nct ob\ner 3\n0, 2\n01 7\n| h ttp\n:// ar\nc. ai\naa .o\nrg |\nD O\nI: 1\n0. 25\n14 /6\n.1 97\n3- 88\n5", + "A random turbulence intensity with an exceedance probability of 0.01 was selected as a design criterion. This probability level corresponds t o an rms gust velocity of 7.0 fps for the design flight conditions.(3,4) Atmospheric turbulence was modeled with a von Karman spectrum. A scale length (L) of 2500 feet was used for the climb and cruise conditions, and 500 feet for the landing ~ o n d i t i o n . ~\nWell-defined ride quality performance criteria for aircraft do not exist presently. However, a Symposium on Vehicle Ride ~ u a l i t y ~ held at the NASA Langley Research Center, July 6 and 7, 1972, produced indications of approximate human comfort motion boundaries. Based on the symposium discussions, conservative motions levels were established for ride control system performance goals, rather than absolute requirements.\nPrimary goals were to reduce vertical acceleration at each flight condition to 0.030 g rrns or less and lateral acceleration to 0.015 g rms or less, at all crew and passenger stations, while subjecting the airplane to the design random turbulence. I n addition, angular accelerations and rates were not to exceed existing free airplane values.\nPitch short period handling qualities were evaluated qualitatively with control column step inputs by comparing pitch rate and normal acceleration of the aircraft and ride control system t o the free airplane response. A minimum damping ratio of 0.04 was selected as a design goal for the phugoid mode.\nLateral-directional handling qualities from M l ~ - ~ - 8 7 8 5 B ( ~ ) (Level 1 for light airplanes) or existing values, whichever were less, were used as design goals.\nIn the area of safety, the system was designed to provide (1) adequate handling qualities and safety for continued flight following a single engine failure, total hydraulic or electrical power failure or a single ride control surface hardover, and (2) safe maneuvering and landing capability following two engine failures.\nI I I. Control Surface Confiaurations\nTrade studies were conducted t o determine control surface configurations providing sufficient authority for ride control, while meeting manual flying qualities criteria. Based on results of these studies, the ailerons, elevators and rudder were each split spanwise into two segments for ride control and manual flight control, as follows:\nManual Flight Control Ride Control Segment Segment\n-Percent Span- -Percent Span-\nAileron 60 40\nElevator 80 20\nRudder 70 30\nManual segments are controlled through existing mechanisms. Ride control segments are controlled by electrohydraulic power actuators that received electrical position command signals from both pilot manual and ride control command.\nSpoiler control surfaces, operating from a biased position, were added t o augment the ailerons for direct l i f t ride control during landing approach.\nThe aileron and spoiler configurations are illustrated in Figure 3. The selected elevator and rudder configurations are\ncompared to the existing configurations in Figures 4 and 5, respectively.\nThe biased, 10 percent chord, 17 percent semispan spoiler has good effectiveness and incremental l i f t characteristics, as shown in Figure 6. The spoiler is biased at 12 degrees and\nD ow\nnl oa\nde d\nby U\nN IV\nE R\nSI T\nY O\nF N\nE W\nS O\nU T\nH W\nA L\nE S\n(U N\nSW )\non O\nct ob\ner 3\n0, 2\n01 7\n| h ttp\n:// ar\nc. ai\naa .o\nrg |\nD O\nI: 1\n0. 25\n14 /6\n.1 97\n3- 88\n5", + "deflected f10 degrees when the ride control system is engaged. The spoiler is not required during the climb and cruise conditions. Location of the outboard edge of the spoiler was limited by the inboard edge of the aileron, so that the aileron effectiveness would not be disturbed. The inboard edge of the spoiler was determined from buffet and pitch tr im effects due to the spoiler wake affecting the flow over the horizontal tail. For the extreme center-of-gravity conditions, only 0.6 degrees of elevator is required t o tr im the pitching moment resulting from the symmetrical spoiler deflection at the bias position.\nEffect of the 12 degree spoiler bias on Twin Otter stall speed is estimated from zero to 3.5 percent increase, depending on how the biased spoiler affects wing stall characteristics. A t most, CLmax is decreased by the incremental CL at the bias position in Figure 6. This corresponds t o a 3.5 percent increase in stall speed for a 40 degree flap landing condition.\nIV. Ride Control System Synthesis\nAnalytical studies were conducted on digital and hybrid computers to synthesize the vertical and lateral ride control systems. Airplane dynamics were described by linear, rigid-body, small perturbation equations of motion, with actuator nonlinearities and l i f t growth dynamics included, as appropriate.\nVertical Ride Control Svstem Svnthesis\nThe vertical ride control system consists of three feedback loops: center-of-gravity vertical acceleration driving symmetrical ailerons and spoilers, and pitch rate commanding the elevator, as shown in Figure 7. The spoilers are active only in the landing phase. The acceleration feedback accomplishes acceleration reduction, and the pitch rate feedback adjusts handling qualities. The high pass filters (washouts) minimize \"fighting\" of pilot inputs. The low pass filter in the pitch rate feedback provides proper phasing for handling qualities.\nFLIGHT CONDITION\nW K i' 2 . 0 g\nELEVATOR . GAIN SCHEDULING\n1 20 j 6~ I COMMAND . S i 20 DEG\nACTUATOR AIRPLANE ~ L E R O N . COMMAND . ACTUATOR\nFig. 7 Vertical ride control system block diagram\nFor this feasibility study, the aileron feedback was gain scheduled. The scope of this study did not allow cost and performance trades of a gain scheduled system versus a constant gain system.\nEffectiveness of center-of-gravity acceleration feedback to the ailerons was determined for each flight condition, as shown in Figure 8, for the cruise condition. This data was digitally computed by power spectral methods for acceleration feedback, without feedback filter or actuator dynamics. A gain of three deg/ft/sec2 was selected for the cruise condition, using full span aft flaps and ailerons. Later, when the configuration was changed to partial span ailerons, the gain was increased to 10.7 deg/ft/sec2\nto compensate for the decrease in control surface l i f t effectiveness.\nThe spoilers, active only in the landing phase, are commanded by the same sensor and filters as the ailerons.\nFeeding back acceleration decreases the short period natural frequency and increases damping as shown in Figure 9. Consequently, this feedback makes the airplane response to pilot commands sluggish. Closing the pitch rate feedback loop to the elevator through a lag filter restores the root to the free airplane root location, as shown in Figure 10.\nCRUISE CONDITION \\ a- '\n-3 -2 -1 10\nD ow\nnl oa\nde d\nby U\nN IV\nE R\nSI T\nY O\nF N\nE W\nS O\nU T\nH W\nA L\nE S\n(U N\nSW )\non O\nct ob\ner 3\n0, 2\n01 7\n| h ttp\n:// ar\nc. ai\naa .o\nrg |\nD O\nI: 1\n0. 25\n14 /6\n.1 97\n3- 88\n5" + ] + }, + { + "image_filename": "designv6_24_0000194_s12239-015-0047-9-Figure9-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000194_s12239-015-0047-9-Figure9-1.png", + "caption": "Figure 9. Planar model including the tire forces.", + "texts": [ + " To calculate the value of the desired yaw moment, the lateral tire forces need to be estimated. L1 ------ c2\u03b4 f 2 d2\u03b4\u00b7 f 2+ cd ---------------------------= = Index for Stability \u03b3desired for agility \u03b3threshold ----------------------------= \u03b3threshold \u03bcg vx -----=\u239d \u23a0 \u239b \u239e Izz\u03b3\u00b7 lfFyf lrFyr wr 2 ---- Fxrr Fxrl\u2013( ) Mz+ +\u2013= s \u03b3 \u03b3d \u03c1\u03b2+\u2013= if \u03b2 \u03b2th then \u03c1 0=\u2264 \u03b2th 10o\u2264 if \u03bc 0.9 ,= \u03b2th 4o\u2264 if \u03bc 0.35= ss\u00b7 \u03b7s2 0 s\u00b7 \u03b7s\u2013=,<\u2013= Mz lfFyf\u2013 lrFyr wr 2 ---- Fxrr Fxrl\u2013( )\u2013 Izz\u03b3\u00b7 Izz\u03c1\u03b2\u00b7\u2013 Izz\u03b7s\u2013+ += wr 2 ---- Fzrr Fxrl\u2013( ) 4.2. Estimator of Lateral Tire Forces Figure 9 shows a planar model including the tire forces and yaw moment control input. This model can represent the lateral dynamics of the vehicle including rotational and translational motion. The corresponding equations of the lateral dynamics are as follows: (13) Equation (13) has two unknown variables Fyf and Fyr that are the front and rear lateral tire forces, respectively. Therefore, as there are two equations for two unknown variables, those lateral tire forces can be determined from equations (13) as follows (Fukada, 1999): (14) Equation (14) needs to be checked for the performance as estimators for the lateral tire forces" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001127_icps.1992.163390-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001127_icps.1992.163390-Figure7-1.png", + "caption": "Figure 7. Space-Qualified SolidInsulation Plug-In Bus Duct Concept.", + "texts": [ + " Finally, there is a fairly high concentration of ions floating around in LEO, which are attracted to charged objects such as bus duct conductors and can initiate short circuits if electrical clearances are inadequate. On the positive side, bus duct in space is immune to moisture, dirt, and most sources of physical damage. On the whole. we expect that qualified bus duct will have considerably better reliability and life expectancy in LEO than industrial duct does in the typical factory environment. To help to illustrate the feasibility of the zero-based systems, we developed two alternative design concepts for the Freedom plug-in bus duct which respond to the design requirements above. Figure 7 shows aceramic-insulatedduct,andFigure 8 illustratesaductdesigncombining ceramic bus supports and vacuum main insulation. 6.0. Evaluationand C) The principal measures of performance for a ultra-compact, highsecurity, isolated power system such as the Space Station's are power quality (comprising voltage regulation and wavefore purity),reliability, availability, maintainability, efficiency, and weight. In the following sections we evaluate the 1990 baseline and the simplified systems with respect to these criteria" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003503_j.actamat.2007.06.033-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003503_j.actamat.2007.06.033-Figure7-1.png", + "caption": "Fig. 7. Schematic depiction of the effect of a fiber-textured material on the associated maximal MDF. The probability of finding a cone angle misorientation between DU dDU/2 and DU + dDU/2 is graphically illustrated by the intersected areas of the dashed ring with the filled ring (see Fig. 2). For increasing randomness, misorientations with larger values of D/1 become more likely to sample, while small values of D/1 become less probable.", + "texts": [ + " Therefore, as a polycrystal becomes increasingly random, the two populations of most probable misorientations approach each other. In the limit of no texture, r = 1, the two populations of cone angles become indistinguishable. Furthermore, the polar angle misorientations, D/1 = \u00b1 p/2, become more likely to occur than other values, and the cone angle misorientation DU = p/2 is the most probable; moreover, even though all the crystallographic orientations are equally probable, all the crystallographic misorientations are not. Graphically, the effect of fiber texture on the maximal MDF is illustrated in Fig. 7. Here, the accessible misorientations with cone angles between DU dDU/2 and DU + dDU/2 correspond to the intersection of the dashed ring and the filled ring in the stereographic projection plane (e.g., Fig. 7). For a polycrystal satisfying Eq. (14), as the degree of texture in the polycrystal decreases, the radius and the thickness of the filled ring increases, thus greater values of D/1 are intersected, while making small values of D/1 less likely to occur. Furthermore, the values of cone angle misorientations that can be sampled become larger with decreasing texture, and the probability of sampling small values of polar angle misorientations decreases. This process makes intermediate values of cone and polar angle misorientations more probable and contributions from an intermediate region less probable" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001387_acdt.2018.8592940-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001387_acdt.2018.8592940-Figure8-1.png", + "caption": "Fig. 8. Axial velocity of the flow at propeller disk of the actual model.", + "texts": [], + "surrounding_texts": [ + "The author would like to thank the aeronautical engineering research team at the Defence Technology Institute for their helpful discussions around the topic of the paper." + ] + }, + { + "image_filename": "designv6_24_0000849_biorob.2008.4762805-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000849_biorob.2008.4762805-Figure4-1.png", + "caption": "Fig. 4. Mass-spring-damper model for cable structure.", + "texts": [ + "u j (5) IKii and IKi j design respectively the sum of tensor Kii associated with the tetrahedra adjacent to node i and with the tetrahedra adjacent to edge (i, j), O(Pi) is the neighborhood of vertex Pi. These tensors, depending only on the rest geometry and Lames coefficients, are constant and can be pre-computed. B. Cable behavior for the tensegrity structure In this section, we present the description of the cable behavior about our tensegrity model (Fig.2). These cables are assumed to be have as viscoelastic mass-spring-damper. Each cable is modeled with two masses interconnected via spring and damper in parallel (Voigt-Kelvin model, Fig.4). In the local frame, the relation between the stress and the strain can be written as follows: \u03c3 = E \u03b5 +\u03b7 \u03b5\u0307 (6) with \u03c3 = F S and \u03b5 = l\u2212 l0 l0 , where F : is the applied load in the extremity of the cable, with l and l0 and S are respectively the resting length of the cable, the initial length and the section of the cable. we replace these parameters in (6), we obtained : Fcab local = ( E S (l\u2212 l0) l0 + \u03b7 l\u0307 l0 ) x (7) In the global coordinates, for the tensegrity structure, we can expand this equation easily and obtain the force applied in each node by summing the forces contributed by all nodes connected (Fig.4). It can be written as follows: Fcab i = \u2211 j {( E S ( |pi j| Li j \u22121 ) +\u03b7 ( vi j . pi j Li j |pi j| )) pi j |pi j| } these cables do not transmit the forces in compression. The force applied to an extremity of the struts of the tensegrity structure, is the sum of the forces exerted by the struts and the cables (Fig.4). 1) Dynamic Model: We suppose that the mass of the cell is lumped at the vertex. The dynamic law of motion can be written for a vertex i as: Mi U\u0308 i + \u03b3 U\u0307 i = F i (8) where Mi is the mass and \u03b3 a numerical structural damping coefficient. We choose a classical second order central finite difference scheme as a good compromise between simulation in real time and accuracy: OPi(t +h) = K1 ( F i(t)+K2 OPi(t)\u2212K3 OPi(t\u2212h) ) (9) where h is the time step; K1 = h2 Mi+\u03b3 i h , K2 = 2Mi+\u03b3i h h2 , K3 = Mi h2 A" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002584_1999-01-0782-Figure11-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002584_1999-01-0782-Figure11-1.png", + "caption": "Figure 11. Second generation prototype", + "texts": [ + " 1000 ][ *SD(u*RR)L* M + = 6 The second generation Kuhl Wheel was an attempt to eliminate some of the weaknesses of the first generation design while continuing to evolve the Kuhl Wheel technology. The improvements needed included a simplified the connection of the spokes to the hub, and a reduction of the overall parts count. In addition, a better method of clearing the brake assembly while accommodating different offsets was necessary. To satisfy these objectives, a second generation prototype was designed and built and is shown in Fig.11. The most significant change made to the overall design was the use of a connection which more rigidly affixed the hub to the spoke. By folding the spokes up into a channel as shown, lug bolts which attach the spoke to the hub could now bear directly on the individual spokes. This direct bearing we believed would help to provide additional reinforcement to some of the more critically stressed areas in this region while reducing the complexity of the attachment. ANALYSIS OF SECOND GENERATION DESIGN \u2013 To get a better understanding of what stresses in the wheel would be like on the dynamic cornering fatigue test, an FEA analysis was performed" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003717_induscon.2016.7874600-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003717_induscon.2016.7874600-Figure1-1.png", + "caption": "Figure 1 - Construction of the machine rotor.", + "texts": [ + " (3) With: = rotor velocity [rpm]; = electric frequency [Hz]; = number of poles; = motor electric position; = motor mechanical position; The three-phase motor developed for this study has an axial structure. The stator has 9 windings connected in series three by three, which possess plastic supports placed in a circle. The wire used in the windings is the 18AWG, 0.815mm2. The rotor is formed by 16 Neodymium magnets fixed in two metallic bases using epoxy resin and glass fiber. At each base, the poles are arranged alternately, as shown in Figure 1. Figure 2 illustrates the structural scheme developed. Each stator phase is represented by a different color. Figure 3 shows the final assembled machine in laboratory. In order to supply the machine with a sinusoidal wave, it is necessary that each branch of the three-phase inverter \u2013 shown in Figure 4 \u2013 to be triggered independently. Analyzing one of the inverter branches, S1 and S2 for example, each switch receives a complementary command signal, which avoids a short circuit on the power supply Vcc" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000349_s10665-004-7788-1-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000349_s10665-004-7788-1-Figure6-1.png", + "caption": "Figure 6. Experimental setup.", + "texts": [ + " It can then be argued that, compared with the ZV shaper, the ZVD shaper not only is more tolerant of parameter fine-tuning errors in suppressing the main vibration mode, but also reduces the amplitudes of the surviving modes to a greater extent. So, the ZVD input-shaping scheme generally is preferable in practice. Here we present some experimental results that demonstrate the effectiveness in residual vibration reduction of the various input-shaping schemes discussed above. For a more detailed discussion of the experimental procedures and results, the reader is referred to the experimentally oriented paper of Chen et al. [12]. As shown in Figure 6, the test structure is a stainless-steel beam having a length of 155 mm, width of 13\u00b76 mm, and thickness of 1\u00b72 mm. One end of the beam (the tip of the cantilever) is free to vibrate, while the other end (the base of the cantilever) is clamped on a piezoelectric actuator (Piezo Jena PA100-12, bandwidth \u223c100 Hz). The base movement of the cantilever is then controlled by digital signals generated by a Pentium III 933 MHz personal computer. A capacitive displacement sensor (MTI ASP50, having a dynamic range of 1\u00b725 mm and bandwidth of 5 KHz) is used to pick up the displacement of the flexural cantilever" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002775_mnl.2015.0198-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002775_mnl.2015.0198-Figure1-1.png", + "caption": "Fig. 1 SEM image of an assembled BCP with integrated bimorph actuators for fine tune angle adjustment", + "texts": [ + " There are two types of out-of-plane structures: those that are designed using hinges and other complex locking structures and those that are hingeless and use their own deformation as the locking mechanism to maintain their assembly. In the first case, a hinged plate is manually lifted out of plane using microprobes and then secured to a fixed position with other hinged structures. Therefore, their assembly process is challenging [8, 9] and requires a skilled operator. Moreover, the hinged structures require a higher number of physical layers in the fabrication process. In the second case, hingeless structures such as buckled cantilever platforms (BCP, Fig. 1) [1, 3, 10] and Tsang suspensions [11, 13, 14] require a simpler procedure. They are essentially compliant mechanisms that are deformed and make use of the elastic properties of the composing materials to maintain this deformation using stoppers or a clever structural design (i.e. using their own reaction forces) [1, 11, 13\u201315]. The procedure to assemble a BCP is depicted in Fig. 2. A microprobe station needle or a wire-bonder head is used to push the front edge of the cantilever towards the structure\u2019s main anchors until the cantilever beams are buckled out of plane causing the attached plate to rise", + " Similarly, thermal accelerometers [11] and microheaters for gas sensing [4] have taken advantage of the BCP\u2019s thermal isolation to operate at a lower power. Magnetic field induction sensors [10] have also been realised in these platforms. In this Letter, we report on BCPs that have an integrated set of MEMS thermal bimorph actuators. This integration allows a precise dynamic tuning of the plate\u2019s angular position beyond its original assembly angle, and additionally allows oscillation of the structure at low frequencies (Fig. 1). A typical MEMS bimorph actuator is a transducer that generates motion when a bilayer structure formed by two materials with different coefficients of thermal expansion is heated. As the temperature increases, the material with the highest coefficient of thermal expansion will undergo a greater increase in length, causing the bilayer structure to bend 545 & The Institution of Engineering and Technology 2015 towards the material with the lower coefficient of thermal expansion [16, 17]. If one of these materials is conductive, passing a current through it will increase its temperature due to ohmic heating, making it act as both a component of the bimorph structure and the heating element for the actuator" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002661_iros.2011.6094731-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002661_iros.2011.6094731-Figure1-1.png", + "caption": "Fig. 1. CAD model of the quadcopter \u201cPelican\u201d. The front and rear propellers are spinning clockwise, the left and right anticlockwise.", + "texts": [ + ". INTRODUCTION Quadcopter systems are more and more used as platforms for experiments in the field of navigation and control. Basically, the quadcopter is a helicopter driven by four rotors, symmetric to the center of mass. The control of the quadcopter is done by changing the rotation speed of the propellers. As shown in Fig. 1, two of the propellers are spinning clockwise and two counterclockwise. A command in the roll axis is e.g. generated by speeding the left propeller up and the right down without affecting the other axes. A yaw command around the vertical axis is done e.g. by speeding front and rear up and left and right down. The total thrust as well as the angular moments around the other axes stay constant. This concept is described in more detail in [1]. With the ongoing popularity of these systems the need of new platforms with higher payload capabilities and a flexible design has grown", + " The theoretical weight limit of the Pelican was 1.2 kg including the solar cell array and a backup battery due to a total average power limit of 180 W. The weight of the solar cell arrangement, mechanical mounts, charging and monitoring electronics accumulated to 350 g. A 1350 mAh 3 s LiPo battery with 100 g weight was chosen. Thus the final configuration had a weight of 1050 g, leaving enough margin to charge the battery. The mechanical design is motivated by fast changing requirements for the payload units. Fig. 1 shows a CAD model of the complete system and Fig. 4 the design of the central part, which is inspired by computer racks. It consists of vertical carriers with slots for horizontal payload inserts. The complete frame is built with 2-D milled carbon fiber and carbon fiber-balsa-wood sandwich material. It is assembled only by screws without the need of gluing components, enabling very fast component changes or repairs. The FCU features a combination of a high speed IMU and two 32-bit 60 MHz microcontrollers for the flight control algorithms" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000460_icem49940.2020.9270851-Figure26-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000460_icem49940.2020.9270851-Figure26-1.png", + "caption": "Fig. 26. Non-uniform rotor yoke.", + "texts": [], + "surrounding_texts": [ + "[1] J. Pyrh\u00f6nen, T. Jokinen and V. Hrabovcov\u00e1, Design of Rotating Electric Machines, Second edition, Wiley (United Kingdom), p. 201, 2014. [2] G. Dajaku, S. Spas, X. Dajaku, D. Gerling, \u201cAn Improved Fractional Slot Concentrated Winding for Low-Poles Induction Machines\u201d, 19th International Conference for Electrical Machines (ICEM), Germany, September 2016, pp. 116-121. [3] K. S. Garner and M. J. Kamper, \u201cReducing MMF harmonics and core loss effect of non-overlap winding wound rotor synchronous machine (WRSM),\u201d IEEE Energy Conversion Congress and Exposition (EC- CE), Cincinnati, OH, 2017, pp. 1850-1856. [4] M. Barcaro, N. Bianchi and F. Magnussen, \"Rotor flux\u2013barrier geome- try design to reduce iron losses in synchronous IPM motors under FW operations\", IEEE International Electrical Machines and Drives Conference (IEMDC), Miami, FL USA, 2009. [5] N. Bianchi, S. Bolognani, D. Bon and M. Dai Pre, \u201cRotor flux-barrier design for torque ripple reduction in synchronous reluctance and PM- assisted synchronous reluctance motors\u201d, IEEE Transactions on Industry Applications, Vol. 45, No. 5, pp. 921-928, May - June 2009. [6] Magnetic materials - part 1-1: classification, International Standard IEC 60404-1:2000. [7] C. L. Xia, L. Y. Guo, Z. Zhang, T. N. Shi and H. M. Wang, \u201cOptimal Designing of Permanent Magnet Cavity to Reduce Iron Loss of Interi- or Permanent Magnet Machine\u201d, IEEE Transactions on Magnetics, Vol. 51, 2015. [8] C. C. Hwang, C. M. Chang and C. T. Liu, \u201cA fuzzy-based taguchi method for multi-objective design of PM motors\u201d, IEEE Transactions on Magnetics, Vol. 49, pp. 2153 \u2013 2156, 2013. [9] Z. Q. Zhu, D. Wu and W. Q. Chu, \u201cInfluence of local magnetic saturation on iron losses in interior permanent magnet machines\u201d, International Conference on Electrical Machines (ICEM), Lausanne, Switzerland, 2016, pp. 1822 \u2013 1827. [10] Z.Q. Zhu and D. Wu, \u201cOn-load voltage distortion in fractional-slot interior permanent magnet machines\u201d, IEEE Transactions on Magnetics, Vol. 51, pp. 1 \u2013 9, 2015. [11] K. Yamazaki, Y. Kato, T. Ikemi and S. Ohki, \u201cReduction of Rotor Losses in Multilayer Interior Permanent-Magnet Synchronous Motors by Introducing Novel Topology of Rotor Flux Barriers\u201d, IEEE Transactions on Industry Applications, Vol. 50, No. 5, pp. 3185 \u2013 3193, 2014. [12] A. Dziechciarz and C. Martis, \u201cNew Shape of Rotor Flux Barriers in Synchronous Reluctance Machines Based on Zhukovski Curves\u201d, 9th International Symposium on Advanced Topics in Electrical Engineering, Bucharest, Romania, pp. 221 \u2013 224, 2015." + ] + }, + { + "image_filename": "designv6_24_0003230_isatp.2005.1511450-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003230_isatp.2005.1511450-Figure2-1.png", + "caption": "Fig. 2. Examples of contact states representations w.r.t. deformation directions: (a) contact involving the concave side of the deformed element: F\u0303A EB (b) contact involving the convex side of the deformed element: F\u0303A EB .", + "texts": [ + " Let UA and UB 3 indicate two surface elements of the elastic tube A and the rigid object B respectively, and each element can be a face (F ), an edge (E), or a vertex (V ) in particular. We now add new symbols to the expression of a PC between UA and UB : \u2022 If UA and UB are in contact while A does not bend, the PC between UA and UB is indicated as UA\u2212UB (just as in the case of contacting rigid objects). \u2022 If UA and UB are in contact while A bends but UA does not deform, the PC between UA and UB is indicated as U\u0303A\u2212UB . \u2022 If UA and UB are in contact while UA deforms, and if the concave side of UA contacts UB , the PC between UA and UB is indicated as U\u0303A UB (as in Fig. 2a); otherwise, if the convex side of UA contacts UB , the PC between UA and UB is indicated as U\u0303A UB (as in Fig. 2b). Now with the added information, there are more types of PCs between an elastic tube and a rigid object than those between a rigid tube and another rigid object, as shown in Fig 3. Consequently, the types of CFs between an elastic tube and a rigid object are also more than those between a rigid tube and another rigid object. We first introduce how to model the deformation of an elastic tube with only bending behavior with the physics-based \u2019bending beam\u2019 model and then describe how to simulate the contact force and deformation shape change for haptic and graphic rendering" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000777_smasis2010-3636-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000777_smasis2010-3636-Figure2-1.png", + "caption": "Figure 2: Thermally activated reconfigurable system", + "texts": [ + " Since conversion of thermal energy to other forms, such as electric through the use of thermoelectric or pyroelectric materials, necessarily involves high losses due to conversion efficiencies of the materials, it is desirable to investigate a way to utilize the thermal energy directly, directing the thermal energy to desired locations within the vehicle rather than converting it to other forms. Presented are the results of ongoing efforts in the feasibility and design of such a system. A conceptual reconfigurable system proposed in this paper is shown in Figure 2. It is composed of thermal transport system and mechanical actuation system. Thermal energy from the environment is directly transferred without energy conversion to trigger the reconfiguration. Each system is represented using an energy metric to optimize the performance of the entire system, the thermal energy modeling of the heat pipe is described in the following section. Copyright \u00a9 2010 by ASME Downloaded From: https://proceedings.asmedigitalcollection.asme.org on 12/06/2018 Terms of Use: http://www" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002773_12.893614-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002773_12.893614-Figure1-1.png", + "caption": "Figure 1. CERES instrument", + "texts": [ + " The purpose of this paper is to provide a brief overview of the instrument, and then to highlight the functions which must be performed by the flight software. 1 Corresponding author: Kelly K. Teague E-mail: kelly.k.teague@nasa.gov, Telephone: +1-757-864-9623 Earth Observing Systems XVI, edited by James J. Butler, Xiaoxiong Xiong, Xingfa Gu, Proc. of SPIE Vol. 8153, 81531S \u00b7 \u00a9 2011 SPIE \u00b7 CCC code: 0277-786X/11/$18 \u00b7 doi: 10.1117/12.893614 Proc. of SPIE Vol. 8153 81531S-1 Downloaded From: http://proceedings.spiedigitallibrary.org/ on 06/28/2016 Terms of Use: http://spiedigitallibrary.org/ss/TermsOfUse.aspx Figure 1 shows a schematic view of the CERES instrument and Figure 2 is the corresponding exploded view. To determine the Earth\u2019s net radiation flux, CERES employs three detector channels. A shortwave channel is sensitive to reflected sunlight, a total channel measures total radiation, from which the Earth-emitted radiance is computed by subtracting the reflected solar radiance, and a 8 \u2013 12 micron channel for measuring radiance in the longwave window.5 Each channel has a telescope which focuses radiation onto a thermistor-bolometer detector" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001320_861355-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001320_861355-Figure4-1.png", + "caption": "Figure 4 - Velocity diagram of a 'spinning", + "texts": [ + " Velocity diagram of a spinning-rolling contact not transmitting torque. (B) (A) creep. divided into two types: (1) those having con tacts with pure rolling motion and (2) those where, in addition to rolling, the contacts are also subjected to a superimposed spin motion, which is a rotary motion about some axis perpen dicular to the contact plane. As will be shown later, that spinning at the contact has very far-reaching effects on tractive capacity and power losses within the rolling contacts. Figure 4 shows a greatly exaggerated and simplified velocity diagram for a \"spinning rolling\" contact under two conditions: (a) no transmitted torque, shown on the left, and (b) transmitted torque, shown on the right [7]. 861355 5 Figure 5 - Velocity diagram of a 'pure rolling' contact in a traction drive. CREfP_1vq- I 'I Vo it- -- 0 ---\\- V2 ~ \\ \" ! I I\u00a3R OTH CONTACTI / .. ARfA V n I if CONTACT ROl Vo LENGTH WI --- :- V2 I, :..l I \"'- CONTACT WIOTH I >Lf-L< lAl (A) (B) Simplified velocity diagram of a rolling contact not transmitting torque" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000197_elk-1306-206-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000197_elk-1306-206-Figure1-1.png", + "caption": "Figure 1. Solid work of the flywheel design.", + "texts": [ + " As a result of this choice, mechanical friction losses need to be considered in the design. As there are many unknowns, it is required that some of the parameter values are chosen and the rest are calculated. The inertia of the flywheel has been taken as 0.008 kgm2 . The weight of the system now depends on this value as well as the material selection and mechanical design. Aluminum material has been used as the material to build the flywheel. Eventually, a disk-shaped flywheel of 20 cm in diameter and 2 kg in weight was obtained. Figure 1 shows the solid work of the flywheel design. As stated earlier, a minimum value of the angular momentum exists for the attitude control system. This value is Hmin = 0.4 Nms for the satellite used in the design. Minimum speed of the operation is obtained now by using these values. Hmin \u2264 J \u03c9min \u21d2 \u03c9min \u2265 (0.4 Nms)/(0.008 kg.m2) = 50 rad/s = 477.46 RPM Although this value will guarantee the operation of the attitude system, the voltage generated at this speed by the flywheel will be very small to regulate the bus" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002572_978-3-319-76276-0_5-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002572_978-3-319-76276-0_5-Figure1-1.png", + "caption": "Fig. 1 Photographs of failed primary inner spring and arrangement of middle axle suspension spring near middle wheel", + "texts": [ + " A primary suspension mounted between wheelbase and bogie with inclined damper at end axles and linkage to restrict lateral movement at middle axle. The middle axle has an assembly of concentric suspension spring of inner and outer opposite handed spring to acquire load without damper. This paper focuses on premature fatigue failures of inner suspension spring due to dynamic effect. The failure region of middle axle primary inner suspension spring and arrangement for mounting on middle axle housing is shown in Fig. 1. The middle axle primary inner suspension spring is subjected to variable loads and hence the fatigue analysis approach is used to investigate the failure of the spring using finite element method. A progression of forward and reverse loading portrayed a fatigue; where plasticity is initiated in each cycle. The fatigue life of a suspension spring can be communicated as the number of loading cycles for the initiation of crack and the number of cycles propagates that crack to failure [5]. A computational model for fatigue examination of suspension spring has been exhibited", + " Hence S-N curve has been plotted for ultimate shear strength of 1152.4 N/mm2 and endurance shear strength of 395.6 N/mm2 as shown in Fig. 5. Considering fatigue strength factor as unity, an equivalent alternating shear stress is determined and it is given in Table 1. The finite element analysis revealed the fatigue life and factor of safety contours as shown in Fig. 6 and the results are tabulated in Table 2. From Fig. 6 and from Table 2, it is observed that, the spring has finite life of 1.89 104 cycles. While examining the failed specimen as shown in Fig. 1, it has been observed that the cross section of failure resembles to that of fatigue failure is shown in Fig. 7 with the fatigue life. The fatigue life varies from 10 cycles to 105 cycles for the zone nearer to the inner side of the coil which indicates finite life. The progressively increases for the cross section slightly away from the inside diameter but still this region is having the finite life. While the rest of the cross section has the life more than 106 cycles and which corresponds to infinite life" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000257_indcon.2006.302820-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000257_indcon.2006.302820-Figure1-1.png", + "caption": "Fig. 1. Geometrically based single bounce one ring model.", + "texts": [], + "surrounding_texts": [ + "Geometrically based single bounce model can be contemplated as a stochastic version of the ray tracing model. The separation between the BS and the MS is assumed to be large enough so that incident waves on the MS from the BS can be considered to be plane waves and ray approximation holds true [4]. It is assumed that a particular ray reaching the MS has been scattered by a single scatterer only. The rays those are scattered by multiple scatterers are neglected as they have very little energy and are comparable with the noise floor level. It has been shown in [5] that the diffused component of the signal between jth antenna of the transmit array and the ith antenna of the receive array for a ULA can be written as T ( ) ( ) ( ) ( ) ( ) max 1 1lim exp cos sin sin exp cos 1 N bs ms msbs ms ij bs bs msn nN n d d h j j N \u03c0 \u03b1 \u03b1 \u03c0 \u03b1 \u03bb \u03bb\u03c6 \u03c6 \u03c6 \u2192\u221e = = + \u2212 \u2211 where, max bs\u03c6 is the angle spread at the BS, ms n\u03c6 is the AOA at the MS from the thn scatterer, N is the number of scatterers surrounding the MS, msd is the antenna element separation at the MS, bsd is the antenna separation at the BS, \u03bb is wavelength of the signal, bs\u03b1 and ms\u03b1 are the angles between the x-axis and the orientation of the BS\u2019s and MS\u2019s antenna arrays respectively. It has been assumed that the base station is elevated and devoid of scatterers but the mobile station is uniformly surrounded with scatterers. The above equation is suitably modified for the other array configurations." + ] + }, + { + "image_filename": "designv6_24_0002422_jra.1986.1087052-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002422_jra.1986.1087052-Figure5-1.png", + "caption": "Fig. 5. (a) Disorientation in a comer. (b) Ultrasonic rangefinder scan.", + "texts": [ + " If the sensor is oriented relative to a smooth, untextured surface at an angle greater than half the angle of acceptance, the sound has a tendency to bounce off of the surface at an oblique angle, rather than returning to the transducer. Eventually, it may be reflected back by another surface along the same path. The result of this is an indicated distance to a particular surface that is much greater than the true distance. Since the wheelchair is designed to operate within hallways, the problems associated with reflections make the raw sensor data very unreliable. This is analogous to a person trying to walk through a maze of mirrors, as can be seen in Fig. 5 and will be discussed in detail later. The ultrasonic rangefinder is used primarily to orient the wheelchair with respect to the walls of the hallways. This is necessary whenever the vehicle becomes lost or trapped. The second use of the rangefinder is for the detection of conditions that result in a change in depth. This may be the detection of a moving obstacle in the halls, or the opening of an elevator door. This sensor is used also in conjunction with the vision system to provide gross information about distances", + " To accomodate inaccuracies in sensing and motion, the process is repeated until the current orientation is within an acceptable range of the path parallel to the wall. At this point, the process terminates and control is returned to the procedure, which visually guides the vehicle down the hall. The situation may arise where the vehicle is not lost in a clear hallway, but is actually trapped in a corner. This occurs in the test building when the wheelchair becomes stuck in one of the many alcoves in the hallways (Fig. 5) . In this case, the MADARASZ et al.: THE DESIGN OF AN AUTONOMOUS VEHICLE FOR THE DISABLED 123 ultrasonic scan does not find a suitable wall-hallway combination. When this happens, the vehicle attempts to maneuver to a better location and tries again. The wheelchair is rotated to the direction of the farthest distance and moved to the minimum of one half of the measured distance or three feet. Prior to rotation, the minimum distance is checked to see if it is greater than the turning radius of the wheelchair", + " In only a few instances were the hallways free of people and obstacles. The obstacles usually consisted of inanimate objects, stationary observers, moving pedestrians that were no longer interested in this activity, and those that actively tried to fool the system (ie: waving hands in front of the sonar). In clear hallway situations, such as shown in Fig. 4, the vehicle is usually able to orient itself in two iterations (two scans and one movement), or a total of approximately 2.5 min. The alcove situation (Fig. 5) is usually accomplished in three to five iterations (3-6 min), depending on the type of corner and its actual initial placement. Each situation may require more time and trials if there is heavy traffic in the hallways, particularly if there are a lot of spectators close to the vehicle. At present there is a limit of 15 on the number of iterations that are allowed, and in some cases, this limit is reached. When this happens, the machine stops moving and starts beeping for someone to come and help it" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000974_2002-01-1347-Figure20-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000974_2002-01-1347-Figure20-1.png", + "caption": "Fig. 20. The customer has selected the \u201cSpecials\u201d tool, clicked onto the Rod, C dimension and changed it to the required 2.25\u201d length. The drawing associatively regenerates and re-dimensions.", + "texts": [], + "surrounding_texts": [ + "This paper is based on the continued evolution of software for Product Selection, Configuration and CAD Drawing Generation. The third of a series of presentations that have followed Fluid Power product software from stand-alone PC based programs to rules based, on-line Internet and World Wide Web enabled systems. There have been several approaches taken by Fluid Power product manufacturers to bring product configuration and/or CAD drawings delivered over the Internet. This paper will review these systems and technologies. The ability to provide dynamic, real time information 24/7 for configured to order and engineeredto-order products, directly from the manufacturers and distributors Web site, to customers is now the focus for fluid power e-business solutions." + ] + }, + { + "image_filename": "designv6_24_0002862_detc2007-34856-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002862_detc2007-34856-Figure1-1.png", + "caption": "Figure 1: Snowmobile CVT", + "texts": [ + " Near the beginning of the twentieth century, cars like the Sturtevant, Cartercar, and Lambert featured friction dependent CVT\u2019s [1]. These friction drive CVT\u2019s were common in automotive use until engines capable of producing higher torques became common and necessitated the move to geared, fixed-ratio transmissions capable of high torque transfer and having better wear characteristics than friction dependent CVT\u2019s. Although friction dependent CVT\u2019s have not disappeared from use, until recently they were only found commercially in recreational vehicles like snowmobiles (see Fig. 1) and ATV\u2019s. Figure 1 shows a common CVT used in snowmobile and ATV applications. It consists of two pulleys, referred to as the driver or primary clutch, and the driven or secondary clutch, and a composite v-belt. Each pulley consists of two sheaves, one mobile and one stationary, which move together or apart according to a supplied axial force. The common system consists of a velocity sensing primary clutch and a torque sensing secondary clutch. While the CVT is designed such that the belt should not slip, some slipping will occur, especially at higher torques" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002025_978-3-319-44087-3_34-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002025_978-3-319-44087-3_34-Figure1-1.png", + "caption": "Fig. 1 Elementary sectors (rotor 1 and 2) and multi-stage assembly", + "texts": [ + " Since stage methods will not predict dynamic behavior of the assembled compressor with adequate accuracy. Stage-to-stage coupling effects may significantly impact eigenfrequencies and modes. A priori, disk-dominated modes are highly affected whereas blade-dominated modes hardly contribute to interstage coupling [4]. This paper summarizes different methods [2, 5, 6] to calculate Multi-stage modes at moderate costs. Aiming at a drastic reduction of the DOF-number, different reduced order models are applied. These techniques are demonstrated by means of academic sample rotors, shown in Fig. 1. Since every ROM implies errors compared to a full FE-model, results from a full analysis are used as reference. First, the dynamic of the assembly is characterized and interesting modes, which demonstrate the need of aMulti-Stage analysis, are identified. Since the main focus of this paper is on disk-dominated Multi-Stage modes and blade individual mistuning particularly modify blade-dominated modes, mistuning is neglected here. However, there are ways to consider small deviations (e.g. material property), if the interaction between blade induced mistuning and stage-to-stage coupling were discussed. Figure 1 depicts a two staged academic compressor, which is evaluated in this paper. The geometry is quite simple; however basic characteristics (e.g. blade-dominated, disk-dominated modes and interactions) appear. Although Rotor 1 consists of 9 blades and Rotor 2 is composed of 13 blades, the assembled mesh is compatible. Thus, FE-meshes of both rotors are matching along the interstage boundary. If the meshes are incompatible, multipoint constraints are used to ensure interstage coupling. The dynamics of the assembled structure is determinate by a modal analysis" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003437_iecon.2016.7793967-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003437_iecon.2016.7793967-Figure6-1.png", + "caption": "Fig. 6 Experimental setup", + "texts": [ + " Scheme 2 Scheme 1 is simple, but fails to control the balance since the gimbal tends to lean toward one direction. In Scheme 1, there is no control on the angle of the flywheel. Therefore, in Scheme 2, control of the flywheel angle is added as shown in Fig. 5. The offset compensator for the vibration is designed as )sin( fc (5) where f is the flywheel angle and is a constant. The tilting torque is given as )()( cfbdpfbdd kk (6) The experimental setup for the balancing control of a single-wheel robot with the vertical configuration is shown in Fig. 6. Parameters for the experimental study are listed in Table 1. Sampling time is 0.01 seconds. The gimbal system is located vertically. An AHRS sensor is located on the top of the gimbal system and a weight is attached for the counter balancing. Two schemes are tested for balancing performance. Two tests are performed in different conditions. The corresponding results are plotted in Fig. 7. Fig. 7 (a) shows the balancing control plots of roll and pitch angles. Initially, both angles are maintained well" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000298_1.1807415-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000298_1.1807415-Figure2-1.png", + "caption": "Fig. 2 Pressure profile of hydrodynamic bearing", + "texts": [ + "97 in.! with the width B525 mm ~0.984 in.! and radial clearance h050.05 mm ~1.97 mil!. The included angle of inlet pocket is a530\u00b0 while its axial land width is b53.75 mm ~0.148 in.!. The supply pressure of lubricating oil is ps5103 000 Pa ~15 lbf/in.2! and its dynamic viscosity is m51.25 31022 N s/m2 (1.8131026 lbf s/in.2). The spinning speed of journal is V51000 rad/s ~'9550 rpm!. The eccentricity ratio of journal is e50.5. Figure 1 is the geometry of the bearing model with 148 K mesh density. Figure 2 shows the pressure distribution obtained from the simulation by CFX-TASCflow on this hydrodynamic journal bearing with the same mesh density. There are many factors that affect the correctness and accuracy of simulation by CFX-TASCflow. Two of these require special attention. The first is the mesh density of the geometry. Several mesh densities, including 52, 148, and 308 K, have been calculated and compared to the results by standard lubrication codes. The comparison of static calculation with different mesh density and the results by standard codes are shown in Table 1" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001888_lmwc.2008.2003454-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001888_lmwc.2008.2003454-Figure6-1.png", + "caption": "Fig. 6. Photograph of the BSF.", + "texts": [ + " For a given , stopband rejection increases with the decreasing . A prototype BSF having as 1.5 GHz and 20 dB FBW as 66.67% is fabricated on a 1.58 mm thick FR4 substrate with dielectric constant and . For these specifications, the impedance values are taken as and . Corresponding width and separation of the coupled lines are approximated to 1.0 and 0.3 mm, respectively. The full wave simulator IE3D has been used to obtain the physical dimensions of the BSF. A photograph of the fabricated filter and its dimensions are shown in Fig. 6. On the present substrate, the approximation of equal electrical lengths in the even- and odd-mode excitations holds good for . However, at lower frequencies, the difference between the electrical lengths is significant. As a result, the number of zeros decreases to two. For example, the full wave simulated -parameters of the present design for three cases, (normal case), , and are shown in Fig. 7. This problem of unequal lengths at lower frequencies can be avoided by etching a ground plane slot [9] or taking periodically non-uniform coupled-lines [10] or using a dielectric overlay [11]. Here, as shown in Fig. 6, the equal electric lengths are achieved by a single rectangular groove of dimension 1.9 0.5 sq. mm at the inside edge of each line. The complete filter structure occupies a compact size of 60.8 2.3 sq. mm. The width of the 50 microstrip line is 3.0 mm. Therefore, the filter is narrower than a 50 line. The computed, full wave simulated and the measured responses of the filter are shown in Fig. 8. An Agilent 8510C vector network analyzer has been used for the measurements. Measured maximum group delay variation within the 3 dB upper and lower pass bands is 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001243_icmmt.2010.5524818-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001243_icmmt.2010.5524818-Figure1-1.png", + "caption": "Figure 1. Schematic view of the proposed antenna", + "texts": [ + " It can also be used to construct a large orthogonal-linear-polarized antenna Array. Two 4\u00d74 array are designed with above elements to implement circular-polarized and orthogonal-linear-polarized radiation. Simulated and measured results are provided to demonstrate the capability of the antenna. II. DESIGN PROCEDURE We choose a rectangular microstrip patch as a radiation element because of its simple structure, ease to be designed and good radiation characteristics of the orthogonal polarization. Figure 1 shows the structure of the antenna. The element consists of three layers, which from top to bottom are as follows: the substrate layer of Rogers4003 with radiation patch and the coplanar-fed microstrip line on its\u2019 upper surface, the ground layer with a rectangular coupling slot and another substrate of Roger4003 with a aperture-coupled feeding line on lower surface . The center frequency of the antenna is 8.64GHz.Through parametric study on the antenna dimensions, we obtain the optimized structure parameters as follows: the dimension of the patch is 8.9mm\u00d77.2mm, the slot in the ground plane is 1mm\u00d75mm.The side view and vertical view with the optimized dimensions of the antenna element are shown in Figure 1.The impedance of the feeding line is 100\u03a9. The simulated results of S parameters of the element are shown in Figure 2(a). It is shown that the bandwidths of return loss less than -10dB at port 1 and port 2 are 1.4% and 2.2%, respectively. The isolation between port 1 and port 2 is 1031978-1-4244-5708-3/10/$26.00 \u00a92010 IEEE ICMMT 2010 Proceedings more than 50dB, which is quite desirable for communication application. When the antenna is feed through two excitations with phase difference of \u00b1\u03c0/2, circular polarization performance can be achieved" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003263_1.4031902-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003263_1.4031902-Figure2-1.png", + "caption": "Fig. 2 Cross-sectional views of seven different sheaves: (a) double web, (b) double web with holes, (c) straight web, (d) straight web with holes, (e) thin web with stiffeners, (f) thin web with stiffeners and holes, and (g) web with decreasing thickness", + "texts": [ + " The S\u2013N curve C in air is recommended to use cast design geometries in order to allow for weld repairs after possible casting defects and possible fatigue cracks after some service life [1,5,9]. Seven different design geometries of wire sheaves were considered for this study. The considered design geometries are distinguished as double web, double web with holes, straight web, straight web with holes, thin web with stiffeners, thin web with stiffeners and holes, and web with decreasing thickness as shown in Fig. 2. The geometry of the wire grove (i.e., shown in Fig. 1) is same for all the design geometries. The wire sheave curved in a plane can be considered as curved thin-walled beams. Therefore, it is comfortable to use the explicit formulas proposed in literature [11,12] for evaluating stresses and critical buckling loads. However, out of the seven proposed sheave geometries, five does not have constant cross section throughout their curvature as they consist of discontinuities, such as holes and stiffeners as shown in Fig. 2. Since this is one of the major assumptions behind the above-proposed formulas, this may be a hindrance for applying these explicit formulas for majority of the proposed sheave geometries. Also, these sheave geometries cannot easily be considered as shallow curved beams. Since this is also one of the major assumptions behind the above-proposed formulas, finally this made a doubtful situation to rely on explicit formulas\u2019 given results. Therefore, FE method was utilized for evaluating stresses and critical buckling loads" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000761_j.jmatprotec.2016.09.004-Figure13-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000761_j.jmatprotec.2016.09.004-Figure13-1.png", + "caption": "Fig. 13. Metal flow of valve forging process.", + "texts": [ + " The rolling torque increased sharply at the knifing zone throughout the CWR rolling process and was maximized at the end of the same zone at approximately 0.66 kNm. The rolling torque in the stretching period was relatively stable and remained unchanged at an average of roughly 0.7 kNm. 4.2. Forging process When Fig. 1 is taken as an example, the height-to-diameter forging ratio is 2, and the friction coefficient is 0.3. Furthermore, the CWR preform rod, which is part of the compensation amount, measures 3 mm. Fig. 13 depicts the metal flow during the valve forging process. This process involves upsetting and extrusion deformation. The factors of die pressure, the flow position at the preform metal end of the valve along the axial direction, and the radial outward flow of metal parts are related to the flow rate of the maximum punch fillet die during the initial stage of deformation in forging. Contact with the metal is mainly made through axial flow, and the flow rate is low. Therefore, movement is rigid. H", + " Thus, we attempted to reduce friction in the die forging process by selecting an appropriately sized blank. When the surfaces of the blank and the die increased, resistance to axial radial flow increased. When the amount of metal increased, the valve head became enlarged. Metal mainly flowed to an unfilled corner of the valve head, such as the bottom of the main flow and of the neck of a metal disk when t = 0.077 s. Metal flow rate increased to 1200 mm/s because of the increase in axial and radial resistance and because of the local extrusion deformation in the metal (Fig. 13). In addition, the metal to axial flow filled the deformed workpiece. The axial flow of the metal rod length increased; therefore, the calculated amount of stem volume compensation related to the billet volume of valve forging must increase as well. Fig. 14 shows the distribution of temperature in the process of valve blank forging. The temperature was 1050 \u25e6C. In addition, the minimum temperature in the billet formation process was detected in the part that made contact with the mold. The maximum temperature was detected in the deformed area because the deformation time was short and because of the heat transfer from the billet, mold, and the environment" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000083_ir-03-2018-0042-Figure36-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000083_ir-03-2018-0042-Figure36-1.png", + "caption": "Figure 36 Degree of freedom of the prototype", + "texts": [ + " Where, RTP is the right thigh pitch, RK is the right knee, RAP is the right ankle pitch, LTP is the left thigh pitch, LK is the left knee and LAP is the left ankle pitch. The prototype is designed based on the main model which has the same degree of freedom of the humanoid robot leg. The wearable vehicle prototype has five degree of freedomwhich are thigh pitch, thigh roll, knee pitch, ankle pitch and ankle roll as shown in Figure 35. The motor specifications are given in Table XII. In addition, it has a revolute joint in torso and two revolute joints in a rear wheels mechanism. The prototype is built by using SolidWorks software as shown in Figure 36. The center of mass (COM) of the proposed system is obtained numerically using SolidWorks software. As noted, to achieve the dynamic stability of the system, the torso has to be slightly Figure 34 Gait generation by KTX-PC humanoid robot at 1.1 s Figure 35 CAD model of the wearable vehicle prototype Support Hip pitch, Hip roll and Ankle pitch Knee Voltage 12 V 12 V Speed (No load) 26 rpm 44 rpm Current (No load) 0.21 A 0.21 A Torque (Stall) 42 kgf.cm 25 kgf.cm Gear ratio 455:1 270:1 Gear material Metal Metal Encoder: cycles per revolution ( motor shaft) 3 12 Encoder: cycles per revolution ( output shaft) 1,365 810 Encoder: Countable events per revolution 5,460 3,240 Encoder type Relative Quadrature Relative Quadrature Encoder sensor type Magnetic (Hall effect) Magnetic (Hall effect) All-terrains wearable vehicle B" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000888_ppc.2013.6627666-Figure13-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000888_ppc.2013.6627666-Figure13-1.png", + "caption": "Fig. 13. Comparison between 2D a) and 3D b) FEA electrostatic field plots for the case V1 = 0 and V2 = 150kV . For 3D results, coils on the right occulted to show the field magnitude on vertical and horizontal planes.", + "texts": [ + " The rise time depends on both capacitance and inductance values and in the the absence of overshoot imposed by the specifications constraints, the errors on the inductance and capacitance values partially compensate each other in the computation of the rise time. Both 3D and 2D models are then in accordance when no pulse overshoot is allowed. The induction spatial distribution plot for short circuit operation (Fig. 12) and the electric field spatial distribution plot for the test V1 = 0 and V2 = 150kV (Fig. 13) with 2D FEA and 3D FEA can be compared. On the plane where 2D FEA is performed, one can notice that the results obtained by 2D and 3D FEA are in accordance. The differences between 2D and 3D FEA mainly occur in the windings corners where one can see that the magnetic field is more homogeneous than the electrostatic field. Consequently the error on the capacitances is bigger than the error on the leakage inductance. There is no significative difference between 2D and 3D approach for the magnetizing inductance estimation because only a negligible amount ot magnetizing energy is stored outside the core under no-load operation " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003985_s10846-013-9874-y-Figure17-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003985_s10846-013-9874-y-Figure17-1.png", + "caption": "Fig. 17 Opening and closing position of micro gripper from top view", + "texts": [ + " When fingers are in operation, micro gripper behavior is separately traced on a graph paper. It is found that when micro gripper grips the object/peg, the closing position of jaw in micro gripper is 0.5 mm because 1 mm diameter of pin is taken during experimentation. When micro gripper is in without operation any condition, the opening position of jaw is 5 mm because IPMC has some initial curvature due to inherent flexible behavior of material. Therefore, the maximum jaw opening and closing positions for micro gripper are achieved up to 5 mm and 0.5 mm respectively as shown in Fig. 17. For obtaining the force behavior of each IPMC finger, experiments are conducted for five times. The experimental data of force with voltage are collected as given in Table 2. The average force data are plotted in Fig. 18. It shows the almost linear relationship. The maximum value of force attains upto 12 mN. Using these force data during peg-in-hole assembly operation, the force of first IPMC finger (F1) and force of second IPMC finger (F2) are applied through voltages. During holding the peg, frictional force components of these IPMCs are dominated along with the weight of peg (W) as shown in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000160_icma.2009.5246163-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000160_icma.2009.5246163-Figure4-1.png", + "caption": "Fig. 4 Micro-fcding Device", + "texts": [], + "surrounding_texts": [ + "automation manufacturing are apparently laid out. However, the disadvantages of this system are also shown. To be exact, too much man-aided operation slashes the functional perfection of the whole system, which, in particular, can be seen in cutting arrangement and costume removing: (1) The automation equipment operates through several cloth spreaders and with a single cloth cutter working together in a coordinated sequence. Any problem arising during this process will result in long waiting or delays which will definitely decrease efficiency. In most cases, the manual arrangement of this process is time-consuming and fails in optimization. (2) After the cutting machine finishes cutting, the layers of cut-parts are to undergo manual handling and then to be hung onto the UPS frame in the right order for sewing. This manual removal is obviously time-consuming, efficiency-slashing and liable to errors.\nWith the tendency of smaller-volume garment manufacturing and shorter-cycle production in garment industry, the manual costume removal is increasingly becoming a bottleneck in a quick response mechanism in clothing industry. Only when patch removing process has been made fully automatic can the overall productivity be improved in garment manufacturing sector. In the academics\u00bb, some researchers have concluded that the automatic costume removing technology will shorten the cutting process by, at least, 30% of its present time while the immediate cost will remain exactly at the same level. Another advantage is to make sure that each finished product is made of the cut-parts form the same cloth .It goes without saying that this technology will make a manufacturer capable of taking an edge in the cost competition as it is more economical due to reduction in demand for workshop space.\nTo make a quick response in garment manufacturing a reality, De Montfort University and four other universities in UK have worked hand in hand with 42 partners from the industry and collaborated in CIMTEX - a project for computerized manufacturing system in clothing industry.with a view to undertaking garment CAD/CAM research and development and to increasing their competitiveness in the industry. The co-established team is now committed to the development of automatic patch snatching system in order to produce an integrated, commercialized CNC cutting system.\nThe automatic snatching patch technology developed by the Garment Automatically Manufacturing Research Centre at SUES includes three parts: a single cut-part removal mechanism, an electromechanical integration design and a robot operating program. The project covers a long list of key research items, including an interface of CAD and CAM, data transmission, intellectualized path recognizing and optimized choosing, electromechanical integration design, micro feeding system design, automation control, sensor and feedback. This paper mainly presents the electromechanical integration design of in-piece tackling system based on the concept of generalized mechanism. The main technologies covered are design of the dividing mechanism, design of micro-feeding mechanism equipped with micro-level pressure\nsensor, and the realizing of computer path for patch snatching.\nIII. Design of the Dividing Equipment Unit\nThe automatically snatching patch technology serves the main function of removing the cut-parts made automatically from the cutting table to the patch hanging system. The order in which the cut pieces are removed is programmed according to the sewing process, thus forming a fully automatic garment production line. The automatic snatching patch system mainly consists of three units, namely, dividing mechanism, tackling mechanism and patch removing mechanism.\nThe research project aims to achieve the following goals: (1) undertaking repeated experiments and testing on the patch tackling motion, and deciding on the optimized mechanical movement leading to the best tackling option; (2) concluding the overall design of the automatic patch snatching system and its prototype made on the basis of the design; (3) developing a micro-feeding machine which feeds according to the different pressure levels required by the natures of different fabrics; (4) developing the software aiming to realize automation of the whole system. The key technology is accurate tackling of a single layer from the multi-layer patch and determination of the different pressure levels required by different fabrics so as to achieve micro-feeding.\nThe researchers firstly undertook a series of experiments on mechanism of single-piece tackling from multi-layers, on an optimized tackling speed, on a micro-feeding program, and on precise tackling pressure required by patch pieces of different fabrics. The data and technologies acquired from these experiments form a reliable foundation for the design of an automatic patch snatching system.\nBased on the results of the pressure experiments, the dividing equipment was devised to mainly divide and snatch cut pieces. The mechanism firstly separates one cut piece of cloth from the block of pieces, and then snatches this single piece and removes it to the designated position and height. The robot arm for patch removing will subsequently take the piece and place it onto a hanger of the hanging system.\nThe key point in designing this mechanism lies in the optimized mechanical movement by a drive device selected from the market for the patch dividing mechanism. By observing the concepts of generalized mechanism and design principles, the researchers have devised the generalized mechanism of three -dimensional freedom which operates in a chain. The servomotor functions as a drive system, driving simultaneously the synchronous belt, gear and rack and exactitude screw rod. Fig.3 presents the structure in a sketch, in which ABC is the generalized PPP mechanism, which is an active open movement chain, and XYZ is an axis system and E, pneumatic chunks and F, a cutting machine.\nThe dividing equipment works in the following procedure: pass on to PC the axis information D (x, y) of the cut-piece on the cutting table acquired through the CAD software and the PC co-ordinates with the servomotor by driving the axis X, Y, Z with a powering circuit. When the", + "10\n0. 590. 1\n26.92\n0. 3\n20.19\n0.2\n13.16\nRange ( rnm ) O. I\nPressure (Mpa) 6. 73\nThe maximum pressure inside the cylinder is 40Mpa. The lever is made of aluminium. The maximum range for movement allowed by the micro-feeding mechanism is 0.59 mm, pinpointing the positions of Ll, the dia. piston of d1 and the dia. of lever d2.\nThickness of a cut piece depends on its fabric (0.15 mm for the cloth of spun count 45 and the cloth of spun count 100 is only O. 01 mm in thickness.)\nThe table I shows the different pressure levels for micro feeding mechanism. These levels enable the mechanism to transmit instruction for the movement range required by cut pieces of different fabrics.\nV . Computer Software for Patch Snatching\nThe key technology in the computer control of the path for patch snatching is to enable the operating unit to identify and process the data of CAD. The dividing equipment is controlled by an offline program of the system. This control program will work on the information from the CAD designs and the description of the sewing process, automatically developing a subroutine which enables the dividing machine to snatch the patch precisely and convey them to the correct positions on the hanging system.\nThe driving instructions for the exact path are given from the graphic design pattern completed by CAD. This pattern generates a data file that includes such information as the axis positions of the patch in the marker. To facilitate the multi layer cutting by a NC cutting machine, the variously-sized patch pieces to be cut off the same fabric for the garments of the same design are placed in a marker on the principle that the margins of the edge line and waste of cloth is minimized, and accurate multi-layer cutting is guaranteed by the NC cutting machine. Fig.5 shows an example of the marker. The graphic design pattern covers all the data, some of which, including small curves and turning positions, the type of the cutter and other sorts of information, are rarely required for the computer program determining the patch snatching path; therefore, it is necessary to collect the relevant data to drive the patch dividing machine by scanning and screening out irrelevant data from all the data files.\nTable I Ranges and Pressures of the micro-feeding Mechanism\n/ f\nL--- ! servomotor reaches Point D (x,y) and touches the top-layer cut-piece, the pressure sensor will transmit a signal which stops the movement of the servomotor in Axis Z. The micro feeding device will continue the action and moves along the Z direction. As long as the pressure sensor receives enough pressure, the micro-feeding mechanism stops working and the pneumatic chunks now tackle the patch piece.\nIV. Design of the Micro-feeding Mechanism\nBased on the same concepts and principles , the researchers have completed the design of a micro-feeding machine that operates on the feedback from micro-pressures received by the pressure sensor. The chief function of the micro-feeding machine is to realize micro-feeding along Z direction by identifying the different pressure levels required by cut pieces of different fabrics.\nTo ensure a simple mechanical device demanding less space, micro-feeding is done through a highly elastic component working under external force. FigA shows a general mechanism which performs the step-by-step feeding process through the flexible lever. The feeding device incorporates elastic components with air control technology, making it a physical kinematics chain.\nThe working principle of the device showed in FigA is: when Pneumatic Chunk 2 loosens its hold' Pneumatic Chunk 3 tightens its hold 0 The lever will stretch itself under Force F exerted by the cylinder, causing Chunk 2 to get tight while Chunk 3 gets loose. The pull force as well as strain within the L1 makes the movable parts work. The displacement can be calculated as follows:\na = (F I EA)LI (1) F: Axis direction of stretchable component A: lever section area E: Elastic modulus of the lever material Ll : Range between two chunks\nFig. 5 Graphic design ofa marker", + "A controller is also developed in the programming language to provide a path planning. This controller is designed to be capable of analysing the data in the graphic and then generating the instruction which drives the dividing device to the designated position along the prescribed path.\nThe function of the path planning controller is to make arrangement of the movements according to their priorities and to work out the best path plan. The planning should perform and conclude the whole process of piece dividing. The output of the planning is actually a sequence of a series of points that are involved in all the movements of dividing process, showing the three-dimensional order in which the dividing process is done. This program produces the data needed by filtrating the design graphics information and ensures that all the useful points in the process will be included by the program. For example, an unloaded hanger is possibly needed on the conveyor rail to hang the garment onto a hanging hook, while in tum, a hanger loaded with garment will run the return path on the conveyor rail. To describe this process, a logic program graphics and numbers of points related to the single cut-piece are presented.\nAnother key technology in the computer control of the patch snatching path is to ensure the data transmission between CAD and CAM through an interface. Most CAD suppliers do not provide such service in data transmission simply because they are trying to keep the key technology in data transmission confidential. For instance, Lectra has developed DIAMINO, software for producing markers, which generates files in PLX or PLA; MODARIS, software for pattern making, generates files mainly in MDL, IBA and VET. When data about marker is transmitted to the cutting table, the control software is VECTORPILOT (Now the Version V2R2). If DIAMINO for marker making as well as cutting is expected to work with a self-developed system for cut-piece dividing, snatching and removing, an interface for data transmission between CAD and CAM must be developed.\nThe data transmission interface is actually a program or approach that will realize the data exchange between the two systems. The key process in data transmission interface is the information' being read and then written from one system document to another. This transference can be carried out by such data transmission tools provided by the Windows as Clip&Paste Plate program and OLE technology. The application programs in Windows have such functions as copy and paste, which are actually one of the static transmission instruments provided and supported by Windows. Applicable in transmission to data of different formats, many of the application programs make possible the exchanges between data in different forms, ensuring much of the processing of mixed files in CAD system.\nBesides the several Windows Programs, some neutral files, say, DXF, can also be used to realize data transmission. DXF file is a type of readable text in ASC II with group as its basic component. All the variables, tables and descriptions are all named by group. Then there are several group showing\ndetails. The variables, tables and descriptions then form section. Although the present CAD software developed by different companies usually store graphic pattern data in very different ways, yet neutral files can transfer all of these data, converting the data from graphic designs or markers into DXF files as the output. The CAD software is also capable of receiving DXF files as the input. In this way, different design data exchange can be made between different CAD systems through those neutral files serving as a medium.\nVI. Conclusion\n(1)The concept of generalized mechanism was initially put forward by Professsor Li Xuerong'\", Professor Zhou Huijun further defined it as: \"An Integration Mechanism consists of a driving unit for movement, both controlled and uncontrolled, and a movement chain made of both flexible and non-flexible parts\" [8] \u2022 Introduced into this mechanism are varieties of components--flexible parts and elastic partstogether with some technologies in electrical, electronic, optical, acoustical and aerodynamic engineering. Also involved are software development for drive and control. Generalized mechanism studies both active and passive mechanism. The concept of generalized mechanism further expands the kinematics to incorporate physical kinematics chain, chemical kinematics chain and biological kinematics. These fresh design concepts have in tum broadened the study scope and enriched the elements of mechanism research. Researchers in this project have successfully applied the concept of generalized mechanism to the system and designed the 3-dimensional patch dividing mechanism (PPP for short) which is a typical active open kinematics chain. The micro feeding mechanism is a typical physical kinematics chain which is represented by the elastic components and the air control techniques. The two samples above have proved the theory in generalized mechanism to be very instructive in designing new machinery.\n(2)The automatic electromechanical patch snatching system, designed on the principle above, is able to separate and take the cut-pieces prepared at the automatic cutting table in the sequence of sewing process. These cut-pieces are then conveyed one by one to the patch hanging mechanism. This automatic garment manufacturing adapts well to the small and medium-sized production and will contribute a lot to the garment manufacturing productivity.\n(3) The CAM automatic patch snatching technology and the relevant equipment designs are a typical electro mechanical integration design process as it incorporates and integrates quite a few fundamental concepts and principles in several disciplines such as generalized functional mechanism, intelligent path recognition and its optimization, robot or manipulator, micro-feeding mechanism, automatic control and precision micro-pressure sensor as well as its feedback.\n(4)This system will greatly enforce the competitive edge of garment manufacturing in the industry. However, there are many problems to be solved to make it perfect. For example, the reliability of the dividing equipment needs improvement;" + ] + }, + { + "image_filename": "designv6_24_0001387_acdt.2018.8592940-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001387_acdt.2018.8592940-Figure2-1.png", + "caption": "Fig. 2. Test stand for propeller static tests", + "texts": [ + " The method showed that this approach was able to predict the propeller forward-speed performance effectively. Here, the aerodynamic properties obtained from the approximate method are transformed for the modeling of the virtual propeller. The CFD computation is employed to investigate the comparing results between actual propeller model and virtual model. The task for this study is divided into three steps. First, the information of sample 13x6 propeller (as shown in Figure 1) and its static test performance is obtained using threedimension scanner and static test stand (as shown in Figure 2), respectively, and the pre-processing on determining the aerodynamic properties of propeller is performed using the technique developed in [1]. Second, CFD computations of the 69 actual propeller and virtual model are carried out. Finally, comparison of results from the two models is performed. The work flow diagram of this study is shown in Figure 3. This paper is organized as follow; section 2 presents the details of virtual propeller disk modeling and CFD environment. Section 3 presents the CFD computation results of the actual and virtual propeller model" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003952_978-3-642-33457-3_7-Figure7.18-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003952_978-3-642-33457-3_7-Figure7.18-1.png", + "caption": "Fig. 7.18 One-eight of the motor 3D FEM and coil conductors of a single phase", + "texts": [ + " From the simulation PM-flux linkage and, accordingly, no-load voltage appear to be noticeably lower (10 percent less) than predicted by the 2D analytical model which is a result of the axial flux leakage. The leakage is primarily enhanced by 1 mm extrusions of the rotor iron which support axial sides of the magnet\u2014their presence partly short-circuits the field of the magnet (see Fig. 7.17). Motor phase inductance is simulated in 3D FEM using non-meshed rectangular coils (it was not possible to create toroidal coils in the used software)\u2014Fig. 7.18. The analytical model (and, also, 2D FEM) greatly underestimates the phase inductance: the analytical prediction is almost an order of magnitude lower than the prediction of the 3D FEM. It would be extremely difficult to directly simulate losses in the conductors in FEM because of their disproportionally small size with respect to dimensions of the motor parts. Eddy-current loss in the conductor strands is, therefore, indirectly simulated. Firstly, the field in the center of the windings is calculated and simulated using 2D FEM" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002642_j.mechmachtheory.2017.02.003-Figure19-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002642_j.mechmachtheory.2017.02.003-Figure19-1.png", + "caption": "Fig. 19. Scheme used to obtain the cam profile starting from \u03b2( \u03b8 P ) function.", + "texts": [ + " { \u03b2 = \u03b2( \u03b8P ) \u2212 \u03b20 \u03b1 = \u03b1( \u03b8P ) \u2212 \u03b10 \u03b2 = \u2212 \u03b1 \u2192 \u03b2( \u03b8P ) = \u03b20 + \u03b2 = \u03b20 \u2212 \u03b1 = \u03b20 \u2212 \u03b1( \u03b8P ) + \u03b10 (12) The initial value \u03b20 is a parameter that should be chosen properly because it affects the shape and the dimension of the cam profile. A first observation can be done on the minimum acceptable value for \u03b2 . In order to have a positive radius of the cam, \u03b2 must be greater than zero for any values of \u03b8P . This affects the choice of \u03b20 that must be greater than \u03b1. Once these preliminary concepts have been fixed, the procedure for the cam design can be described. In Fig. 19 two different configurations for the mechanism are showed in the kinematic inversion in which the locomotion unit and the cam connect with it are fixed. Variables with subscript zero refer to the initial configuration of the climbing sequence. The notation with apostrophe indicates variables values in a different and generic configuration of the mechanism. According to this kinematic inversion, during the locomotion unit rotation, the wheelchair frame PC moves around P with a rotation of \u03b8P = \u03b8 \u2032 P \u2212 \u03b8P0 . Meanwhile the seat RC rotates around P and moreover the relative orientation between PC and RC changes according to the desired angle \u03b2( \u03b8P ), in order to remove the oscillation of the seat. The trajectory of point R around P describes the desired cam profile. The cam profile can be described in polar coordinates (h CAM , \u03b4) with respect to a reference frame fixed on the locomotion unit and centered in P. By applying the cosine and the sine theorems on triangle P C\u2019R\u2019 of Fig. 19 , Eqs. (13) and (14) can be 0 derived. Finally, Eq. (15) can be written through geometric considerations on the scheme of Fig. 19 . These equations define the polar coordinates of the profile, starting from the mechanism and the wheelchair parameters (l PC , l RC , \u03b20 , h C ). h CAM = \u221a l 2 PC + l 2 RC \u2212 2 l PC l RC cos \u03b2 (13) \u03b5 = sin \u22121 ( l RC h CAM sin \u03b2 ) (14) \u03b4 = \u03b5 + \u03b8p (15) In summary, the cam design process can be related to the synthesis of a motion generator cam mechanism as defined in [30] that imposes a straight line translational motion to the output link RC. The procedure described in the previous paragraph defines a cam profile capable of removing the seat oscillation at least for the nominal stairs", + " 20 the mechanism is represented in another kinematic inversion in two different configurations: the initial one (P 0 C 0 R 0 ) and the configuration associated with the maximum value for h P (P 0 C 0 R\u2019), that corresponds to the maximum value of \u03b1 and to the minimum values of \u03b2 and h CAM . The wheelchair frame PC is fixed, the seat RC rotates around C and the cam and the locomotion unit rotate around P. This representation can be associated with the generic representation of a cam mechanism with swinging follower. The mechanism (see Fig. 19 ) has four independent parameters: l PC , l RC , \u03b20 and h C . The other quantities are functions of these parameters, as can be observed in Eqs. (13) \u2013(15) . The parameter h C has been fixed to the best value according to Fig. 16 . The remaining three parameters can be redefined by introducing the dimensionless quantities as in Eq. (16) . R = l RC l PC I = l PC l L (16) The parameters R and \u03b20 change the cam shape while I affects the mechanism dimension respect to the locomotion unit size. A complete analysis of the cam mechanism cannot be done only focusing on the dimension and geometry of the cam profile. Important quantities that must be taken into account are the pressure angle ( \u03b8PRESS ) and the radius of curvature ( \u03c1). According to [31] , Eqs. (17) \u2013(20) can be written. These equations are known from the cam mechanism synthesis theory and can be understood referring to Fig. 19 . In details, Eq. (17) allows the computation of the pressure angle of the cam profile from the geometry of the mechanism. The angle \u03b2 represents the angle between the frame PC and the rocker arm RC, while K is the center of curvature of the pitch curve. Then, further parameters can be introduced: the radius of curvature ( \u03c1), and the angle \u03c6 that represents the angle between the segment KC and the direction defined by the frame PC. The Eq. (19) allows calculating the radius of curvature while Eqs" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003248_ifost.2016.7884321-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003248_ifost.2016.7884321-Figure1-1.png", + "caption": "Fig. 1 Scheme of boiler BKZ-220-100ZhSh 1, 3 \u2013 steps of an airheater; 2, 4 \u2013 steps of the economizer; 5 \u2013 convective step; 6, 7 \u2013 output and furnace screens; 8 \u2013 furnace camera; 9 \u2013 nozzle of secondary air; 10 \u2013 eparator of a mill; 11 \u2013 shot cleaning system", + "texts": [ + " Further it is considered that process proceeds in two directions: oxidation of nitrogen to nitrogen oxide and formation of molecular nitrogen owing to a recombination of atomic nitrogen or as a result of recovery of an oxide of nitrogen. The settlement scheme of generation of fuel oxides is described by system from three equations [4, 5]. III. RESULTS AND DISCUSSION The numerical models of generation of N x described above were applied to operating BKZ-220-100ZhSh which configuration is provided in fig. 1. In fig. 2 distribution of concentration of nitrogen oxides on BKZ-220-100ZhSh copper fire chamber height is provided in case of from excess of air The boiler BKZ-220-100ZhSh has the prismatic furnace camera. Height of the boiler is 24.9 m, width of a fire chamber is 8.64 m, depth of a fire chamber is 7.74 m. The furnace camera is completely screened pipes diameter 60 mm. Poorly inclined symmetric under ( = 15z), formed by pipes of front and back screens, corresponds to system with a liquid slag removal, and the lower part of a fire chamber up to the height of 5" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002258_tap.2017.2730250-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002258_tap.2017.2730250-Figure4-1.png", + "caption": "Fig. 4. Vector of electric field distribution at center frequency in the middle of the substrate between two water layers.", + "texts": [ + " It can been seen that the electric field excited by the vertical disk-loaded probe are mainly confined between the water patch and water ground due to the larger difference of dielectric constants between water and air. The energy propagates from the central to the edge of water patch and then radiate into the free space through the surrounded circular open slot. In contrast, very weak electric field exists inside water so that this kind of water patch antenna does not work in the dielectric resonant mode. The electric field vector in the middle plane of air substrate at the center frequency were simulated and depicted in Fig. 4. For circular patch antenna, in order to achieve a conical beam radiation pattern with a null in the broadside direction, a TM02 mode needs excited based on the cavity mode analysis method [22]-[24]. The electric field distribution in the substrate is same as the electric field distribution in metallic patch operating in TM02 mode, which confirms that this proposed water patch antenna works in a kind of resonant mode of a conventional patch antenna. The measurement on S-parameters, gain, radiation patterns and radiation efficiency was achieved by Agilent E5071C network analyzer and SATIMO complex antenna measurement system" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002920_glocom.2017.8254866-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002920_glocom.2017.8254866-Figure4-1.png", + "caption": "Figure 4: A eight-node three-dimensional radiation pattern in 60GHz", + "texts": [ + " Then for a certain receiver \ud835\udc56, the signal-tointerference-and-noise ratio (SINR) in the LOS is given by \ud835\udc46\ud835\udc3c\ud835\udc41\ud835\udc45\ud835\udc56 = \ud835\udc43\ud835\udc56\ud835\udc54\ud835\udc56\ud835\udc51\ud835\udc56 \u2212\ud835\udefc \ud835\udf0e2 \ud835\udc5b + \u2211 \ud835\udc57 \u2215=\ud835\udc56 \ud835\udc43\ud835\udc57\ud835\udc54\ud835\udc57\ud835\udc56\ud835\udc51\ud835\udc57\ud835\udc56 \u2212\ud835\udefc (1) where \ud835\udf0e2 \ud835\udc5b is the noise power, \ud835\udefc is the path loss factor, \ud835\udc43\ud835\udc56 is the transmit power, \ud835\udc51\ud835\udc56 is the transmission distance to receiver \ud835\udc56, \ud835\udc54\ud835\udc56 is the corresponding transmit antenna gain determined by the angle and beam radiation pattern, while the distance and antenna gain from the interferer \ud835\udc57 to the receiver \ud835\udc56 are \ud835\udc51\ud835\udc57\ud835\udc56 and \ud835\udc54\ud835\udc57\ud835\udc56, respectively. The total interference at the receiver \ud835\udc56 is given by \ud835\udc3c\ud835\udc56 = \u2211 \ud835\udc57 \u2215=\ud835\udc56 \ud835\udc43\ud835\udc57\ud835\udc54\ud835\udc57\ud835\udc56\ud835\udc51\ud835\udc57\ud835\udc56 \u2212\ud835\udefc. Then by considering the 3-dimensional radiation pattern in Fig. 1 and the nodes\u2019 locations in Fig. 4, we calculate the total interference field in a 8-node network in Fig. 5. C. Comparison with Traditional Networks Based on the interference calculation explained above, we carry out a comparison on the distribution of interference between traditional omni-directional communication networks and directional mmWave networks. For the traditional networks, we consider small scale fading (Rayleigh fading and Rician fading in the simulation). We also consider the power gain of directional mmWave antenna as a special type of fading, which is random due to the randomness of the angle between the interferer and the interfered node" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002574_s0266-3538(02)00044-1-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002574_s0266-3538(02)00044-1-Figure5-1.png", + "caption": "Fig. 5. Stress distributions for layer 1.", + "texts": [], + "surrounding_texts": [ + "In order to validate the aforementioned approach, a numerical analysis on a pole specimen with external diameter equal to 60 mm and wall thickness equal to that of the 80 mm pole was performed making use of the FE model already set up (i.e. same mesh and material properties). By applying a monotonically increasing load it was possible to determine the curve load-diametrical deformation which, being in elastic regime, is a straight line of equation: y \u00bc 0:5719p \u00f03\u00de where p indicates load per unit length (N/mm). Initial specific ring stiffness experimental tests were then performed on the 60 mm pole inducing a diametrical deformation of 1.73 mm (2.9% of the diameter) with a mean load recorded value of 587.73 N which corresponds to p=2.937 N/mm. This last value, introduced in Eq. (3), returns y=1.68 mm (see also Fig. 4) showing a good agreement between the numerical results and the experimental ones. The numerical model of the studied Table 5 Elastic constants for (a) layer 1, EF 3D-solid, (b) layers 3 and 5, EF 3D-solid, (c) layers 2 and 4, EF shell a. Young\u2019s moduli (GPa) Ea Eb Ec 6.13 6.13 121 Poisson\u2019s ratio ab nac nbc 0.019 0.019 0.2 Shear moduli (GPa) Gab Gac Gbc 0.7 0.7 0.8 b. Young\u2019s moduli (GPa) Ea Eb Ec 7.13 7.13 140 Poisson\u2019s ratio ab ac bc 0.019 0.019 0.2 Shear moduli (GPa) Gab Gac Gbc 0.7 0.7 0.8 c. Young\u2019s moduli (GPa) Ea Eb Ec 5.1 5.1 98 Poisson\u2019s ratio ab ac bc 0.019 0.019 0.2 Shear moduli (GPa) Gab Gac Gbc 0.49 0.49 0.56 Fig. 4. Displacements along the diametrical direction for the 60 mm diameter prototype. structural elements, based on the calibration of the material parameters by comparison with experimental results obtained for a specific prototype, seems to be an effective tool for predicting the mechanical behaviour of a typology of such elements. The model can in fact be utilised for the design of \u2018\u2018new\u2019\u2019 elements made of the same material but having different geometrical dimensions and even slightly different cross-sectional shapes. Moreover, a design procedure driven by a numerical simulation allows one to define the \u2018\u2018right\u2019\u2019 geometry as well as the \u2018\u2018right\u2019\u2019 layer arrangement when loading conditions, often not reproducible experimentally, have to be considered so avoiding expensive trial-error manufacturing approaches. Obviously an accurate, i.e. experimental-based, FE analysis gives detailed information on the mechanical behaviour of the structural element in terms of stress and strain distributions. Such results are given, for the 80 mm pole specimen, in Figs. 5 and 6 were, for sake of brevity, only the stress and strain distributions in the external layers are reported. Analogous distributions were found in the other layers and the obtained numerical results can be qualitatively summarised as follows. In each layer the maxima normal stresses, either in compression or in traction, are recorded along the directions of the fibres. So that for layers 1, 3, and 5 higher normal stresses are evidenced parallelly to the pole\u2019s axis, while for layers 2 and 4 these values are maxima along the hoop direction. All the layers are subjected to variations from tensile to compressive stresses; precisely, on the loading plane the tensile stresses increase going from the outer versus the inner layer while the compressive stresses decrease. An opposite trend is observed perpendicularly to the loading plane. A similar qualitative behaviour is exhibited by the shear stresses. Finally, the computed strains are consistent with the stress distribution and are practically vanishing in the axis pole\u2019s direction and attain their maximum absolute values in the hoop direction." + ] + }, + { + "image_filename": "designv6_24_0003345_amm.813-814.964-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003345_amm.813-814.964-Figure5-1.png", + "caption": "Figure 5.FEA Test Setup of Brake Pedal.", + "texts": [ + " In the brake pedal, following loads are assumed - Generally human weight is about 700N - 1100N (as per the Knorr Bremse company standard) which forms main consideration for design of brake pedal. In emergency situation, driver stands on the pedal making pedal prone to bend or fail. Thus the maximum load considered is 1250N. Length of human foot is 190 mm, but while applying force driver makes use of its half foot, so only 90mm length is considered. The brake pedal rotates about the pin and the pin is fixed in base plate so it acts as a hinge support about pin, as shown in figure 5. Degree of freedom for both the holes where pin is fixed is considered as zero in X and Y directions and the rotation about z direction are allowed. Result of FE Analysis The prepared deck is solved into ANSYS software. As per the material specification, for checking its failure mode, displacements and stress plots are drawn. The similar load is applied on original brake pedal to compare the results of displacements and stress. Table 2 and Table 3 shows the results of Original and developed model respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001943_0890-6955(95)00104-2-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001943_0890-6955(95)00104-2-Figure4-1.png", + "caption": "Fig. 4. 4-axis vertical machine tool with a rotary angle head attachment.", + "texts": [ + " Equation (13) is still valid, provided d, = 2r, A = 0, and m = kd. 5. MACHINE TOOL SE'Iq'INGS The manufacturing of screws is complicated and precise work. It depends upon the generating method and types of the employed machine tools. To investigate the machine tool settings, the ability function of the machine tool should be determined first. 5.1. Ability function of a 4-axis machine tool equipped with rotary angle head attachment A 4-axis vertical machine tool equipped with rotary angle head attachment is shown in Fig. 4. It consists of 4 rigid links typically connected in series by prismatic or revolute joints. Here, again, in order to obtain its ability function, all the links should be numbered sequentially starting from the workpiece marked as \"0\" and ending at the cutter link marked as \"4\". Once the frames (xyz)~ (i = 1,2,3,4) have been assigned to the machine tool, the various link parameters can be tabulated as in Table 2. Here we always use underlined parameters to distinguish them from those of VPSTM. Note that O~", + " Referring to Fig. 5 again, when the tip of the cutter coincides with the origin of workpiece frame (xYZ)o the position and orientation of tool frame (xyz)t with respect to the workpiece frame is given by: ii0 1 ~ ? - C z \u00b0A-' - ? -S'y (17) 0 0 1 By equating the corresponding elements of equations (16) and (17), one obtains the following expressions of origin of workpiece frame in terms of joint variables of machine tool: _01o=0 (18) _b2o = - b l + dS'y (19) bso = -a_4-_dC7 (20) b4o = - a l (21) Referring to Fig. 4, the desired NC position commands of A, X, Y, and Z axes, expressed with respect to screw frame, are now given by the corresponding differences of equations (A9)-(AI2) and (18)-(21): A = __01-_01o (22) X = -(_bz-b2o) (23) NC Data Generation for 4-axis Machine Tools 349 y = - ( b 3 - b 3 o ) (24) Z = -(b4-b_4o) (25) The values of Y and Z are constants and are referred to as positional constraints. Positional constraints halt the machine axis motion by specifying the position of the axis and take the form: q = constant" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002782_mop.24669-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002782_mop.24669-Figure5-1.png", + "caption": "Figure 5 Human arm phantom model", + "texts": [ + " The radiation behaviors are similar to monopole with a nul in the direction collinear to the feeding line. However, as for the IIFA, an important part of the near field is radiated in the body making again the antenna sensitive to it. The working bandwidth is around 120 MHz. The antenna gain is around 4 dBi. The dimensions are as follows: lp 10.5 mm, Lp 13 mm, L 32 mm, l 24 mm, Wc 2 mm, s 0.7 mm, e 1 mm, d 2 mm. To consider the electrical properties of biological tissues [6], a human arm phantom is modeled (Fig. 5) in HFSS\u2122. The phantom is a planar multilayered structure (Table 1) representing the skin, fat, muscle, and cortical bones of the arm in a simple way. To properly evaluate the near field effects and to comply with SAR characteristics, a -sized phantom has been realized with 120 mm 120 mm lateral dimensions at 2.4 GHz. The HFSS phantom model shows good agreement with human arm measurements as shown in Figure 1. This validates the electrical parameters of biological tissues in Table 1. Modeling-embedded fluid (blood vessels) will further increase the accuracy of the model" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001324_ausms.2016.7593478-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001324_ausms.2016.7593478-Figure1-1.png", + "caption": "Fig. 1. Ray Diagram of Single Mode Propagation with Critical Beamwidth.", + "texts": [ + " SIMULATION SETUP The simulation uses an analytical model, which considers path-loss, reflection attenuation and wave interaction. The transmitting and receiving antennas are assumed to have a radiation pattern with a given beamwidth and they are always directed towards the receiver with the path of maximum gain. In the simulation model, each path undergoes free space path loss that is proportional to the distance squared over which the path propagates. In addition, each reflection that exists is attenuated and phase shifted by 180 degrees. Fig. 1 depicts the 2D model of a typical tunnel with a single directed beam radiating from the transmitting antenna, Tx, where R1 is the path of the main lobe with maximum radiation intensity delivering the maximum power along the tunnel to the receiving antenna Rx. It should be noted that additional reflections \u2013 R\u20192 and R\u20193 - also exist; they may reflect a number of times between the walls before reaching the receiver. The tunnel width, x, and length, y, define the minimum beamwidth required to have a single propagation path. Using the dimension in Fig. 1, the relation is defined as yx R 2 2 3 1 2 cos2 If the beamwidth of the propagating wave is greater than , the channel will exhibit multipath propagation. Close to the transmitter, these multiple beams rapidly change according to spatial location [5] which can be problematic for fixed directional antennas [11]. However, with a single directed beam in the channel, the destructive behavior of multipath reflections in the near field is significantly reduced. III. SIMULATION RESULTS The performed simulations are based on a typical metal mine tunnel [12], such as iron ore, which has an average width of 5 m" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003247_9781119258827.ch7-Figure7.6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003247_9781119258827.ch7-Figure7.6-1.png", + "caption": "Figure 7.6 The Merritt skid steering model.", + "texts": [ + " Large differential longitudinal track forces must be developed by the tracks to overcome these moments. Predicting these forces and the interface slips is fairly complex. Up until the Second World War (WW2) the design of steering systems appears to have been largely by trial and error. Merritt [7.3] produced the first analysis of steering behaviour that could usefully be used for design purposes. It gave fairly accurate predictions of track forces and rates of turn for small radii of turn at low speeds. Figure 7.6 shows the Merritt model for a tracked vehicle executing a turn, where W is the vehicle weight, 2c is the distance between the track centres and 2l is the track ground contact length, generally taken as the distance between the front and rear road wheels. Initial assumptions are uniform load distribution on the tracks and simple Coulomb friction between the tracks and the ground. As noted above, large differential tractive forces are required to overcome the friction moment. This requires the outer track to slide backwards over the ground and the inner track to slide forwards", + "3) M W l a a a = +( ) \u2212 \u22120 25 1 12 0 5 2 1. sinh . (7.4) The moment produced by the equal and opposite track forces Fx is equated to the friction moments M. These two relationships can then be inserted into an Excel spreadsheet together with the dimensions of an example vehicle, and the equations solved by using the Solver routine in Excel. The track speed difference dv required to achieve a radius of turn R for a given forward speed V is readily calculated by considering the geometry of Figure 7.6, and can be shown to be: d o iv V R c a a l= + + 2 ( ) (7.5) This equation can be inserted into the spreadsheet, where the term ( )a a lo i+ represents the effect of track slip. Steeds [7.4] extended the method to include the effect of lateral forces, from cornering forces for example, and lateral load transfer between the tracks. Solution of the Steeds method required a tedious trial and error process; various graphical solutions of the method have also been produced including that by Wormell and Purdy [7", + " The most suitable control regime would probably be established by an experimental programme. For pivot turns the model can be set up in one of two ways. The first is by taking the centre of the vehicle as the turning centre for both the hull and the tracks. The difficulty here is that large slip angles are obtained for the first and last wheels. The variables in the Solver routine are the (equal) track slips. Results, for new pads and with track pretension, show a slewing moment of 142 kNm with a slip of 0.49. The second method is by using the Merritt method [7.3] Figure 7.6 with an instantaneous centre for the hull at the vehicle centre and two instantaneous centres at equal distances outside of each track. Slip angles are calculated from these centres. The variables in the Solver routine are the (equal) track slips and the distance of the track instantaneous centres to the track. Results, for new tracks, show a slewing moment of 140 kNm with a slip of 0.4. The power required by the steering system is usually governed by the required rate of turn \u03c9p while performing a pivot turn" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001153_amr.189-193.1882-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001153_amr.189-193.1882-Figure5-1.png", + "caption": "Figure 5", + "texts": [ + " Click the icon to fix the Qianzhoulian and Chetoucheshen,Houzhoulian and Chetoucheshen. Click the icon \u201ctranslation\u201d to add translational joint between Huakuai1 and Chetoucheshen. Click the icon to add revolute joint between Huakuai1 and Changgan.Click the icon to add contact force between Huakuai2 and Chexiang.Click the icon to add revolute joint between Chexiang and Duangan. Click the icon to add revolute joint between Changgan and Duangan, Changgan and Chexiang, Duangan and Chetoucheshen.The added joints is shown in figure 5. Move Jushengyougang to the proper postion and Click the icon to add the revolute joint between Jushenyougang and Changgan.Click the icon \u201ctranslation\u201d to add the translational joint between Jushengyougan and Jushengyougang.Click the icon \u201c translation motor\u201d to add the motion on this translational joint and rename the motion as \u201cJushengbang\u201d.Set the motion function as \u201c-130* (max(time,5)-max(time,15)+10) \u201d.Similarly move Qingxieyougang to the proper postion and add the revolute joint between it and Changgan" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003869_851385-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003869_851385-Figure4-1.png", + "caption": "Figure 4. Combined Environments Acoustic Facility", + "texts": [], + "surrounding_texts": [ + "851385 5\ncombustion of SOFI decomposition products. Aft-dome combustion is sustained by turbulent mixing which provides sufficient oxygen at the SOFI surface to promote burning. Wind-tunnel tests evaluated char recession under co~bined aerodynamic heating and shear, and determined ablation performance at protuberance heating levels. Minitanks are 3 ft diameter X 5 ft high aluminum tanks used to assess TPS performance under repeated cryogenic fill/drain cycles and pressurization. Minitank tests evaluated substrate-primer-TPS strain compatibility, primer and TPS adhesion, ab1ator and SOFI cracking and susceptibility to cryopumping. Design features such as c10seout/ repair configurations and instrumentation wire bending were also evaluated with these tests. To assure that the applied TPS materials did not delaminate and falloff with catastrophic consequences the flexstrain also known as the cryof1ex test was developed. The tester is a single axis device capable of applying any combination of tensile and flexural loads simultaneously in an environmental test chamber which allows the simu1atation of thermal loading (Figure 2). The key element in the TPS qualification was the design- verification tests conducted on 4 ft X 4 ft combined-environments panels. The biaxial test panels were representative of specific critical areas of the external tank structure and TPS design. In these tests, the TPS experiences the timed phased combined effects of cryogenic SUbstrate, biaxial strain, acoustics, ascent heating and pressures. Figures 3 and 4 show a typical combined enviroments test panel and test apparatus. The qualification methodology was to combine the ablation wind-tunnel tests with specific design-verification tests conducted under combined environmental conditions to achieve confidence in the total TPS system.", + "6 851385\n~~TERIAL/PROCESS VERIFICATION\n\"The verification test program for the ET TPS included the wind tunnel tests discussed above together with several cryogenic, radiant heating, and combined environment tests. These tests were designed to verify the TPS integrity under the various predicted flight induced environments. There was no one test (other than flight) which simulated all of the pertinent flight parameters. Confidence in the TPS system was achieved by the successful results of these tests taken together. \"Minitank tests were used to evaluate TPS cryo-strain compatibility, primer and TPS adhesion, and TPS cracking and susceptibility to cryopumping. The minitanks were 3-ft diameter aluminum tanks with TPS applied and were tested under repeated cryogenic fill, drain and pressurization cycles. These tests did not simulate any ascent pressure, heating or acoustics loads. \"A larger 10-ft cryogenic tank was also tested similar to the minitanks to assess any large scale application issues. The 10-ft tanks, like the mini tanks, were tested with LH under repeated cryogenic cycles. The 10-ft tank also included a radiant heat test to assess TPS recession and propellant quality on a large scale tank. Note that the 10-ft tank was the largest scale application of flight type TPS prior to STS-l. \"Radiant heating tests were conducted to verify the TPS recession characteristics under the aft dome environment where heating was due primarily to the exhaust plume radiation and recirculation, rather than aeroheating as simulated in the windtunnel tests. These tests were conducted in two facilities; one simulated radiant heat and acoustics, and the other\nradiant heat and ascent pressure decqy. \"The key element in the TPS qualification was the \"combined environment\" tests conducted on four TPS panels configured to represent the substrate and TPS in specific critical areas on the ET. These panels were subjected simultaneously to biaxial substrate loads, cryogenic backface temperature, ascent heat load (radiant), and either acoustics or ascent pressure. The panels were tested in a thennal-vacuum chamber, and/or in a thermal-acoustic facility depending on the specific test objective. Both facilities employed a large load cell structure which could be programmed to induce biaxial load profiles in either tension or compression. These were used to simulate various degrees of predicted flight substrate loads to demonstrate the TPS structural margin. The panels were cooled with liquid helium to simulate hydrogen tank substrate temperatures and the flight heat loads were simulated by an infrared lamp bank. These unique tests allowed the TPS to be subjected to nearly all the flight conditions (except aeroheating) and on a scale large enough to verify production methods (Bachtel et al., 1983).\"\nFLIGHT RESULTS\nActual flight data from the first six ETs provided flight verification of the predicted thermal environments. Specially designed instrument islands measured aerodynamic heating and boundry-lqyer conditions at 30 locations on the ET sidewalls and 8 locations on the LHZ aft dome. In addition, the temperature of the tank structure was measured at 59 locations. Flight instrumentation was augmented by cameras mounted in the orbiter which photographed the", + "851385\nET during ascent and separation from the orbiter. Photographs confirmed the locations of higher heating regions. Initially the flight instrument data was used to compare the environments predicted with the actual environments. This data together with the photographs confirmed the adequacy of the TPS materials and design. Currently this same data is being used to further reduce cost by further reducing the coverage of the ablator based on the actual temperatures, not the higher predicted heating.\nSUMI4ARY\nSince the start of the project, the External Tank has progressed from development to production. The thermal protection system has changed with changing requirements and improvements in materials and processes (Table 2), however, it is still true to the original concept. Today's changes represent an upgrading and maturing of an already efficient design. Figure 5 shows the current TPS configuration. As the production rate increases, most changes are directed toward\nproducibility improvements and elimination of sole-source dependencies. Increased use of robotics and net-molding techniques will greatly reduce touch labor and material usage. The qualification of NCFI 22-65 for use on areas other than the aft dome would provide an alternative to CPR-488 and could potentially reduce the amount of ablator used. Other efforts are underway to develop cheaper/lighter ablators and/or cheaper foams with improved thermal capability to replace SLA-56l ablator completely. Based on flight and production experience, the already efficient design will continue to be optimized to enhance producibility, consistency and reliability. Continuing concerns include the streamlining of launch operations by minimizing the number of TPS closeouts and simplification of active thermal- control systems (purges and heaters). With the forthcoming launches from Vandenberg Air Force Base (VAFB), a new set of environmental parameters must be considered. Vandenberg's more severe weather conditions and launch profile present new challenges.\n7" + ] + }, + { + "image_filename": "designv6_24_0001686_bf03397240-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001686_bf03397240-Figure3-1.png", + "caption": "Fig. 3-Nuclear thermionic power system.", + "texts": [ + " Anal yses of interplanetary missions have indicated that a high specific-impulse electric propUlsive system must have a specific weight as low as 10 to 20 lb per kw at levels of 1 Mw and above to be com petitive at short mission times with chemical or nuclear rockets.~,3 Therefore, further discussion of power systems will be restricted to the nuclear turbogenerator and thermionic systems. 638-JOURNAL OF METALS, SEPTEMBER 1963 Nuclear thermionic power A typical thermionic power system is depicted in Fig. 3. The system operates as a heat engine, ex cept that electrons rather than a gas or vapor are the thermodynamic fluid. Electrons boiling off the hot cathode travel across a gap to the cooler anode and return to the cathode via an external load. The waste heat resulting from the thermodynamic cycle plus the heat lost because of radiation from the cathode to the anode are extracted by a coolant (heat-transfer fluid) that flows around the outer surface of the anode. The coolant is passed through a radiator where a portion of its heat is given up to the radiator, which in turn radiates this heat into space" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001572_tmag.1981.1061290-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001572_tmag.1981.1061290-Figure8-1.png", + "caption": "Fig. 8 S t r u c t u r a l Model. The d i s t o r t i o n s a r e f rac t ions of Ro and a r e shown exaggerated.", + "texts": [ + " STRUCTURAL ANALYSIS Two practical problems are now addressed ( 1 ) The effect of deformations (in both the coil and structural supports) on the internal force distribution, and ( 2 ) A realizable structure quivalent o the schematic arrangement shown in Fig. 2 The cylinder and pinned link configuration is difficult to construct due to the congestion where the links converge. The inverted arrangement shown in Fig. 7 is simpler. The coil structure is extended axially in the shape of a triangular blade ( B ) and a massive ring (A) restrains the radial motion. The blade then pushes towards the midplane at R1. A coil shape was geperated for = 0.7 ch = 0.2 and f = - .2. A structural model is shown in Fig. 8 . &e analysis, using $ structural code,6 indicates that distortions reduce the comprewive force at R1 to = - 0.17. All forces are within - 5% of the predscted values. fP . 1945 # 01 \u20ac0038 r Circu lar Co i ls 1 Dee Coi ls \\ n tant Stress Bow Coi ls l Demountab le Bow Coi ls Meon Length 2.72 2.13 1.67 1.28 2.19 1.56 1.09 0.74 1.6 1. 5 0.97 0.74 0.80 0.70 0.60 0.50 Z Max. 38.0 27.0 1 .7 12.3 31.6 20.5 12.8 7.3623.9 16.7 11.4 7.36 12.6 9.67 7.11 4.93 Volume 12.2 9.76 7.77 6.06 10.3 7.65 5.65 4.08 8" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002855_971458-Figure9-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002855_971458-Figure9-1.png", + "caption": "Figure 9 : Deformation pattern of a Y-shaped seat leg.", + "texts": [], + "surrounding_texts": [ + "Figure 7. The weights and heights of these ATD models are tabulated in Table 1.\nFigure 5 : Y-shaped seat leg design.\nAnother design of the seat leg is the Y-shaped seat leg (see Figure 5). The Y-shaped seat leg has two Figure 7 : Three types of ATD's. sections. The upper section is an u-shaped structure\nformed from Al2024-T351 sheet. The edges of upper seat leg section are welded to the seat pan structure. The bottom section of the seat leg is a rigid I-beam. During impact, the U-shaped seat leg is capable of deforming as much as 5 inches from the vertical stroke before the seat pan structure makes contact with the Ibeam structure. Figure 6 shows an illustration of the Yshaped seat leg design generated by MADYMO. The seat leg structure as appeared in CAD drawing are coincided with the seat leg position in MADYMO drawing.\nFigure 6 : Superimposed view of the Y-shaped seat design.\nOCCUPANT MODELS\nThree type of ATD models were used for the analysis including the sh percentile female, the 5oth percentile male, and the 95Ih percentile male, as shown in\nTable 1 : Specifications for the ATD's.", + "Three types of seat legs were analyzed based on the FAA's proposed 32 G. They were straight leg, Sshaped seat leg and Y-shaped seat leg.\nStraight seat legs\nLumbar load for the straight leg seat found in the 5Ih percentile female, the 5oth percentile male and the 95Ih percentile male ATD's all exceeded their injury limits by more than 450 pounds.\nS-shaped seat legs\nDesigns iterations were carried out to find a suitable thickness of S-shaped seat legs that passed the proposed 32 G vertical test condition. A soft seat leg would easily \"bend-over\", which results both of its ends to meet together. In such a case, the seat pan stops immediately and results in a high impact load. A stiff leg on the other hand, will result in high lumbar load due to little amount of stroke. Figure 12 shows three lumbar load history curves of a 5oth percentile male ATD based on three seat legs. Straight leg design resulted in the highest value of lumbar load among the seat leg designs. The S-shaped seat leg designs with 0.12 inches thickness of the legs resulted in a higher lumbar load value than the S-shaped seat leg with 0.10 inches.\n4", + "Improved S-shaped Seat Legs - S-shaped seat leg with constant sectional thickness tends to show different performance for different size of occupants. An S-shaped seat leg with two thickness sections was developed to alleviate the problem of weight differences. This type of S-shaped seat leg has a soft structure at the upper section and a stiff structure at the lower section. During impact, a small occupant will benefit energyabsorbing from upper section of the seat leg and a large occupant will benefit energy-absorption from both section of the S-shaped seat leg. Figure13 shows the lumbar load history curves for the sth percentile female, the 5oth percentile male and the 9Sth percentile male on the Sshaped seat leg design with two section thickness. The 50Ih percentile male and the 95'h percentile male show two peaks in their lumbar load history curves. The first peak in the lumbar load occurs when the upper section of the seat leg deforms. The second peak is resulted from the deformation of the second section of the seat leg. Figure 14 shows the dynamic response of the 5dh percentile male ATD on the S-shaped seat leg design with two thickness section.\nFigure 13 : Lumbar load history curves for three ATD's on S-shaped seat leg with two thickness sections." + ] + }, + { + "image_filename": "designv6_24_0002164_taes.1986.310776-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002164_taes.1986.310776-Figure2-1.png", + "caption": "Fig. 2. Mode shapes. Radial displacement at outside radius.", + "texts": [ + " A single electrode covers the whole of the undersurface and is held at Earth potential by terminal To. BURDESS & WREN: PIEZOELECTRIC DISC GYROSCOPE THEORY 411 Since the disc is thin, a voltage applied to terminal TI will produce an axial electric field in that region of the disc which is defined by the shape of electrodes (1) and (5). If periodic, this field can be used to drive the disc into resonant vibration via piezoelectric action and can excite a combined radial and torsional mode of vibration in which the radial displacement has the form shown in Fig. 2. The nodal lines of this mode will occur at + 450 with respect to OX. Using the inverse piezoelectric effect, the modal response produced by this excitation can be measured directly by taking the current produced by electrodes (3) and (7) through a high input impedance amplifier A1. Because of the virtual Earth configuration of this amplifier (Fig. 3) terminal T2 may be considered to be held at Earth potential. If the excitation at T1 is now arranged to be determined by this current, via positive feedback it is possible to drive the disc into selfoscillation", + " The resulting vibration will take place at the resonant frequency of the mode and can be held at a preset amplitude by employing active gain control in the feedback circuit. The details of one such oscillator scheme can be found in [5]. A second set of measurement electrodes (4, 8) is provided by connecting T4 to a second amplifier A2 (similar to A,). Since these electrodes are centered precisely on the 450 nodal lines of the foregoing mode they will register no output current as a result of the oscillator vibration. If the disc is now rotated about OZ Coriolis, inertia forces will excite a secondary motion, as shown by the dotted line of Fig. 2. This motion will cause an output to be generated by A2. If the voltage applied to T4-and hence electrodes (2, 6)-is now provided by A2 via negative feedback, it is possible to drive the secondary motion to a null value. The value of the voltage at T4 is taken as a measure of the applied rate of turn. IV. ASSUMPTIONS AND CONSTITUTIVE RELATIONSHIPS In order to simplify the field equations which describe the mechanical and electrical behavior of the gyroscope, it is assumed that the disc is thin. This permits the mechanical problem to be treated as one of plane stress [7]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002990_s0094-114x(02)00037-x-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002990_s0094-114x(02)00037-x-Figure2-1.png", + "caption": "Fig. 2. Center-point-surface and circle-point-surface of a five-link independent suspension (solution 2 in paragraph 4) in perspective view (a) and top view (b).", + "texts": [ + " The method described above for solving the position problem it is likely to reduce this amount of time, since requires solving a system of only five nonlinear equations. P.A. Simionescu, D. Beale / Mechanism and Machine Theory 37 (2002) 815\u2013832 823 Considering the instantaneous motion of the five-link suspension, the wheel carrier motion relative to the car body is a screw motion of the circle-point-surface fixed to the wheel carrier with respect to the center-point-surface fixed to the car body [16]. The common tangent of these two surfaces is the instantaneous screw axis of the spatial motion (see Fig. 2), and corresponds to the points of minimum velocity of the wheel carrier relative to the car body. Therefore, the parameters positioning the momentary screw axis can very well be determined by formulating a minimization problem. A different approach is to solve the system of equations expressing the condition the linear velocity ( _x, _y, _z) of a point (x; y; z) attached to the wheel carrier is parallel to the angular-velocity vector (xx, xy , xz): _x=xx \u00bc _y=xy \u00bc _z=xz \u00f015\u00de 824 P.A. Simionescu, D" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002448_2011-01-0862-Figure13-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002448_2011-01-0862-Figure13-1.png", + "caption": "Figure 13. MAHLE 900 cc, twin-cylinder, Range Extender engine design.", + "texts": [ + " After carrying out an Eigen mode and response analysis a transfer matrix was created using MAHLE in-house tools, details of the inlet system acoustic analysis for this engine are described by Bassett et al. [29]. A previous study investigating the requirements for a range extender engine suitable for a C-segment E-REV found that approximately 30 kW is required for this application. This paper presents the results of a detailed design investigation which has culminated in the design of a 900 cc, twincylinder, 4-stroke gasoline engine. Figure 13 shows the final RE engine design with the main external dimensions and Table 2 summarises the key engine target parameters. During the detailed design process attention has been paid to minimising the engine package volume, weight and projected cost. Several engine architectures were examined, and it was found that the in-line, two cylinder, arrangement would yield the smallest package volume, lowest weight and lowest cost solution. To enable maximum installation flexibility the engine can be installed either vertically or horizontally mounted" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000512_icaiet.2014.27-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000512_icaiet.2014.27-Figure3-1.png", + "caption": "Figure 3. Design of air duct model by SolidWorks software.", + "texts": [ + " 1 depicts the fuzzy heat controller that requires capturing various fan speed and heating rate to adapt the air flow and temperature as they evolve during the CFD modeling. Embedded within this adaptive system approach is a desire to automatically identify and quantify the uniform airflow and airflow turbulence kinetic energy in order to reduce energy consumption. V. MODEL STRUCTURE AND SPECIFICATIONS The main objectives of this paper are to design the tubular heater (Fig. 2) and the air duct (Fig. 3) of the Carbon Nexus single tow research line at Deakin University using SolidWorks software and then simulate the airflow and heat transfer between heater and air flow with the help of the ANSYS/CFD package. The internal cross-sectional dimensions of the duct are 800mm x 270mm. The length of duct used for our design purpose is 1m; though, the true length of the duct is longer. The energy consumption of three phase electrical heater is 32 kW, with maximum current of 44.5 amp, and 415 V voltages at 50 Hz" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001255_3297097.3297110-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001255_3297097.3297110-Figure2-1.png", + "caption": "Figure 2. Mechanical model of the system", + "texts": [ + " rmwFN 2 )()sin()( 2211 2 21 LmLmwmmgAN (3) marFT )()cos()( 221121 LmLmaAmmg T (4) actuation In Figure 1, the lengths 1L , 2L and 3L , are the distances between the CoM of the body segment and the knee engines axis of rotation. Masses 1m and 2m are the leg mass and foot mass respectively. TA and NA are the tangential and normal components of the force reaction at the motor shaft and MT the motors torque. Equations (3)\u2013(5) are the movement equations applicable to the leg and foot actuated on the knee [14]. Taking into account equation (5) and Figure 2, we can obtain the relationship between the torque and speed of the system as: ,)()sin()cos()( 2 2 1 322211 a N NILgmLmLmgMM TM (6) where a is the acceleration in motor shaft, M is friction torque, MM is motor torque, 1N and 2N are the proportion of the torque between the motor drive and the stick, respectively. The motor torque MM , also called electromagnetic torque, is proportional to the current through the armature winding and can be written as [15]: ,ikM mM (7) where mk is the torque constant" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000083_ir-03-2018-0042-Figure17-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000083_ir-03-2018-0042-Figure17-1.png", + "caption": "Figure 17 The wearable vehicle while moving on uneven terrain with different amplitudes", + "texts": [ + "16N.s/ mm. The simulation analysis is carried in SolidWorks simulation study. The surface is uneven with different amplitudes as shown in Figure 16. The weight of the person impacts on the chair of the wearable vehicle. Through the simulation analysis, the value of damping coefficient, stiffness of the spring, length spring, Di, wire diameter dw and number of coils are provided from the program software. The simulation result is satisfied at different friction coefficients of the road as shown in Figure 17. The vertical line describes the distance between trunk mechanism and the ground, while the oscillation describes the amplitude of the uneven terrain. The wearable vehicle can across easily on uneven terrain and the shock absorber can reduce the fluctuation after passing uneven terrain which increases the level of comfort (Figure 18). Ccr \u00bc ffiffiffiffiffiffiffiffiffiffiffi k:msp p (25) z \u00bc C Ccr (26) The design of the wearable vehicle differs from the exoskeleton suit, because of the two modes of motion which is implemented in this structure" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000124_aps.2016.7696396-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000124_aps.2016.7696396-Figure4-1.png", + "caption": "Fig. 4, Detail of the patch feeding mechanism.", + "texts": [ + " 2 shows the main body of the Radar Instrument Structure (RIS), including the top deck with both the L-Band and the S-Band feed arrays. The L-Band and S-Band TRMs are located both inside and outside the RIS. Each L-Band tile is covered by a radome. The radome shell is made of a 1mm thick Astroquartz layer and is painted for thermal control and protection from atomic oxygen corrosion. Inside the radome shell, a foam layer 1379978-1-5090-2886-3/16/$31.00 \u00a92016 IEEE AP-S 2016 with cut-outs for the patches provides additional thermal insulation. Fig. 3 shows a single L-Band tile with and without the radome. Fig. 4 shows the patch feeding mechanism in more detail. The four feed probes in each stacked patch are soldered to the feed network board. They then support a dielectric bushing which supports the lower patch. An upper dielectric bushing in combination with a metal washer is used to capacitively couple power to the lower patch. The upper patch is directly soldered to the four probes and the center post. Since each patch pair is fed 300W of peak power when operated in circular polarization, power breakdown was a concern during the design" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003953_tia.1986.4504787-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003953_tia.1986.4504787-Figure1-1.png", + "caption": "Fig. 1. Construction of disk-type PM synchronous motor. (a) Section view. (b) Rotor. (c) Stator.", + "texts": [ + " Permanent-magnet synchronous motors have been increasingly used in machine tool and robotics applications which require motors with low inertia, high power/weight ratio, and flexible control for variable-speed operation. These motors can have cylindrical rotor construction or disk rotor construction. Each type has its own advantages and drawbacks with regard to construction difficulties, bulkiness, robustness, etc. The Laboratoire d'Electrotechnique de Grenoble has designed and built a disk-type PM synchronous motor that fulfills the foregoing requirements [3]. The structure of this motor is illustrated in Fig. 1. The rotor is a titanium disk with windows carrying samarium-cobalt magnets. The stator is built from two stacked copper disks on which the conductors are etched. The cavities between the conductors are filled with laminated iron forming the magnetic circuit. This structure is not the sole possible for disk motors. Other structures are found in commercially available motors. Disk-type PM synchronous motors have the following typical characteristics. 1) The air gap length is constant and large since the magnets and their supporting disk have the same permeability as air" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001573_j.apm.2014.11.058-Figure11-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001573_j.apm.2014.11.058-Figure11-1.png", + "caption": "Fig. 11. Magnetic flux lines distribution of the coaxial magnetic gear mechanism.", + "texts": [ + " In this study, a commercial FEA package Ansoft/Maxwell is employed for the 2-D magnetostatic field analysis. Fig. 9 presents the coaxial magnetic gear mechanism to be analyzed; it was created by the Ansoft/Maxwell 2-D pre-processor according to specifications listed in Table 1. Fig. 10 shows the finite element meshing of a coaxial magnetic gear mechanism with 11,300 elements. The mesh quality is high enough to guarantee the accuracy of the numerical computation. The distributions of magnetic flux lines excited by both permanent magnet rotors for the coaxial magnetic gear mechanism are depicted in Fig. 11. The magnetic flux density distribution of this magnetic gear mechanism is presented in Fig. 12. It can be found that the yoke of the low-speed rotor is magnetically saturated. The yoke radius of the low-speed rotor Ryl should be increased to avoid magnetic saturation for this design. Radial components of magnetic flux densities within the inner and outer air gaps, obtained using the FEA and the 2-D equivalent magnetic circuit network, are illustrated in Fig. 13(a) and (b), respectively. It is found that a good agreement exists between the analytical and the numerical computations" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001650_pcicon.2015.7435110-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001650_pcicon.2015.7435110-Figure8-1.png", + "caption": "Fig. 8 Five possible stator modes. Mode 4 is shown on the stator at the bottom.", + "texts": [ + " The stator slot passing frequency and associated force is applied to the rotor and is rarely a problem. The amplitudes in Figs. 3 to 7 are small but magnified to show more detail. These amplitudes will be higher for a loaded motor. Amplitude of 5 is not uncommon but represents no challenge to the motor reliability. The stator natural modes and its high frequency response is a fundamental part of the motor design against magnetic noise [7, 8]. Five possible stator core natural modes are shown in Fig. 8. The magnetic force harmonics in the air gap are fine-tuned by proper selection of rotor and stator slot combinations. However, the vibration will be more magnified if the excitation of RBPF or its harmonic side bands of +/- 2FL, 3FL or4FL, are in the vicinity of a natural mode that it can excite. The situation is further complicated by the fact that the stator natural modes shift with the addition of the housing frame and the supporting base. The spectra in Figs. 9 through 11 are high frequency vibration from a bearing pedestal of a 3000 KW 2 pole motor" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000099_12.373196-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000099_12.373196-Figure5-1.png", + "caption": "Figure 5. O erall AIRS Instrument illustration showing key hardare elements. The spectrometer is small compared to the olume required to support the space radiators and electronics.", + "texts": [ + " The heart of the instrument is a cooled (155 K), array grating spectrometer operating over the range of 3.7 - 15.4 p.tm at a spectral resolution (X/AA) of 1200. The spectrometric approach uses a grating to disperse infrared energy across arrays of high sensitivity HgCdTe detectors operating at 58 K. The concept requires no moving parts for spectral encoding and provides 2378 spectral samples, all measured simultaneously in time and space. Simultaneity of measurement is an essential requirement for accurate temperature retrievals under partly cloudy conditions. As shown in Figure 5, AIRS views the ground through a 282 Downloaded From: http://proceedings.spiedigitallibrary.org/ on 07/18/2016 Terms of Use: http://spiedigitallibrary.org/ss/TermsOfUse.aspx 2t Downloaded From: http://proceedings.spiedigitallibrary.org/ on 07/18/2016 Terms of Use: http://spiedigitallibrary.org/ss/TermsOfUse.aspx 284 Downloaded From: http://proceedings.spiedigitallibrary.org/ on 07/18/2016 Terms of Use: http://spiedigitallibrary.org/ss/TermsOfUse.aspx cross track, rotary scan mirror which provides ground coverage along with views to on board spectral and radiometric calibration sources every 2" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003682_eumc.2018.8541699-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003682_eumc.2018.8541699-Figure6-1.png", + "caption": "Fig. 6. FSS absorber 3D CAD model", + "texts": [ + " In the literature several publications on frequency selective surfaces (FSS) for microwave absorbers are reported [6], [3], [10]. These structures are often implemented using laminate-based processes like PCBs or similar. More advanced processes have also been developed targeted for microwave absorbers. In [11] a resistive sheet is utilized to make a thin meta-material absorbers. With 3D printing technology and G/PLA composite filament a high resistive and high permittivity material is available for making arbitrary 3D shapes not easily available with laminate-based solutions. In Fig. 6 a 3D model of the FSS unit element and 80 mm x 80 mm section and in Fig 10 the simulated reflection of the absorber with conductive back plate is shown. With a conductive back plate the measured reflection is now proportional to the inverse absorption. With an reflection of less than -10 dB from 6 to 8.5 GHz this absorber design is suitable for the UWB signal absorption. The FSS absorber was printed in 80mm x 80mm square sections and four sections was glued to a 160 mm x 160 mm metal backing plate" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001814_mfi.2014.6997693-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001814_mfi.2014.6997693-Figure3-1.png", + "caption": "Fig. 3 the sktech of the sensors board", + "texts": [ + " It is a 9-axis Motion-Tracking device that combines a 3-axis gyroscope, a 3-axis accelerometer, and 3- axis magnetometer in a small 3\u00d73\u00d71mm package. We use this sensor to make the sensor board, the size is 10\u00d715\u00d72.6mm, as shown in the Fig.2. Then the sensor boards are deployed on the each section of the hand. One sensor board is put on the palm, and each finger is deployed three sensor boards. There are 16 sensor boards are used to measure the hand pose, and the measurements of the sensor are sent to the processor to computer the pose. The sketch of the design are shown in the Fig.3. The green represents the sensor board, and the red represents the processor board. Thumb Index Middle Ring Pinky Palm Proximal phalanx Middle phalanx Distal phalanx x z y According to the three kinds of sensors, there are two independent ways to determine the attitude and heading. One is obtained from open-loop gyros. It has high dynamic characteristic, however, the gyro errors would create wandering attitude angles and the gradual instability of the integration drifting. The other way is determined from open-loop accelerometers and magnetometers" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001658_j.ifacol.2015.07.064-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001658_j.ifacol.2015.07.064-Figure4-1.png", + "caption": "Fig. 4. The photographs are taken from different positions using one camera", + "texts": [ + " In the equations (3) and (4) there are further used parametr pixel density ppmm which is computed as CCDR/CCDS . apx and bpx represent object positions in pixels relative to the sensor center. The position of unknown point C referenced to the focal point FA is calculated by equation (5). CFA (x, z) = ( e sin\u03b1 sin \u03b2 sin\u03b1+\u03b2 tan\u03b1 , e sin\u03b1 sin\u03b2 sin\u03b1+ \u03b2 ) (5) In practice, there is only one camera which takes the photographs from different positions, with different orientations and in a different time. The situation is illustrated in Fig. 4. It should be noted that the above equations are valid in the case that a flight altitude is constant and the camera sensor is horizontal to the ground in all cases. It is also required that the measured object does not move during shooting. The position and orientation of the camera (sensor) can be measured using onboard GNSS receiver and intertial measurement unit (IMU). This approach in photogrammetry, when the exterior orientation is obtained using onboard sensors, is called direct georeferencing", + " e = eA + eB = d tan\u03b1 + d tan\u03b2 (1) Mostly is solved inverse problem when the triangle altitude d is unknwn and the distance of the points (or their spatial position) is known. In this case is used equation (2). 1 d = 1 e ( 1 tan\u03b1 + 1 tan\u03b2 ) = sin(\u03b1+ \u03b2) e sin\u03b1 sin\u03b2 (2) This method was widely used for measuring the position of the significant points in the landscape (hills, towers etc.) in the past. In times when there were no GNSS it was more accurate to measure angles instead of distances. Today we can measure spatial position very accurately by placing the GNSS receiving antenna, but aerial stereophotogrammetry enables passive contactless measurement. Fig. 4. The photographs are taken from different positions using one camera According to equation (2) two variables must be measured: the angles \u03b1 and \u03b2 and the distance e. Angle measurement is performed using camera where the angle is calculated on the basis of the position of the object on the photography. The angles are calculated by the equations (3) and (4). \u03b1 = 90 + tan\u22121 ( apx ppmm \u00b7 lf ) (3) \u03b2 = 90\u2212 tan\u22121 ( bpx ppmm \u00b7 lf ) (4) Camera sensor size CCDS , resolution CCDR and focal length lf are constant parameters which can be obtained from technical documentation or can be determined during a calibration", + " In the equations (3) and (4) there are further used parametr pixel density ppmm which is computed as CCDR/CCDS . apx and bpx represent object positions in pixels relative to the sensor center. The position of unknown point C referenced to the focal point FA is calculated by equation (5). CFA (x, z) = ( e sin\u03b1 sin \u03b2 sin\u03b1+\u03b2 tan\u03b1 , e sin\u03b1 sin\u03b2 sin\u03b1+ \u03b2 ) (5) In practice, there is only one camera which takes the photographs from different positions, with different orien- tations and in a different time. The situation is illustrated in Fig. 4. It should be noted that the above equations are valid in the case that a flight altitude is constant and the camera sensor is horizontal to the ground in all cases. It is also required that the measured object does not move during shooting. The position and orientation of the camera (sensor) can be measured using onboard GNSS receiver and intertial measurement unit (IMU). This approach in photogrammetry, when the exterior orientation is obtained using onboard sensors, is called direct georeferencing" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002052_tie.2007.892610-Figure16-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002052_tie.2007.892610-Figure16-1.png", + "caption": "Fig. 16. (a) Complete schematic of the prototype. (b) Photograph of the prototype.", + "texts": [ + " This is mainly due to the fact that a readily available test chip was used, which includes the PWM, the drivers, the second-order filter, and an internal clock at a fixed frequency of 115 kHz. Therefore, all these constraints made unpractical the design presented in Section IV, which was tailored to the loudspeaker presented in [5]. The opamp used in the current feedback circuitry was a Burr-Brown OPA-27, which was preferred because of its good noise figure. The power stage uses STP14N N-channel diffused metal\u2013oxide\u2013semiconductor (DMOS) and STP12P P-channel DMOS, both by STMicroelectronics. Fig. 16(a) shows the complete schematic of the prototype used for measurements, whereas Fig. 16(b) shows a picture of it. Fig. 17 shows the measured gain versus load impedance for the prototype when the load resistance varies within the range from 1 to 100 \u2126. This graph confirms the behavior of the mixed feedback topology. Instead, Fig. 18 presents the circuit frequency response when two different resonant circuits were used as load. In this case, no target frequency response exists. The purpose of these measurements is just to show that the proposed system is capable of modifying its frequency response as a function of the load impedance and that it is able to do this regardless of the frequency at which the resonant peak occurs in the audio bandwidth" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000514_1.c034448-Figure10-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000514_1.c034448-Figure10-1.png", + "caption": "Fig. 10 Static aeroelastic deflection of distributed propulsion aircraft.", + "texts": [ + " The propulsive force and moments are given by fpz Xn i 1 \u03b4 x \u2212 xi \u2212mi \u22022w \u2202t2 \u2212miyi \u22022\u03b8 \u2202t2 \u2212mig \u2212 Ti sin\u039b \u2202w \u2202x Ti cos\u039b\u03b8 (21) mp x Xn i 1 \u03b4 x \u2212 xi \u2212Ii \u22022\u03b8 \u2202t2 \u2212miyi \u22022w \u2202t2 \u2212migyi \u2212 Ti sin\u039b \u2202w \u2202x yi Ti cos\u039bzi Xn i 1 \u03b4 x \u2212 \u03be \u2212Ti cos\u039b xi \u2212 x \u2202w \u2202x (22) mp y Xn i 1 \u03b4 x \u2212 xi \u2212Ti sin\u039bzi Ti sin\u039byi\u03b8 (23) where Ti is the propulsor thrust, mi is the propulsor mass, Ii is the mass polar moment of inertia of the propulsor, (xi, yi, zi) is the coordinate of the propulsor thrust center such that xi is the wing station along the elastic axis, yi > 0 forward of the elastic axis, and zi > 0 below the elastic axis, and \u03b4 x \u2212 \u03be , where 0 \u2264 \u03be \u2264 xi is the Dirac delta function that relates the distributed force fpz andmoments mp x and mp y to the concentrated force and moments at x \u03be asZ L 0 \u03b4 x \u2212 \u03be f x dx f \u03be (24) Note that the propulsive force and moments are influenced not only by the thrust force and the propulsor weight but also by the aeroelastic deflections. The aeroelastic deflection-dependent terms give rise to the thrust-induced stiffness. A three-dimensional beam finite-element model is developed for both static and dynamic aeropropulsive-elasticity. Themodel computes the aeroelastic deflections due to the aerodynamic and propulsive forces andmoments, as shown in Fig. 10, as well as the structural dynamic mode shapes and flutter speed. The aeroelastic wing shaping control analysis is conducted using a coupled analysis framework that includes the propulsion performance model, aerodynamic model, finite-element model, and automated geometry deformation tool that generates new deformed aircraft wing geometry from the output of the finite-element model.Fig. 9 Local wing elastic reference frame. D ow nl oa de d by U N IV E R SI T Y O F M A N C H E ST E R o n M ar ch 1 2, 2 01 8 | h ttp :// ar c" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003715_aero.2004.1368030-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003715_aero.2004.1368030-Figure1-1.png", + "caption": "Figure 1 - Six Discrete Spacecraft Docking With A ReSupply And Transportation Mother-Ship", + "texts": [], + "surrounding_texts": [ + "The proposed approach to system software for reconfigurable spacecraft focuses on the use of a virtual machine, executed on all participating node control processors. Each node control processor runs an instance of the virtual machine with the same instruction set, memory topology, input-output access, and mechanisms for guaranteed performance.\nIn addition to enabling the basic re-configurable spacecraft mission, a properly implemented virtual machine itself adds new capability to the platform. This paper will explore the advantages of a virtual machine based software system for re-configurable spacecraft.\nTABLE OF CONTENTS 1. INTRODUCTION ................................................................ 1 2. RE-CONRGURABLE SPACECRAFT s/W DEMANDS ......... 2\nSPACE-FLIGHT SOFTWARE ................................................. 2\nSYSTEMS .............................................................................. 3\n6. HARD REAL TIME SUPPORT ........................................... 4\n5. OPTIMIZATION FOR CONTROL SYSTEMS APPLICATIONS4\n7. TIGHT COUPLING OF DEVELOPMENT TOOLS, VIRTUAL MACHINE, AND APPLICATIONS ........................................... 5 8. CONCLUSION ............................................................ BIOGRAPHY ....................... .......................................... 6\n1. INTRODUCTION Recently, interest in satellite assets, which can assume a multiplicity of unplanned-for roles, has surfaced. The driving goal for these vehicles is a reduction in response time to global events. To accommodate this need, a new generation of spacecrafi will need to be created, ones that can rearrange existing capability and capacity to handle the dynamic role they are assigned to.\nIt has been identified that the most effective way to solve this requirement is to implement a system composed of discrete vehicles that can work together and maneuver as a team. Analogous to a special forces squad, the reconfigurable platform acts as an atomic unit and can adapt to new missions because that\u2019s what they\u2019re created for. Every member of a special forces team is trained in basics, navigation, communications, observation, and armaments. Likewise, each vehicle node is created fiom a common slice containing a core set of capabilities, including propulsion, communication, and computation. Most importantly, the core slice provides the needed elements to perform close proximity maneuvering and docking. These core\n\u2019 0-7803-8155-6/04/$17.0~ 2 04 IEEE \u2019 EEEAC paper#1352, Version I , Updatcd 1/13/2004\n0-7803-8155-6/04/$17.00 02004 IEEE\n1\n2364", + "capabilities are distributed to every discrete vehicle in the system, and are built upon to create specialty vehicles.\nWhile all members of a special forces team have a common training background, each specializes in a skill which the others must rely upon to complete a mission. These skills include long range shooting, artillery spotting, language, and leadership. Specialty nodes in a re-configurable satellite provide long-range communication, radar, refueling mechanisms, or optical tracking. To augment the team, specialty members may be integrated on a moments notice when the native skill set is insufficient to guarantee success.\nIt is in this context that this paper presents a computational foundation for a distributed real-time control system.\n2. RE-CONFIGURABLE SPACECRAFT SOFTWARE DEMANDS\nTruly re-configurahle spacecraft software makes new demands of system and software engineering. The very nature of re-configurability is at odds with traditional spacecraft deployment thinking. A fundamental assumption for most modern satellite missions is that the base run-time environment will not change in the course of normal operations. It is not expected that totally new features will be added to the system at run-time, while other features are removed. These assumptions are not just related to the payload or the mission parameters, but encompass the very hardware the software executes, including processing, communication, attitude, and power subsystems.\nThe re-configurable platforms recently proposed include multiple independent vehicles working together to achieve a single mission. The critical difference between these collective platforms and existing constellation systems is that the independent vehicles can mate and de-mate with each other and other vehicles. Complexity is added due to the desired autonomous nature of this interaction. Within existing constellations, such as Iridium or the Global Positioning System, the constellation is composed of elements that do not physically contact or maneuver relative to one other. When a re-configuration occurs and multiple nodes rearrange themselves, the controlling software must adapt to the new structure.\nA critical implication of the re-configurahle capability is that over time, new nodes may be introduced to the system, while others are retired. A system could theoretically be continuously upgraded, meaning it doesn\u2019t have a fixed end-of-life. This is only possible if the newer nodes can work seamlessly with the older nodes. A generation issue must be addressed, where backwards compatibility is assumed and built into the system from its inception. However, backwards compatibility must not handicap the system by arbitrarily limiting the hardware to a \u2018standard\u2019. As newer space-rated parts are made available and the\noriginal components end-of-lifed, upgrade nodes will improve in capability while they are still required to maintain backwards compatibility with the existing nodes. This long term upgrade path has not been possible with any other space vehicle, including complex systems like the Space Shuttle. As a result, this requirement is completely new to the industry, and can not be addressed properly with existing space software practices.\nWhile existing spacecraft have been designed with a software capacity for graceful degradation, and even onorbit reprogramming, this capability is usually implemented as mitigation against failure. There is an important philosophical architectural and design difference between planning for failure and building a vehicle for quick response to changes in operational mission. The primary difference for software is how software controls hardware.\n3. EXISTING SPACECRAFT HARDWARE CONSTRAINTS ON SPACE-FLIGHT SOFTWARE\nThere are two major design drivers involved in computer architecture for space flight, and a handful of secondary factors. The first is a lack of a foreseeable reduction in price per kilogram costs to launch a spacecraft. At a delivered-to-orbit price of $2000 to S22,OOO per kilogram, severe constraints are placed on the system engineering staff to reduce launch mass. The second, and arguably more powerful driver is the operational environment of the satellite, which implies a level of radiation, shock, and thermal protection. Without a volume market, space rated components which are able to survive the total radiation dose over the lifetime of the craft are expensive, rare, and lag the leading edge of terrestrial technology. These two factors, compounded by power and thermal restraints yields in a minimal performance system design, which provides the software engineering team with sparse computational resources.\nA perpetual shortfall of on-orbit computing resources, outlined above, has shaped the space industries view of software\u2019s role in space and their approach to software construction. This view tends to be very point-solution oriented, driven to optimal performance, and extremely horizontal in nature (unable to be adapted to other roles). The net result is a heritage of flight software that works well for the mission hardware of today. Unfortunately, the reality is that this approach will not work for reconfigurable spacecraft of tomorrow. The fundamental reason for the shortfall goes back to the lack of sufficient on-orbit computing power to implement serious levels of abstraction. Today, software control of flight hardware is so tightly coupled to the mission hardware due to the scarcity of computing resources, that small changes lead to large system redesigns, requiring costly human effort. Before generic software can be written for re-configurable spacecraft, the computing hardware must be made powerful\n2" + ] + }, + { + "image_filename": "designv6_24_0002388_2015-01-0090-Figure12-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002388_2015-01-0090-Figure12-1.png", + "caption": "Figure 12. Strain probes on a side rail in longitudinal direction", + "texts": [ + " In the simulation, vertical displacement of yf\u2212 yi (Figure 9) was assigned to the left end of the front axle. Figure 10 is a contour plot of equivalent strains on the truck frame predicted by the simulation. For convenience, lateral and longitudinal strains were defined to comprehend our discussion later on. The lateral strain was normal strain measured in a direction along the length of cross members near their ends as showed in Figure 11. The longitudinal strain is normal strain located on the parallel flanges of side rails (U-shaped channel), Figure 12. With difficulty in mounting the strain gauges to a truck due to limit access, it was necessary to attach a strain gauge inside the channel for the top flange and outside the channel for the bottom flange. of ramp Two cases of a one-wheel ramp condition were investigated, i.e. ramping of a front left wheel and a rear left wheel. Practically one left wheel was on the top position of the ramp while another right spring experienced its maximum bump travel. The vertical distance between the wheel centre and level road of the lifted wheel was 296 mm in height" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000512_icaiet.2014.27-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000512_icaiet.2014.27-Figure2-1.png", + "caption": "Figure 2. Design of tubular heating element model by SolidWorks software.", + "texts": [ + " An intelligent control system is built including a Fuzzy Heat Controller (FHC) to control the heater. Fig. 1 depicts the fuzzy heat controller that requires capturing various fan speed and heating rate to adapt the air flow and temperature as they evolve during the CFD modeling. Embedded within this adaptive system approach is a desire to automatically identify and quantify the uniform airflow and airflow turbulence kinetic energy in order to reduce energy consumption. V. MODEL STRUCTURE AND SPECIFICATIONS The main objectives of this paper are to design the tubular heater (Fig. 2) and the air duct (Fig. 3) of the Carbon Nexus single tow research line at Deakin University using SolidWorks software and then simulate the airflow and heat transfer between heater and air flow with the help of the ANSYS/CFD package. The internal cross-sectional dimensions of the duct are 800mm x 270mm. The length of duct used for our design purpose is 1m; though, the true length of the duct is longer. The energy consumption of three phase electrical heater is 32 kW, with maximum current of 44" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000932_icuwb.2016.7790496-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000932_icuwb.2016.7790496-Figure7-1.png", + "caption": "Figure 7. The physical layout of the lumped equivalent circuit of modified branch-line directional coupler.", + "texts": [ + " 5 can be evaluated using (1), and can be optimized by simulation with the results of Lm = 2.22nH, Lb = 3nH, Ld = 3.15nH, Ls = 1.66nH, C1 = 2.66pF, C2 = 3.2pF, C3 = 2.2pF, Cs = 1.66pF. The S parameter of the proposed coupler is shown as the dashed line in Fig. 8. The structure of the GaAs-based IPD process is shown in Fig. 6, where MIM capacitor is made up of the metal layer M1 and M2, and the spiral inductor is formed by M1 layer. The GaAs substrate, with a higher permittivity, makes the propoed coupler smaller. As shown in Fig. 7, the coupler in Fig. 5 is realized by GaAs-IPD process. The final layout of the proposed coupler has a size of 2.49mm\u00d73.60mm, and its performance is shown as the solid line in Fig. 8. The insertion loss of direct port and coupled port are 4.89dB and 4.93dB at 2.45GHz, and the maximum isolation within operating band is 60.1dB. By observing Fig. 8, we find that the S parameters of the lumped equivalent circuit of modified branch-line directional coupler agree well with its physical model implemented by GaAs-IPD process" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003423_ecce.2019.8913146-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003423_ecce.2019.8913146-Figure5-1.png", + "caption": "Figure. 5: Axial flux DFIG", + "texts": [ + " One of the major drawbacks of DFIG topology is that the power conversion circuitry is complex unlike the asynchronous generators and on top of that the slip rings connected to the rotor windings demand regular maintenance which proves to be inconvenient and expensive for ensuring long term operation of the system. Radial flux DFIG topologies are quite common in the literature as already discussed. The prime contribution of this work lies in making the introduction of AF-DFIG topology which to the authors knowledge is yet to be explored in detail. Before diving in the design details this subsection aims at laying out fundamental structures used for the work, for both axial and radial flux topologies. Fig. 4 shows a 2D view of the inner rotor radial flux DFIG topology used in this work, while Fig. 5 displays a single stator single rotor AF-DFIG 3D structure. Envelope dimensions, used for subsequent optimization, of the AF-DFIG are also marked in Fig. 5, where is outer diameter of the machine, is inner diameter of the machine and is the stack length of the machine. In addition to these parameters several other slot dimensions are optimized for the AF-DFIG as discussed in the next section. The fundamental difference between the two topologies lies in the air gap flux direction as clearly suggested by their names. III. AXIAL FLUX DFIG: DESIGN PROCEDURE The general-purpose sizing equation of axial flux machines take the form of equation (1) [18] for output power " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002584_1999-01-0782-Figure16-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002584_1999-01-0782-Figure16-1.png", + "caption": "Figure 16. FEA analysis of stamped wheel center design.", + "texts": [ + " Finally, the material between the spokes of each pair is removed and discarded. MATERIAL CONSIDERATIONS \u2013 The complexity of this design made further traditional analysis problematic at best, so our design iterations were carried out using intuition and finite element analysis. Typical of the wheels developed is the 4-spoke, 4 lug 5X14 design shown in Fig. 15 above. This wheel is intended to sustain a 1,588 Nm cornering load for a minimum of 200,000 cycles. Wheel stresses corresponding to this load are shown in Fig. 16. Peak stresses at the outer fiber are less than 315Mpa, a reasonable level for formed work-hardened steel. Some fairly high stresses are also evident at the base of the curved web between spoke pairs. Addition of shape features to the hub area in the vicinity of this web appears to reduce these stresses. Careful optimization of material thickness and details of spoke shape have resulted in designs with weight reduction of 10 to 15% percent based on the whole wheel weight, compared with stamped center designs for the same application" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003952_978-3-642-33457-3_7-Figure7.16-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003952_978-3-642-33457-3_7-Figure7.16-1.png", + "caption": "Fig. 7.16 Field lines and magnitude of the armature field modeled by 2D FEM", + "texts": [ + " Calculations of those electromagnetic parameters of the motor that are directly derived from the magnetic field are compared to the results of the analytical model. The FEM geometry and mesh is given in Fig. 7.15 and results are shown in Table 7.2. While the analytical representation of the permanent-magnet field is maintained\u2014 estimations of the PM flux density and the winding linkage of the PM flux of both models match very well\u2014the total phase inductance of the motor seems to be significantly higher than analytically predicted. According to the FE model the leakage of the armature field outside the stator iron is significant (see Fig. 7.16) which is not surprising taking into account large effective air gap of the machine. This result shows inadequacy of the developed analytical model to represent the armature field in a toroidally-wound machine. Rotor iron loss was modeled in a transient FE simulation for maximum expected current (2 A) and rotational frequency. The result partly confirms suitability of the magnetostatic approach in motor modeling since the losses in the rotor iron are extremely small. However, it was not possible to model eddy-current loss in the plastic-bonded magnet because no suitable means to represent the magnet electrical conductivity has been found" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002388_2015-01-0090-Figure15-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002388_2015-01-0090-Figure15-1.png", + "caption": "Figure 15. Illustration of a side rails experienced bending deformation", + "texts": [ + " case of Front-Left wheel on ramp (Lateral direction) case of Front-Left wheel on ramp (Longitudinal direction) To explain the discrepancy the thickness of the side rails had to be considered. The strain gauge at location 11 measured the strain at a bottom surface of the upper flange of the rail while the strain gauge at location 12 measured the bottom surface of the lower one. Once the parallel thick flanges were bent with the same curve, there are two appearances of surface definition, i.e. convex surface (stretching surface) and concave surface (shrinking surface) (Figure 15). The actual strains at location 11 and 12 were positive because both locations is on the stretching surface. However the side rails with thin shell assumption was invoked in the simulations. Consequently one side of flange would predict the positive strain with the negative strain on another flange due to no convex or concave surface definition for thin shell. Hence, the overall trend of these longitudinal strains was consistent in the sense of tending toward the right rail curving up (from location 11 and 12) and the left rail curving down (from location 13 and 14)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003635_s12206-009-0321-8-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003635_s12206-009-0321-8-Figure3-1.png", + "caption": "Fig. 3. Three dimensional beam element model.", + "texts": [ + " So, by dividing the finite element model into seven pairs of symmetric sides, a total of fourteen areas, the section properties, such as cross-sectional area, area moment of inertia and shear coefficient are extracted for constructing an FOA equivalent model. To verify the accuracy of the FOA model that only consists of beam elements, the modal properties obtained with the model are compared to those obtained with a full scale finite element model. The geometry and the coordinate systems employed for a beam element used for modeling are shown in Fig. 3. The element stiffness matrix of a beam element having arbitrary orientation in space is given as follows: 11 12 21 22 \u02c6 \u02c6 \u02c6 \u02c6 \u02c6 \u23a1 \u23a4 = \u23a2 \u23a5 \u23a2 \u23a5\u23a3 \u23a6 e K K K K k (1) where the 6\u00d76 sub-matrix is comprised of bending, axial and torsional stiffness matrices by direct superposition. The element mass matrix for the beam element is given by \u02c6 = \u222b\u222b\u222be T V \u03c1 dVm N N (2) In this study, a consistent mass matrix, which is derived from the same shape functions that are used to obtain the stiffness matrix, is used for modeling" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003926_tmech.2011.2121917-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003926_tmech.2011.2121917-Figure1-1.png", + "caption": "Fig. 1. Exploded view of the proposed 6-DOF positioner.", + "texts": [ + " Section II describes the design concept for achievement of the desired motion behavior in the new 6-DOF precision-positioning system. The force formulation, mathematical model, and dynamic behavior of this positioning system will be derived and discussed in Section III. Proper controller designs for a robust adaptive sliding-mode controller are developed in Section IV. To test the practical performance of the new design, the results of extensive experiments are provided in Section V. Finally, the conclusions are given in Section VI. The proposed 6-DOF precision positioner, seen in an exploded view in Fig. 1, includes a stator, a moving platen, and eight sets of electromagnetic actuators. Four sets of horizontal actuators, formed with rectangular magnets and rectangular coils, are used to provide the lateral forces for planar motion. The other four sets of vertical actuators, formed with circular magnets and cylindrical coils, are used to generate the vertical forces for vertical motion. The magnets are mounted on the moving platen, while the coils are mounted on the stator. With this kind of hardware arrangement, the moving parts can be made wireless and Coulomb friction effects can be avoided" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002748_j.ymssp.2015.11.019-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002748_j.ymssp.2015.11.019-Figure5-1.png", + "caption": "Fig. 5. Angles of inclination of the accelerometer, in xz and yz planes.", + "texts": [ + " This will result in errors being introduced in the following operations (Fig. 4). Detecting such kind of error sources is fundamental to have a robust and reliable starting dataset. The starting 0.84 s of the accelerations time-history, shown in Fig. 4, refer to conditions of null velocity for the vehicle; it is possible to notice that during this phase, the average values of the acceleration, ax0 and ay0 are non-zero and this could be caused by an improper installation of sensors. The angles of inclination of the accelerometer (Fig. 5), \u03bbx in xz plane and \u03bby in yz plane, can be estimated as follows: \u03bby \u00bc sin 1 ay0 g \u03bbx \u00bc sin 1 ax0 g \u00f01\u00de consequently, assuming a horizontal road, the measured values of the acceleration can be corrected taking into account of the cited angles \u03bbx and \u03bby: ax \u00bc ax ax0 cos \u03bbx ay \u00bc ay ay0 cos \u03bby \u00f02\u00de To ensure the availability of data useful for the correction of the offsets, the first step of the test procedure consists of 3 s of still standing acquisition. Additionally it is possible to zero any offset seen in the wheel speed sensors at this point" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000219_estc.2006.280002-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000219_estc.2006.280002-Figure3-1.png", + "caption": "Fig. 3 Sketches of (a) bonding wire and (b) solder bump, and (c) their equivalent circuit", + "texts": [ + " \u00a3die is the permittivity of the dielectric layer between the metal layers and tdie is the thickness of the dielectric layer. Modeling ofPackaging Parasitics Packaging parasitics always influence responses of circuits significantly, especially in RF and highspeed digital circuits [13,14]. In the design of high frequency/high speed SoP, parasitics of package must be considered. Wire bonding and solder bonding are two of the most popular chip bonding techniques. HFSS model of bonding wire and solder bump are shown in Fig. 3 (a), and (b), respectively. Bonding wire Because of its convenient fabrication process, bonding wire is widely implemented. In this work, bonding wires are first simulated with a 3D EM (electro-magnetic) simulator--HFSS, and then PBM are extracted in terms of the heights and lengths of bonding wires. The equivalent circuit of bonding wires and bonding pads are modeled as the circuit shown in Fig. 3 (c) in terms of the heights, lengths, diameters of bonding wires and the sizes of the bonding pads. Simply, parasitic inductance of a bonding wire could be roughly estimated as lnH/mm. Solder bump Solder bumps are preferred in RF and highspeed digital systems because of their weak parasitics. Electrical responses of solder bumps are also simulated with HFSS, and then PBM are extracted in terms of the diameters and heights of the solder bumps. The equivalent circuit of solder bumps is similar to that of bonding wires" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000850_j.matlet.2008.05.060-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000850_j.matlet.2008.05.060-Figure1-1.png", + "caption": "Fig. 1. Parallel arrangement of experimental set up for explosive welding process.", + "texts": [ + " In this study, it was investigated that microstructure and mechanical properties of explosive welded aluminum\u2013HSLA steel and aluminum\u2013dual phase steel. l rights reserved. Thechemical compositionsof thealuminumandHSLAsteel aregiven in Table 1. To produce dual phase steel the 100\u00d7150\u00d710 mm square HSLA specimens were intercritically annealed for 20 min at 724 \u00b0C (these temperaturescorrespondto~21%MVF)with instantquenching in icebrine. The parallel arrangement was used for experimental set up for explosivewelding as schematically shown in Fig. 1. Aluminumwas used as flyer plate and HSLA and dual phase steels were used as parent plate. Aluminum and steel plates were designed with dimension of 100\u00d7150\u00d72 and 100\u00d7150\u00d710 mm respectively. The ELBAR 5 (ammonium nitrate 90%, min 4.5% fuel-oil and min 3.0% TNT) was chosen as explosive material which was supplied by M.K.E. Barut Company, TR. The detonation velocity of the explosive material is 3000\u20133200 ms\u22121. All samples were ground and polished to 1 \u03bcm finish. A 3% nital solution was used for etching of HSLA and dual phase steel" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001472_9781118354179.auto089-Figure16-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001472_9781118354179.auto089-Figure16-1.png", + "caption": "Figure 16. Cross section and major components of the pull and push type clutch designs. (Reproduced with permission from ZF Friedrichshafen AG.)", + "texts": [ + " This is the most common type, but it Encyclopedia of Automotive Engineering, Online \u00a9 2014 John Wiley & Sons, Ltd. This article is \u00a9 2014 John Wiley & Sons, Ltd. DOI: 10.1002/9781118354179.auto089 Also published in the Encyclopedia of Automotive Engineering (print edition) ISBN: 978-0-470-97402-5 is important to mention that \u201cpull-type\u201d can also be found on a number of vehicles. The principle of operation for the two types is schematically shown in Figure 15 (for more information, see Shaver, 1997), with springs in the preinstallation shape indicated on (a). Figure 16 (from ZF Sachs) presents cross sections and major components of the two designs. There are no clear, overall advantages and disadvantages between the two designs, as not only the clutch assembly but also the actuating mechanism and installation must be considered. Still, from Figure 15, it is obvious that the pull-type clutch has a higher internal lever ratio, hence requires less force (but more stroke) for disengagement. For this type, it is more important to ensure concentricity of the (de-)actuating force; therefore, the thrust bearing is directly installed on the inner diameter of the spring \u201cfingers" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001778_2017-01-1563-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001778_2017-01-1563-Figure1-1.png", + "caption": "Figure 1(b). View of motorcycle with rider in rolling position", + "texts": [ + " The first section of this paper describes the mathematical model used for the numerical analysis of stability along with the results derived from it. The next section describes the test motorcycle used for the investigation and test experiment results, followed by the simulation results obtained from the Lotus Suspension Software. Then, the results drawn from the mathematical model, analysis and simulation are presented. In the last section, the conclusions drawn are mentioned. Also, future works that can be done on the instability of two wheeler vehicles are presented. Figure 1(a), 1(b) and 1 (c) depict the view of a motorcycle with a passive rider in upright position, rolling position and cornering position receptively. Let a motor cycle is moving at a roll angle \u201c\u03d5\u201d of mainframe, with a steer angle \u201c\u03b4\u201d of handlebar about the steering axis. The handlebar creates a rake angle \u201c\u03b5\u201d with respect to vertical line in the upright position, projecting a caster length \u201ctc\u201d in the ground. \u201cA\u201d is taken as the point of reference. It is located exactly below the centre of gravity of the mainframe, which moves forward with a velocity \u201cu\u201d and in lateral direction with velocity \u201cv\u201d" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002764_wocn45266.2019.8995132-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002764_wocn45266.2019.8995132-Figure3-1.png", + "caption": "Fig. 3. Geometry of stub loaded closed loop filter", + "texts": [ + " For the filter to be integrated in MMIC or in any planar circuitry the lumped model is modified to a planar configuration. To reach the proposed design, a closed loop microstrip line is taken as the primary design. This design fails to perform as a resonator. So T shaped stubs are introduced diagonally to the closed square loop. This leads to open ended transmission lines in the design thus providing band stop characteristics. The geometry of the designed microstrip line based stub loaded closed loop filter is shown in Fig. 3 and the parameters that affect the filter performance are length of the square loop \u2018Ls\u2019, width of the square loop \u2018W\u2019, length of the stub \u2018l\u2019, width of the stub \u2018d\u2019 and gap between the stubs \u2018S\u2019. The geometric and substrate parameters influence the filter characteristics considerably. The variation of each of these parameters are analyzed. At high frequencies the geometry of the structure is getting reduced, therefore the length of the square loop \u2018Ls\u2019 is having an inverse relation with frequency and is shown in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003415_iros.2014.6942881-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003415_iros.2014.6942881-Figure5-1.png", + "caption": "Fig. 5. Dimensions of the parts related to the estimation method", + "texts": [ + " By using these non-dimensional parameters, the attraction force between the magnets Fa is given by the following function, Fa(z) = \u03c0\u03bc0M2r2 M 1 \u2211 i, j=\u22121 i \u00b7 j \u00b7A(\u03b6 + i\u03c4m + j\u03c4s), (1) where \u03bc0 is the permeability of vacuum and M means the magnetization corresponding to the magnetic polarization of the magnet. A(\u2217) is the function defined as follows: A(\u03c9) = 1 2 \u2212 (2+ 0.5\u03c92)\u03ba\u03c9 2\u03c0 K(\u03ba2)\u2212 \u03c9 \u03c0\u03ba E(\u03ba2), (2) where \u03ba = \u221a 4/(4+\u03c92), K(\u2217) and E(\u2217) are complete elliptic integrals of the first and the second kind, respectively. Here, to evaluate the performance of the HMPM the hopping velocity of the rover generated by the HMPM just after impact is focused on. The hopping velocity is calculated by using (1). In advance of the calculation, the parameters of the HMPM are defined as follows (see Fig. 5). the distance between the stationary magnets, the inside length of the slide bracket and the thickness of the bracket are denoted by lS, lw and la, respectively. First, we consider the state in which the center of the movable magnet is located on the central position between the stationary magnets, as shown in Fig. 6. The origin of the calculation is set to the upper side of the movable magnet in the initial state, and the displacement of the movable magnet from the origin is denoted by x. The total magnetic force FM(x) acting on the movable magnet is obtained from the following equation, FM(x) = Fa ( lS 2 + lMs 2 \u2212 x ) \u2212Fa ( lS 2 + lMs 2 + x ) " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001529_iwat.2016.7434821-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001529_iwat.2016.7434821-Figure2-1.png", + "caption": "Fig. 2. (a) Illustration when wearing the antenna. (b) The antenna on two frequency dependent phantoms used in simulations to model the human body.", + "texts": [ + " Then, to extend the bandwidth in the low band, three slots with the length of less than a quarter of the effective wavelength at the lowest target frequency operation (4 GHz) are added. Lastly, an arc-shaped parasitic patch is positioned in the proximity of the main patches\u2019 radiating edges to improve the impedance matching throughout the bandwidth, therefore enabling a combined wideband resonance for the proposed antenna. The design is mainly intended for body-worn textile antenna candidate integrated into garments as illustrated in Fig. 2(a), thus the size of the antenna is not the main issue due to the large available area on clothing [11]. The antenna were simulated and studied using CST Microwave Studio 2014 full-wave simulator. To study the performance of the antenna in its real operating environment, i.e. near the human body, two simplified human muscleequivalent phantoms, rectangular and cylindrical phantoms, with frequency-dependent properties given in Fig. 3 were used. The antenna was located above both phantoms with a gap of d as depicted in Fig. 2(b). The cylindrical phantom was used to analyze the antenna performance when it becomes conformal to the human body surface. Thus, the total of the phantom radius (Rc) and the gap (d) denotes the bending radius of the antenna. This study is important as body-worn antennas need to be able to cope excellently with dynamic nature of the human body. The comparisons between the computed |S11| of the proposed antenna in free space and that on the phantoms are shown in Fig. 4. As can be seen in Fig. 4(a), the antenna has a robust input impedance against the human body loading showing by a very stable |S11| regardless of the change in the antenna position relative to the rectangular phantom" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001733_icelmach.2018.8507255-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001733_icelmach.2018.8507255-Figure1-1.png", + "caption": "Fig. 1. schema of the studied axial flux PM machine structure", + "texts": [ + " This stator iron cores can be made with Soft Magnetic Composite (SMC) pieces [17] (with a limited size for each of the SMC pieces) or with magnetic iron steel laminations rolled in a circular form (tape wound cores) [18], [19]. It can be noted that the stator core can also be made with an hybrid solution including both SMC and steel laminations as in [20]. Each stator disk supports independent multiphase windings (3 or more phases) which are located in the stator slots. An ironless rotor disk which supports alternatively axially oriented permanent magnets is located in the center of the structure between the two stator disks. So the machine have two airgaps (one in each side of the rotor). Figure 1 shows the geometry of the studied structure for a 18 poles device (in fig. 1, windings and rotor mechanical assembly are not represented for clarity reason). One of the interest of using this particular configuration is that the axial forces exerted on the rotor disk are balanced if the rotor remains perfectly centered. This allows to minimize the mechanical constraints which limits the use of PM Axial Flux machines. Another interest of such structure is to increase the intrinsic reliability of the system [4]. If a fault appears in one of the stator windings or cores, the system can be used at half power by disconnecting one of the two stators", + " The main dimensions of the prototype are given in table II. The two stator cores are made of ATOMET EM1 soft magnetic composite. The main physical properties of this material are described in [24]. Each of the stator is equipped with a quite common 2-layer tooth-concentrated winding where the number of slots per pole and per phase is spp = 0.5. This tooth-concentrated winding provides a fundamental winding factor of about 0.866 [25]. The ironless rotor integrating the magnets is located in the center of the system as illustrated by figure 1. Figure 6 shows a snapshot of the machine different parts. Rotor is located on the upper left corner of the picture and the two stators are located in the center (in up) of the picture. The distance between the two stators can be set with a screw system located in the machine. Figure 7 shows a picture of the prototype in the test bench. The black scroll wheels located in each side of the structure allow to tune the values of the air gaps. This prototype machine is mechanically coupled to a DC motor which for load or lead purpose" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001379_s10556-007-0042-8-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001379_s10556-007-0042-8-Figure1-1.png", + "caption": "Fig. 1. Bimetallic tube with rolled finned sheath.", + "texts": [ + " The surfaces of bimetallic finned tubes (BFT) are the main heat-transfer surfaces for tube bundles in heat-transfer parts of air cooling equipment (ACE) for natural gas as used at compressor stations of major gas pipelines, as well as in equipment for dividing oil into components, and in plant for petrochemical and gas-processing industries, and in nuclear engineering and thermal power. BFT are also used in tube bundles for calorifiers, condensers in large refrigerating and heat-pump plants, and in air conditioning systems for public buildings and living quarters. A BFT (Fig. 1) consists of a load-bearing tube 1 (mild or stainless steel, brass, copper, or melchior stainless steel) and a finned sheath 2 (made of material of high thermal conductivity, usually aluminum), which is mechanically coupled to the bearing tube. The finned sheath may be formed by rolling from a thick-walled tube or by winding with aluminum strip [1]. In the absence of a homogeneous metallurgical joint between the outer surface of the load-bearing tube and the inner surface of the finned sheath, there is a temperature drop \u2206tc [1], and thus there is a difference from a monometallic finned tube in that there is an additional thermal contact resistance (TCR) Rc = \u2206tc /qc = \u2206tc / (Q /Fc), in which qc = Q /Fc is the heat flux density in the contact zone of the BFT in W/m2; Q is the heat flux transferred by the tube in W; Fc = \u03c0doL is the nominal area of the contact zone surface in the tube in m2; do and L are the outside diameter of the load-bearing tube and the length in the contact zone" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001490_radar.2013.6585992-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001490_radar.2013.6585992-Figure1-1.png", + "caption": "Fig. 1. Antenna Configuration and Photograph", + "texts": [ + " But, the lack of sufficient effective designs for satisfying array demands while keeping UWB characteristics, directed us toward the design of this novel single element. This structure has been utilized to construct a 1 \u00d7 4 linear antenna array in order to achieve a better SLL and a more directive pattern than mentioned in previous designs. Simulation results are presented and discussed in detail. II. SINGLE ELEMENT A UWB antenna was designed initially. The top and bottom view of the proposed UWB antenna is shown in Fig. 1. It is in the x-y plane (W1 along x-axis and L1 along yaxis). The antenna contains a tapered microstrip line, a semielliptical radiating patch, and a defected ground structure. The width (W2) of the feeding microstrip line is set to have the impedance of 50 \u2126. Salient parameters of the proposed antenna are shown in Table 1. The substrate chosen here is TACONIC TLC-30. Its thickness is 1.58 mm, the relative permittivity ( r) is 3 and the metal cladding thickness is 35 \u03bcm. 2013 IEEE Radar Conference (RadarCon13) 978-1-4673-5794-4/13/$31" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003247_9781119258827.ch7-Figure7.5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003247_9781119258827.ch7-Figure7.5-1.png", + "caption": "Figure 7.5 (a) The Renk transmission (gearbox and steering) as used on Leopard 2 and (b) a simplified diagram of the steering system. Source: Courtesy of RENK AG.", + "texts": [ + " As shown the differentials are assumed to be 2:1:1 torque split (as used in automobile drive axles). For the Figure 7.4b arrangement, shafts C and D are shown driven by a 0.5 torque reduction from the drive and steer input shafts in order to depict similar input and output conditions for the two arrangements. For the Figure 7.4a arrangement, the differentials must be of the 2:1:1 type to ensure the necessary symmetry across the vehicle. The Figure\u00a07.4b arrangement is normally used with other torque split ratios. Figure 7.5 shows the layout of the Renk transmission as used in the Leopard 2 MBT, together with a simplified diagram of the steering system. A feature of note is the use of hydrodynamic couplings to assist the hydrostatic variable displacement pump/motor unit when slewing moments are high; this means the hydrostatic unit can be smaller and operate at lower maximum pressures. The speed difference between the output shafts is still controlled by the hydrostatic unit. Another feature is the use of a hydrodynamic retarder for use at higher vehicle speeds, with friction disc brakes operating at lower speeds" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002826_aps.2009.5172236-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002826_aps.2009.5172236-Figure3-1.png", + "caption": "Fig. 3. Normal E-field distribution of the proposed antenna.", + "texts": [ + "00 \u00a92009 IEEE of the proposed antenna structure, impedance matching is easily achieved and it shows a resonator frequency of 2.4 GHz. By changing the length of the stubs and slots on the ground plane, the operating frequency of the antenna can be changed. The L-shaped slot edge on the ground is optimized for more improving the bandwidth and radiation efficiency, and four slots are located symmetrically for achieving the balanced radiation power pattern. Figure 2 shows the geometry of the L-shaped slot on a ground plane. To achieve wide bandwidth the location of the microstrip line and L-shaped slot is very important. Figure 3 shows the simulated normal E-field distributions of the proposed antenna. It shows a balanced field distribution on the ground plane. Figure 4 shows surface current distributions. The normal E-field is concentrated between conductor strip line and L-shaped slot on a ground. The surface currents flow along the conductor strip lines and through the edges of the L- shaped open slots. The proposed antenna structure is optimized by the 3-D field simulation tool, CST MWS (Micro Wave Studio) [5]. Experimental results of the Fabricated Antenna Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002572_978-3-319-76276-0_5-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002572_978-3-319-76276-0_5-Figure6-1.png", + "caption": "Fig. 6 FE analysis of inner spring for its fatigue life and safety factor", + "texts": [ + " A stress ratio is given as input for fatigue analysis in ANSYS and it is given in Table 1. Stress-Life (S-N) curve with low cycle and high cycle fatigue life has to be provided in material property of spring material. Hence S-N curve has been plotted for ultimate shear strength of 1152.4 N/mm2 and endurance shear strength of 395.6 N/mm2 as shown in Fig. 5. Considering fatigue strength factor as unity, an equivalent alternating shear stress is determined and it is given in Table 1. The finite element analysis revealed the fatigue life and factor of safety contours as shown in Fig. 6 and the results are tabulated in Table 2. From Fig. 6 and from Table 2, it is observed that, the spring has finite life of 1.89 104 cycles. While examining the failed specimen as shown in Fig. 1, it has been observed that the cross section of failure resembles to that of fatigue failure is shown in Fig. 7 with the fatigue life. The fatigue life varies from 10 cycles to 105 cycles for the zone nearer to the inner side of the coil which indicates finite life. The progressively increases for the cross section slightly away from the inside diameter but still this region is having the finite life" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002814_tvlsi.2008.2008392-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002814_tvlsi.2008.2008392-Figure4-1.png", + "caption": "Fig. 4. Statistical channel model for a single column AND gate: (a) crossbar implementation and (b) channel model.", + "texts": [ + " It is thus important to evaluate the capability and limitation of defect-tolerant techniques in a rigorous manner. In this section, we present the central thesis of our approach, where we consider molecular crossbars as a defect-prone information processing medium and employ information-theoretic measures to investigate some key issues associated with defect tolerance in molecular electronics. A. Information Transfer Capacity for a Single Output Column We begin with an output column consisting of crosspoints, as depicted in Fig. 4(a). The logic function to be implemented is an -input AND gate, where the input bits can take any of the crosspoints. Thus, there are redundant crosspoints in this column that can be exploited for defect tolerance. The corresponding channel model of this AND gate is given in Fig. 4(b), where , i.e., the original input is while due to defects at one or more crosspoints the effective input seen by the output is . As shown later, the probability is a function of defects and may not be symmetric. We need to determine in order to obtain the mutual information and information transfer capacity. Without loss of generality, we may assume that the defects are independent and uniformly distributed. This is a commonly employed assumption for theoretical analysis, which allows us to focus upon the essence of the proposed method instead of the physical details of the defects" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001174_icit.2017.7913098-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001174_icit.2017.7913098-Figure4-1.png", + "caption": "Fig. 4. Flux density distribution for (a) preliminary, (b) minimum-weight, (c) minimum-loss, and (d) multi-objective designs.", + "texts": [ + " COMPARATIVE STUDY The specifications and dimensions of the preliminary design and the three optimized IPMSMs are given in Table IV. The minimum-weight design has the lowest weight of 7.353 kg while the minimum-loss design shows the highest efficiency of 93.624%. The efficiency and weight for the multi-objective design are 93.185% and 7.413 kg, which lie in between the minimum weight and minimum loss designs. A series of FEA are carried out using ANSYS-Maxwell to predict the performance of the designed machines. The physical structures and flux density distributions for four machines at rated power and speed are shown in Fig. 4. Maximum Torque per Ampere (MTPA) is employed to verify the speed range of the machines taking into account the current rating (5A) and the maximum available voltage (VDC=320V). ANSYS\u2019s electric machines design toolkit is used to evaluate the speed range and efficiency contour maps for each machine and the results are shown in Fig. 5. The preliminary design shows the widest speed range extending to 6000 rpm, followed by the multi-objective design with a range of up to 3600 rpm at rated power. The minimumloss and minimum-weight designs have speed ranges of up to 3750 rpm and 3150 rpm at the rated power, respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001095_20.280999-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001095_20.280999-Figure3-1.png", + "caption": "Fig. 3 Radd force density distribution for la alignment. (Stator TVR 1:l Rotor TVR 1:l)", + "texts": [ + " 1 is selected for the force calculation because any further increase in the number of element results in only a 5 % change in the radial force value. The radial forces are calculated for full alignment, 3/4 alignment, 1/2 alignment, 1/4 alignment and is the permeability of free space. 0018-9464/93$03.00 1993 IEEE 2414 mis-alignment positions between stator and rotor teeth for various tooth shapes, tooth to valley ratios and potentials. The d t of radial force density distribution for full alignment between stator and rotor teeth with rectangular tooth shape and stator and rotor tooth to valley ratio (TVR) of 1:l is shown in Fig. 2. Fig. 3 shows the radial force density distribution for 112 aligmmnt between stator and rotor teeth and Fig. 4 for mis-alignment between stator and rotor teeth. From the figures, it can be seen that when the stator and rotor teeth are in full alignment the radial force is high at the centre of the overlapping area between the stator and rotor teeth hence corresponding a high a radial force. On the other hand for half alignment, the overlapping area between the stator and rotor teeth is reduced, hence the radial force is lower than the full alignment situation" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002095_cobep.2013.6785220-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002095_cobep.2013.6785220-Figure2-1.png", + "caption": "Fig. 2. Synchronous machine model on MagNet.", + "texts": [ + " Figure 1 shows the comparison of the flux density components (normal and tangential) solution in the air gap obtained through FRM and FE method. In this method, the electrical machine is first modeled in finite element software accordantly to machine technical features (constructive parameters), therefore the effects of the magnetic circuit geometry, saliency, winding layouts and slot shape are included. Table I summarizes the parameters of a commercial salient pole synchronous machine used while Fig. 2 depicts the FE model for the target machine developed in the MagNet software. The FRM formulation is the same for different machine types (IM, PMSM, SRM). However the basis function determination procedure is particular for each one. The basis function determination steps for wound-rotor synchronous machine are described in [11] and [12]. Herein, since the focus of analysis is the generator mode, the FRM concept is used to create a flux reconstruction formulation, which allows the output voltage calculation" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001960_tec.2016.2609338-Figure17-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001960_tec.2016.2609338-Figure17-1.png", + "caption": "Fig. 17 Test rig for torque-angle characteristic measurement.", + "texts": [ + " CPM-MG is slightly severer than the SPM-MG1 since its Tratio is a little lower than that of the SPM-MG for all stack length. For instance, when the stack length is 60mm, Tratio of the CPM-MG and the SPM-MG1 are 82.5% and 88%, respectively. As a consequence, the advantage of high toque density of the CPM-MG is decreased due to the severe end effect. Therefore, the 3D verification should be conducted in the MG design process. The optimized CPM-MG and SPM-MG1, parameters of which have been listed in Table IV, are prototyped and investigated on the test rig shown in Fig. 17. In the test, both the magnetic modulation ring and the outer rotor are fixed at still with an arm and a mass. The inner rotor is held by a dividing head. By turning the dividing head, the inner rotor can rotate precisely. At each position of the inner rotor, the torque on the outer rotor is measured with a scale, using the calculation of product of the net force on the scale and the arm length. The torque-angle characteristics can thus be obtained, and the pull-out torque which represents the MG torque transmission capability is just the maximum value of the curve" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002392_smasis2018-8132-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002392_smasis2018-8132-Figure2-1.png", + "caption": "Figure 2. Composite layout with plystack showing the fiber orientation for each layer for the [0\u00b0/45\u00b0]s case. The top three layers correspond to the top piezoelectric stack, the middle three layers correspond to the GFRP layer stack, while the bottom three layers correspond to the bottom piezoelectric layer stack.", + "texts": [ + " The top three layers were assigned PZT-5A, the three central layers were assigned S2 glass fiber, and the bottom three layers were again assigned PZT-5A. The fiber orientation alternated between the layers, depending on the assigned stacking sequence as shown in Table 3. The fiber orientation was chosen as general stacking of [0\u00b0/45\u00b0]s. Previous research has shown that this stacking sequence is the most appropriate for leaf springs, as first-ply failure occurs at higher loads than in other stacking sequences [7]. Figure 2 shows the stacking sequence and ply numbering sequence. The simulations were then repeated for a beam comprised entirely of GFRP-layers for comparison and to determine the effect of the piezoelectric layers on the stress and deflection. Table 2. Material properties of composite GFRP and piezoelectric layers. S2 Glass/Epoxy Elastic Modulus (GPa) E11 43 E22 = E33 8.90 Poisson\u2019s Ratio \u03bd23 0.45 \u03bd13 = \u03bd12 0.27 Shear Modulus (GPa) G23 3.07 G13 = G12 4.50 PZT-5A Elastic Modulus (GPa) E11 (fiber axis) 30" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002877_tencon.2019.8929336-Figure2.1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002877_tencon.2019.8929336-Figure2.1-1.png", + "caption": "Fig. 2.1: Launch vehicle geometry in pitch plane", + "texts": [ + "This paper discusses an LQG based load relief control for a 13th order model of a launch vehicle combining rigid body, flexibility, actuator, nozzle and sensor dynamics using the estimates of the unmeasured states for control. II. SYSTEM MODELING Despite of a launch vehicle being a highly flexible body, the initial phase of controller design begins with the rigid body representation of the vehicle so as to reduce design complexity. A rigid body model gives the basic response characteristics. The vehicle motion is expressed as short period dynamics [14]. The launch vehicle geometry in pitch plane is as shown in Fig. 2.1. All coordinate systems we associate with the vehicle are conventionally taken to be right handed. A body fixed right handed coordinate system is fixed to the vehicle with origin at the geometric center of the vehicle. The motion of the vehicle in pitch plane is described with reference to an inertial coordinate system. A. Rigid Body Dynamics The nonlinear equations of motion for a launch vehicle were derived from the Newton\u2019s second law of motion [4]. ( ) I F d mV dt = (1) ( ) I H Hd dt = (2) where, the sub-script \u2018I\u2019 refers to the time rate of change of the vector with respect to inertial space" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001929_20.120032-FigureI-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001929_20.120032-FigureI-1.png", + "caption": "Fig. I. Cross section of the designed PM motor.", + "texts": [], + "surrounding_texts": [ + "V.S. Ramsden, \"Permanent magnet motor developments and markets,\" JOWMI ofEIectrica1 and Electronics Engineering, Australia, vol. 9, No. 3, pp. 118-123. September 1989. M.A. Rahman, G.R. Slemw, \"Promising applications of neodymium boron iron magnets in electrical machines,\" IEEE Trans. on Magnetics, vol. 21, No. 5, pp. 1712-1716, September, 1985, A. Ishizaki, Y. Yamamoto, \"Asynchronous performance prediction of AC permanent magnet motor,\" IEEE Trans. on Energy Conversion, vol. I, No. 3, pp. 101-108, September 1986. E. Richter, T.J.E. Miller, T.W. Neumann, T.L Hudson, \"?he ferrite permanent magnet AC motor - a technical and economical assessment,'' IAS Annual Meeting, pp. I3S3-1358,1984. S.A. Nasar, \"Handbook of electric machines,\" McGraw-Hill Book Company, 1987. \"The PE2D Rference manual,\" Vector Fie& Ltd.. Ogord, England, 1991. J. Luomi, \"Numerical field analysis,\" Lecture material, Tampere University qf Technology, Tampere. I990 (in finnish)." + ] + }, + { + "image_filename": "designv6_24_0000855_0954406220936732-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000855_0954406220936732-Figure1-1.png", + "caption": "Figure 1. Structural of heavy-load-bearing flexible force sensor.", + "texts": [ + " Based on the superposition principle under small deformation, starting from the flexibility matrix of double-axis flexure hinge and the stiffness matrix of rectangular beam, the whole stiffness matrix of the sensor and the mapping relation between each limb and the six-dimensional external force are calculated. The theoretical deduction process is verified by the numerical calculation and finite element simulation, and the prototype is calibrated. The design concept of hybrid limb and the method of precise force mapping model can provide a new idea and theoretical basis for the design of multidimensional force sensors. Structural characteristics of the hybrid limbs parallel force sensor The configuration of the hybrid limbs heavy-loadbearing flexible parallel force sensor is shown in Figure 1. Unlike the conventional Stewart platform parallel mechanism, the force sensor adopts the redundant configuration of 8-SPS flexible limbs, flexible link acts as prismatic joints, the spherical joint adopts the biaxial circular flexure hinges. The eight limbs are divided into four groups and distributed around the platform symmetrically. Four load-bearing limbs of the configuration of rectangular beam are interlaced arrangement with the flexible limbs symmetrically. The measuring limbs are of the same size, and so are the bearing limbs" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002878_eej.22684-Figure12-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002878_eej.22684-Figure12-1.png", + "caption": "Fig. 12. Prototype.", + "texts": [ + " 10, the current amplitude slightly exceeds its target of 4 A, which is explained by the integral deviation remaining in PI control. These results were obtained by adjusting the integral gain at a fixed load increase rate. The phase difference between the rotors at the target phase currents of 4.24 Arms and 2.83 Arms is shown in Fig. 11. As can be seen from the diagram, the high-speed rotor slips at the target phase current of 4.24 Arms but not at 2.83 Arms. 4.1 Prototype motor and N\u2212T characteristics measuring system We built a prototype magnetic-geared motor (Fig. 12) and conducted experiments using an N\u2212T characteristic measuring system (Fig. 13). The prototype motor was connected to a hysteresis brake via a torque meter, and the characteristics were measured while increasing load from rotation in the no-load state. 4.2 Measured N\u2212T characteristics First, we applied sufficient load to the brake and measured the maximum transmission torque while rotating the high-speed rotor with a servo motor, with the low-speed rotor fixed. The angular difference between the rotors was zero" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002188_amr.338.491-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002188_amr.338.491-Figure5-1.png", + "caption": "Fig. 5 Forth-order Fig. 6 Fifth-order Fig. 7 Sixth-order vibration mode vibration mode vibration mode", + "texts": [], + "surrounding_texts": [ + "Harmonic response analysis is to calculate the exciting force frequency response of structure, the response displacement , response stress, and obtain the curve of the dynamic response of the system and system vibration frequency,which is called amplitude-frequency [5]. After modal analysis, the amplitude of three directions of node K is large. frame of the maximum displacement at the upper front of the frame, Therefore, in this region pick up a node K excitation force applied, amplitude 20N, frequency (50 ~ 500) Hz, Divided into 10 steps for harmonic analysis, Obtained amplitude-frequency curve of the node K, shown in Fig. 8. Abscissa is the frequency. the vertical axis is the displacement. Can be seen in (250 ~ 350) Hz the amplitude of the three directions of node K is large, while a sharp change in response to the displacement, Thus resonance may occur in the 300Hz or so, should be taken to avoid resonance." + ] + }, + { + "image_filename": "designv6_24_0000318_detc2005-85463-Figure15-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000318_detc2005-85463-Figure15-1.png", + "caption": "Figure 15 : F.E. Model for LS-DYNA.", + "texts": [ + " ]1][ln1][[ * 0 mn TCBA \u2212++= \u03b5 \u03b5\u03b5\u03c3 & & (15) roommelt room TT TTT \u2212 \u2212 =* sec/1=o\u03b5& 1st term : strain hardening 2nd term : strain rate hardening 3rd term : thermal softening Figure 14 shows true stress-strain curves from Hopkinson bar experiments in order to achieve dynamic material properties of impact body with high speed. As strain rate is increasing from near 0 to 5000/sec, material characteristics are changing considerably because of hardening effect. 5 Copyright \u00a9 2005 by ASME rms of Use: http://www.asme.org/about-asme/terms-of-use Do The quasi-static analysis applied Johnson-Cook model is carried out and predicted the impact force prediction using commercial program LS-DYNA. Figure 15 shows LS-DYNA finite element model. Figure 16 is LS-DYNA explicit finite element analysis result. Compared with the experimental data, Figure 16 draws Johnson-Cook analysis considered strain rate is more probable than general elastic analysis. The contrast between JohnsonCook analysis considered strain rate and general elastic analysis illustrates the importance of the dynamic material properties, which should be performed the Hopkinson bar experiments. However, it costs high and takes a long time to be accomplished" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001046_062024-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001046_062024-Figure2-1.png", + "caption": "Figure 2. modularized vertebral detection and therapy instrument. (a) A sectional view of the assembly. (b) Sectional view of the connecting part of the rear module and the middle module. (c) Sectional view of the connecting part of the middle module and the front module.", + "texts": [ + " Depending on the degree of stiffness of the body part, the peak value of the reaction force received by the sensor 2 is different, thereby determining whether the part needs treatment. During the treatment of the patient, when the preload force received by the sensor 1 reaches a predetermined value, the electromagnet pushes the massage head to hit the diseased part at a predetermined frequency and number of times, and the pulse force triggers resonance of the human tissue to achieve the therapeutic effect. Figure 2(a) is the actual assembly diagram of the vertebral detection and therapy instrument. When in use, the three modules are assembled into a vertebral detection and therapy instrument with complete functions. Each module is circled with a dotted frame. Figure 2(b) is a cross-sectional view of the connecting part of the rear module and the middle module, in which the slider slides into the slot under the push of the spring to connect the middle module and the rear module. Figure 2(c) is a crosssectional view of the connection between the middle module and the front module. The two modules are connected by threads. The stop part is used to avoid the thread looseness caused by vibration and improve the reliability of the connection. ISCME 2020 Journal of Physics: Conference Series 1748 (2021) 062024 IOP Publishing doi:10.1088/1742-6596/1748/6/062024 The function of the front sensor is to judge the stiffness of the patient's muscles by reading the value of the sensor 2 during detection" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003952_978-3-642-33457-3_7-Figure7.24-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003952_978-3-642-33457-3_7-Figure7.24-1.png", + "caption": "Fig. 7.24 Stress in the magnet at 200.000 rpm, 85 \u25e6C", + "texts": [ + " However, what concerns is a large stress concentration at the line close to boundary between the iron shaft and glass fibre ring (pointed by arrows at Fig. 7.22). Further, according to this model, if the rotor remained at room temperature, the contact between the magnet and iron would be lost beyond 180.000 rpm (Fig. 7.23). Maximum possible speed is increased, though, if the operating temperature rises. Maximum equivalent stress in the magnet at the maximum speed and temperature is at the outer magnet surface and amounts to 110 MPa (Fig. 7.24) which is still below the compression limit. Maximum tensile stress in carbon fibres\u20141147 MPa\u2014is in a very good agreement with results from 2D modelling. The chapter presents the design of the high-speed spindle motor, from a conceptual design to electromagnetic and structural optimization of the motor. Two new spindle concepts6 are presented; in both concepts a slotless toroidallywound PM motor with a short rotor is combined with 5DOF frictionless\u2014active magnetic or aerostatic\u2014bearings. The motor is spatially integrated with bearings without merging their functions even in the example of active magnetic bearings" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000587_0029-5493(65)90101-9-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000587_0029-5493(65)90101-9-Figure2-1.png", + "caption": "Fig. 2. Yield polyhedron.", + "texts": [ + " (8) This y ie ld locus has the o r ig in of coo rd ina t e s as a c en t e r of s y m m e t r y . I t has uniquely d e t e r - mined suppor t ing p lanes at a l l po in ts except those on the p a r a b o l i c a r c s (4) and (5), the s e g - ment CD, and the poin ts obtained f rom these by s y m m e t r y with r e s p e c t to the or igin . An exact y ie ld locus for a sandwich she l l , if M 0 and N O a r e t aken to r e p r e s e n t the y ie ld m o - ment and the y ie ld fo rce of the sandwich she l l , was d e s c r i b e d by Hodge [9] and is shown in fig. 2. If, however , M 0 and N O a r e given the va lues c o r r e spond ing to a so l id she l l , th is po lyhedron r e p r e s e n t s an approx ima t ion to the exact y ie ld locus of fig. 1. The c o r r e spond ing f a c e s of the po lyhedron l ie in the fol lowing p lanes : face I n O = 1 II n O - n x = 1 H I n x - m x = - 1 I V 2 n 0 - n x + m x = 2 V 2 n 0 . n x - m x = 9 (9) face I nO = 1 II n O - n ~ o = 1 III m ~ o = l i v n \u00a2 = 1 Other p l anes can be obtained by s y m m e t r y with r e s p e c t to the or ig in " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003205_ecce.2014.6954109-Figure15-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003205_ecce.2014.6954109-Figure15-1.png", + "caption": "Fig. 15: Optimal machine design with ANN", + "texts": [ + " 13 that there is very good agreement between the predicted results using ANN and the coupled model data. The trained network has been subsequently built into the optimization program. An EM-only model combined with the predicted maximum current densities provided by the trained ANN is sufficient to replace the full coupled EM/thermal model for the remaining 1,200 designs after the initial 300 designs required for training. This updated optimization with the ANN has been run, and the resulting optimal design (Fig. 15) is compared with the optimal design from Section III (Fig. 14). The key parameters of these two machines are listed in Table IV. The close match between the two designs demonstrates the usefulness of the ANN technique for further accelerating the optimization process. As discussed in Section III, the total number of analyses required to accomplish the optimization without ANN is approx. 7,500 transient EM plus 7,500 static thermal FE analyses. Applying the ANN technique, the machine design parameters and their associated maximum current densities found from the first 10 generations \u2013 that is, the first 300 designs \u2013 can be used to train a network" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000372_tmag.2010.2073455-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000372_tmag.2010.2073455-Figure1-1.png", + "caption": "Fig. 1. Annular ring microstrip patch antenna on metamaterial substrate.", + "texts": [ + " This paper is organized in the following way. In Section II, the boundary value problem is formulated in the Hankel transform domain leading to the dyadic impedance Green\u2019s functions. Galerkin\u2019s technique is used to solve the unknown currents and resonant frequencies. In Section III, numerical results are presented and discussed to verify our theoretical predictions. The conclusions are presented in Section IV. The cross-section view of the proposed antenna employing a metamaterial as substrate is depicted in Fig. 1. The annular ring patch is taken as perfectly conducting and has inner radius and outer radius . The ground plane is assumed to be infinitely long and perfectly conducting. Dielectric region (1) is air-filled, with and denoting the free-space values of the permittivity and permeability, respectively. The metasubstrate (region 2) with thickness is anisotropic and characterized by uniaxial permittivity and permeability tensors given by [3] (1) The electric and magnetic Hertz potential vectors, and , respectively, are defined along the -direction" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000431_j.engappai.2004.09.003-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000431_j.engappai.2004.09.003-Figure5-1.png", + "caption": "Fig. 5. Optimized stator/rotor lamination.", + "texts": [ + " Korous\u030cic\u0301-Seljak / Engineering Applications of Artificial Intelligence 18 (2005) 47\u201355 53 lengthy process. It took around 3000 runs for the optimization to converge, and since one finiteelement analysis took about 7min on a Pentiumbased computer, the whole optimization lasted for 2 weeks. Most of the solutions that were given by the DOptiMeL program show a significant reduction of the iron and the copper losses, in comparison with the losses in the initial motor. The best solution results in a power-loss reduction of 24%, and gives us a motor with iron and copper losses of 239W (see Fig. 5). The main differences between the initial design (Fig. 4) and the optimized design (Fig. 5) are: (a) the height of the rotor-and-stator laminations is increased by 13%, (b) the rotor radius is increased by 5%, (c) the slot (copper) areas in the stator and the rotor are larger, and (d) the iron area in the rotor is larger. A comparison of the magnetic flux densities in the initial and the optimized motor shows a clear reduction of the areas with the highest levels of magnetic flux density in the optimized motor. In the optimized lamination, the copper losses in the rotor and the stator are significantly lower than in the initial lamination" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002886_kem.373-374.770-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002886_kem.373-374.770-Figure2-1.png", + "caption": "Fig. 2 Sketch of robotic polishing process Fig. 3 Elastic deformation of the soft tool", + "texts": [ + " In the rapid spray metal tooling, the metal film is formed by plasma spraying and acts as working surface of the rapid metal tool. The material of the metal film is Ni-Fe alloy, which is wearable and its hardness can reach HRC50. So the soft polishing tool and free abrasive are appropriate choice, which can decrease the fluctuation caused by the polish force and the robotic moving path as far as possible. During the robotic polishing process, the angle \u03b1 between the tool axis and normal direction of the polished surface should be set between 5\u00b0 ~ 15\u00b0 to make the polishing process more smoothly, which is shown in Fig. 2. And the common material of the soft polishing tool used in this paper is unwoven cloth or wool, so the compensation for the elastic deformation and the abrasion during the polishing process of the soft tool is the key to the control of the robotic polishing path. In order to test the tool character of elastic deformation, the width of the tool impression is measured by different pre-compressed value along the normal direction of the surface. The material of the soft polishing tool is unwoven cloth, the diameter is 30mm, and the scope of the pre-compressed value is from 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002605_973233-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002605_973233-Figure5-1.png", + "caption": "Fig. 5. A Circular Segment", + "texts": [ + " A straight segment shown in Fig. 4 is used to demonstrate how the compliance is derived. The segment is subject to two loads: a force p and a moment pa during insertion, where a is the length of moment arm. The compliance ej from Eq. (8) can be calculated as where Mj is the moment at any cross-section, s is the distance along the segment, El is the flexural stiffness of the clip, and L and or are respectively the length and the angle of the segment. For the circular segment under the same loads (see Fig. 5), the compliance ej can be calculated as 3 c: + 3c; g2 = r [2c1c2 + (--- )8 - 2ci sine - 2clc2 cose 2 2 1 2 -CIC~ sin 0 + -(c2 - c;) sin281 4 (13) where p is the initial angle, and 8 is the total angle of the segment. The maximum insertion force is known to occur between the points A and B as shown in Fig. 6, where point A (8 = ei) is the beginning of the arc and point B (0 = 0) is where the largest clip deflection occurs. The clip deflection is the amount of the interference between valve body and clip opening" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002224_3527608117-Figure5.4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002224_3527608117-Figure5.4-1.png", + "caption": "Fig. 5.4 Principle of powder flame spraying (Linde AG).", + "texts": [ + " However, oxidation reactions occur unfavorably during coating of metallic sprayed layers at high temperatures. 5.2.6 Thermal Spraying Processes 5.2.6.1 Flame Spraying During flame spraying a fuel gas oxygen flame is used as the heat source, into which the spraying additive (powder, wire or rod shape) is inserted. Acetylene is mostly used as the fuel gas, but ethene, methane, propane, propylene, natural gas or hydrogen may also be used. 5.2.6.1.1 Powder Flame Spraying In powder flame spraying (Fig. 5.4) only small particle speeds are obtained (<50 m s\u20131), with a relatively long interaction time between hot fuel gases and powder particles. Powder flame sprayed layers therefore show a relatively high porosity 120 5 Production of Composites or Bonding of Material by Thermal Coating Processes (5 %), a high oxide content and a high gas cavity. In general, spraying powders with a diameter between 20 and 100 \u00b5m are used. For metallic spraying powders deposition rates between 3 and 6 kg h\u20131 are obtained, while for ceramics 1 to 2 kg h\u20131 is common" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003205_ecce.2014.6954109-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003205_ecce.2014.6954109-Figure7-1.png", + "caption": "Fig. 7: EM-only optimal machine design [6]", + "texts": [ + " This configuration can accelerate the computation speed approx. 5 times higher than with a singlecore computer. The optimization was run and converged after the 50th generation, with a total number of 1,500 designs evaluated. The best design has a total active mass (stator and rotor) of 13.5 kg. It can continuously produce the required torque at its maximum current density of 9.84 A/mm2. The optimized machine using the coupled EM/thermal model (Fig. 8) has a mass reduction of 7.17 kg compared to the machine optimized previously using the EM-only model (Fig. 7), in which the current density is a fixed value of 4.6 A/mm2. The key parameters of these two machines are listed in Table III, demonstrating a significant improvement in torque density for the machine using the coupled EM/thermal model. A large share of this improvement can be attributed to the action of the coupled model to significantly raise the current density within the maximum temperature limits of the winding insulation, set at 155\u2070C for this exercise (assuming Class F wire insulation). As noted above, the optimization evaluated 1,500 designs" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000116_itoec.2018.8740499-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000116_itoec.2018.8740499-Figure4-1.png", + "caption": "Fig. 4.Binocular recovery space depth of field schematic", + "texts": [ + "Binocular camera back to the principle of depth of field map Algorithms for binocular camera calibration and optimized feature matching are implemented. Calibration is the operation that the binocular camera must perform before use. The accuracy of the binocular camera calibration directly affects the accuracy of the matching [12]. For binocular reconstruction, the depth map is first obtained by using the semi-global matching algorithm, and then the spatial depth image is restored by the principle of triangulation [13]. The principle is shown in Fig. 4. As shown in Figure 4, Q is a point in the line of sight of the binocular camera. P' and P'' are the two optical centers of the binocular camera respectively. The imaging plane of the camera is rotated and placed in front of the lens. Q is in two cameras. The imaging points on the photoreceptor are Q' and Q'' respectively, f is the camera focal length, Y is the center distance of the lens, Z is the depth of field, and the imaging point distance is D, then the formula of D is: D Y XL XR (1) According to the similarity principle of the triangle, the formula for the depth of field Z can be obtained as: Y XL XR Y Z f Z (2) The formula for obtaining Z from equation (2) is: Z Yf XL XR (3) Therefore, only the parallax (XL-XR) is measured, and the depth of field is reflected in the binocular camera" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002642_j.mechmachtheory.2017.02.003-Figure11-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002642_j.mechmachtheory.2017.02.003-Figure11-1.png", + "caption": "Fig. 11. Representation of the locomotion unit trajectory on nominal stairs with different slopes.", + "texts": [ + " The step geometry that simplifies the process of cam design (i.e. the nominal stair) is the one that generates the simplest trajectory for point P. This condition is obtained when the locomotion unit moves only with consecutive rotations around the blocked wheel of the locomotion unit. It means that after a 120 \u00b0 rotation of planet carrier the front wheel comes in contact with both riser and tread of the following step. In this condition, the trajectory of the locomotion unit center is a sequence of circular arcs as represented in Fig. 11 independently from the stair slope. This condition occurs when e is equal to the distance between two wheels of the locomotion unit as stated in Eq. (2) . e = e n = l L \u221a 3 \u223c= 277 . 1 mm with l L = 160 mm (2) The value of the length of the locomotion unit arm (l L ) has been obtained through the scheme of Fig. 12 by considering the configuration in which the highest step is climbed with a safety margin in the upper wheel contact. By applying geometrical relations on the scheme of Fig. 12 , Eq. (3) can be derived" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002773_12.893614-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002773_12.893614-Figure2-1.png", + "caption": "Figure 2. Schematic representation of the CERES instrument.", + "texts": [ + " 1 Corresponding author: Kelly K. Teague E-mail: kelly.k.teague@nasa.gov, Telephone: +1-757-864-9623 Earth Observing Systems XVI, edited by James J. Butler, Xiaoxiong Xiong, Xingfa Gu, Proc. of SPIE Vol. 8153, 81531S \u00b7 \u00a9 2011 SPIE \u00b7 CCC code: 0277-786X/11/$18 \u00b7 doi: 10.1117/12.893614 Proc. of SPIE Vol. 8153 81531S-1 Downloaded From: http://proceedings.spiedigitallibrary.org/ on 06/28/2016 Terms of Use: http://spiedigitallibrary.org/ss/TermsOfUse.aspx Figure 1 shows a schematic view of the CERES instrument and Figure 2 is the corresponding exploded view. To determine the Earth\u2019s net radiation flux, CERES employs three detector channels. A shortwave channel is sensitive to reflected sunlight, a total channel measures total radiation, from which the Earth-emitted radiance is computed by subtracting the reflected solar radiance, and a 8 \u2013 12 micron channel for measuring radiance in the longwave window.5 Each channel has a telescope which focuses radiation onto a thermistor-bolometer detector. CERES science data is acquired by continuously scanning its radiometers over the Earth surface" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002325_eucap.2017.7928215-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002325_eucap.2017.7928215-Figure4-1.png", + "caption": "Fig. 4. Radiation pattern of port 1. (a) 3D pattern at 2.45 GHz, (b) 3D pattern at 3.5 GHz.", + "texts": [ + " 3 shows the S-parameters of the prototype. In general, the simulated and measured results show a reasonable agreement. A little discrepancy is observed for the reflection at port 2 (inner patch), which is probably mainly due to the fabrication tolerances and the unavoidable slight misalignment of the substrates, due to the manual assembly process. For both feeding ports, the return losses show the capability of dual-band operation around 2.4 GHz and 3.5 GHz. The far field patterns of the antenna are given in Fig. 4 for port 1 and in Fig. 5 for port 2. In the lower band, the pattern is a standard broadside pattern for port 1 and a standard omnidirectional pattern for port 2. The cross-polar components are below -20 dB. However, in the upper band, side lobes appear because the radius of the annular ring is larger than one wavelength at this frequency. The main performance of the antenna is summarized in TABLE I. The correlation coefficient between the two ports is extremely low, i.e. below 0.04 and 0.05 for the lower and upper bands in measurement, and below 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003331_9781119546924.ch7-Figure7.1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003331_9781119546924.ch7-Figure7.1-1.png", + "caption": "Figure 7.1 Schematic of a 3-phase synchronous machine.", + "texts": [ + " There are a number of excellent texts on synchronous machines [6, 93]. The discourse in this chapter follows the materials and organization in [94]. A synchronous machine consists of a rotor generating a constant-speed rotating magnetic field inducing alternating currents in the windings on the stator. The magnetic field on the rotor is generated by the field winding controlled by the field voltage. The schematic of a rotor with two poles and a stator with three stator windings, separated equally by 120\u2218, is shown in Figure 7.1, with the rotor turning in the counter-clockwise direction at an angular velocity of \ud835\udf14r . Power System Modeling, Computation, and Control, First Edition. Joe H. Chow and Juan J. Sanchez-Gasca. \u00a9 2020 John Wiley & Sons Ltd. Published 2020 by John Wiley & Sons Ltd. Companion website: www.wiley.com/go/chow/power-system-modeling 178 7 Steady-State Models N S Rotor a b\u2032 a\u2032 b c c\u2032 Stator Air gap Amortisseur bar d-axis q-axis r b-axis c-axis \u03c9 \u03b8 The direction of the rotor magnetic field defines the direct- or d-axis. The quadratureor q-axis is perpendicular to the d-axis and rotate in the direction of the rotor. The three windings of stator circuits are called phase a, phase b, and phase c. The axes of the phases are defined with respect to the d-axis, as indicated in Figure 7.1. The relative positions of the three stator windings are such that a 120\u2218 counterclockwise rotation after the d-axis is aligned with phase a, the d-axis will be aligned with phase b, and then another 120\u2218 counterclockwise rotation later, the d-axis will be aligned with phase c. Synchronous machines are mainly of two types depending on the construction of the rotor. Round-rotor machines have a solid iron rotor with a uniform air gap between the rotor and the stator (Figure 7.2). These machines normally have two poles, and are primarily intended for high-speed synchronous machines (3600 or 3000 rev/min) such as steam turbines", + " For a p-pole machine with the rotor rotating at 3600 \u00d7 2\u2215p rev/min, with p a multiple of 2, the relationship between the electrical and mechanical angles is \ud835\udf03 = p 2 \ud835\udf03m (7.1) In this definition, the electrical angle between two adjacent north and south poles will always be 180\u2218. With this definition of the electrical angle, the derivation of a mathematical model of a synchronous machine can be based on a two-pole configuration. Amortisseur bars or damper windings are normally added to the rotor surface to damp harmonic oscillations [93, 95]. They are typically embedded in the pole face, as shown in Figure 7.1. The bars are connected at the end by rings to form a short circuit. Their effect is captured in detailed synchronous machine models. Following the schematic of the synchronous machine in Figure 7.1, Figure 7.4 shows the dq-circuits on the rotor and the individual phase circuits on the stator. On the rotor, the field winding is aligned with the d-axis. The amortisseur circuits are included on both the d- and q-axes, and are assumed to be decoupled. To avoid clutter, only one d-axis amortisseur circuit and one q-axis amortisseur circuit are shown in Figure 7.4. If more than one amortisseur circuit is modeled, it may be denoted by k = 1, 2, ..., for as many circuits as needed (although in practice k is seldom greater than 2)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003283_870182-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003283_870182-Figure7-1.png", + "caption": "Figure 7", + "texts": [ + " A small incremental pressure loading of 5 psi was applied inside the pressure vessel to establish relative stress distribution and stiffness characterlstics tor the two designs. A linear elastic detormation assumptlon was used tor this analysls. By not1ng the relative response of the pressurized tank and header 1n each des 190, load transmitting boundary conditions between the two components wer\"8 iteratively determined. The analysis a 150 showed where the tank and the header were 1n contact and where they pulled away from each other. Refer to Figure 7 for further clarIfication of final model geometry and boundary conditions. from the 2 1/2-0 model results for each design. An Internal load of 5 psi pressure was applIed, and a I inear static analysis response was calculated. All models were generated wIth a coarse mesh. since def I ect 10ns (node d I sp I acements) were of Immed I ate lntere3t. Deflection calculations do not require as fIne a mesh for convergence. as S 1m 11 ar Iy accurate stress data do. The 3-D mode 15 became more coarse as the distance fran the mid-length section lncreased" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000143_isemc.1982.7567736-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000143_isemc.1982.7567736-Figure5-1.png", + "caption": "FIGURE 5. MAGNETIC FIELD SENSOR CONFIGURATION", + "texts": [ + " The metallic glass magnetostrictive alloys for which the highest sensitivity has been achieved are amorphous metal alloys produced in ribbons 20 to 50 pm thick and from 1 to 50 mm wide. Using an unbiased strip of metallic glass 7 cm long to which the optical fiber was bonded, a magnetic field detection o f 3.3 x ]0_7 Oe was demonstrated. Furthermore, the measurements, carried out to above 1 0 1 Oe, were linear over the entire range. A sensor configuration suggested for a magnetic probe is shown in Figure 5. Both the sensor and the reference fiber are sandwiched between two strips of properly annealed magneto strictive amorphous metal. The maximum value of dT is obtained by annealing the sample at 374\u00b0C with a magnetic field value of 5000 Oe applied at right angles to the axis of the strip. This socalled \u201c magnetic anneal\u201d increases the value o f dT by as much as an order o f magnitude compared to a conventionally (without the magnetic field) annealed sample. Thus, significantly increased sensitivity is achievable", + " The reference and sensor leads, extending from the Mach Zehnder interferometer to the sensing element and back to the interferometer, must be potted together into a plastic tube. This arrangement allows the leads to follow nearly identical paths and thus to be subjected to the same noise field (both thermal and acoustic). The plastic tube, potting material, and fiber jacket also serve as a low pass filter of thermal fluctuations. The Fourier components of the thermal fluctuations above 1 Hz or less are rejected. The use o f phase-locked loop or heterodyne detection compensates for thermal drift. A magnetic sensor fabricated as shown in Figure 5 is expected to exhibit the desired sensitivity over the entire frequency range although laminar construction will probably be required for high-frequency operation. Electric Field Sensor Element The E-Field sensor element, shown in Figure 6, is sand wiched between two strips of the piezoelectric polymer, polyvinylideneflouride (PVDF). The sensor fiber is bonded in place, while the reference fiber is free. The reference and sensing leads would be treated as described for the H-Field sensor in order to minimize thermal and acoustic noise" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002314_12.902237-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002314_12.902237-Figure1-1.png", + "caption": "Figure 1: Design overview of the Isara 400 CMM; a) 3D design concept, without covers, b) photograph of the complete machine.", + "texts": [ + " The Isara 400 CMM is the latest development of IBS Precision Engineering for coordinate metrology of large, complex parts, featuring a traceable measuring uncertainty of 50 nm in 1D (2\u03c3) and 100 nm in 3D, both specifications are including probe and within the complete measuring volume of 400x400x100 mm. The Isara 400 CMM is capable of measuring complex surfaces such as aspheres, free-forms or integrated optics with nanometre accuracy in 3D. In addition, application areas include geometrical inspection of a wide range of industrial parts, similar to conventional CMMs, but with much higher accuracy. An overview of the complete machine is shown in figure 1. Optical Fabrication, Testing, and Metrology IV, edited by Angela Duparr\u00e9, Roland Geyl, Proc. of SPIE Vol. 8169 81690T \u00b7 \u00a9 2011 SPIE \u00b7 CCC code: 0277-786X/11/$18 \u00b7 doi: 10.1117/12.902237 Proc. of SPIE Vol. 8169 81690T-1 Downloaded From: http://proceedings.spiedigitallibrary.org/ on 07/03/2016 Terms of Use: http://spiedigitallibrary.org/ss/TermsOfUse.aspx Three plane mirror laser interferometers are applied as measuring systems for the machine axes. The interferometers each measure against the sides of a mirror table, on which the work piece is mounted" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001650_pcicon.2015.7435110-Figure15-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001650_pcicon.2015.7435110-Figure15-1.png", + "caption": "Fig. 15 Thermal Bow on one end of machine", + "texts": [ + " 1) Case A: Though it is rare, thermal changes inside the machine have been known to manifest themselves only on one end of the machine. This can happen when there is a thermally sensitive shaft with rigid welded spider arms that keep the shaft from bowing along the core length. The shaft is only able to bend externally to the spider arms. In some cases the shaft significantly bends only on one side. Not bending equally on both ends could indicate a different bending length outside the core area with a greater change on the longer end (Fig. 15). If the increase in vibration is on the non-drive end, then clearly the external sources are less likely a factor. If the high vibration is on the drive end, then the external source, e.g. coupling, shaft extension, alignment or driven equipment, could be eliminated by measuring the vibration hot uncoupled. If the higher vibration still exists then the external sources are eliminated. 2) Case B: Another thermal bow example is when the rotor gets hot and the vibration on each end of the motor increases but is 180 degrees out of phase" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002469_cdc.1990.203965-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002469_cdc.1990.203965-Figure4-1.png", + "caption": "Figure 4. Heavy Wrap Type Small Diameter Power Grasp", + "texts": [ + " Assuming the case of a balanced rod, the analysis is again reduced to a planar case with two virtual fingers contacting the cylinder at 5 points. Keeping the grasp orientation fixed with 0, = 0 (Fig. 2), the maximum value of cylinder weight W is calculated as its radius is varied from 1.732 in. to 0.577 in. and imposing limits on the applied joint torques. The corresponding grasp configuration varies from a large diameter power grasp (6' = 60\") to a small diameter power grasp (0 = 120\u00b0), shown in Fig. 4. The problem is again formulated using linear programming as defined by Eqs. (1)-(3). Moreover, friction constraints at the 5 contacts are also included in the formulation as - /iN 5 F 5 /IN (4) where ,U is the coefficient of friction and F is a (5 x 1) vector of tangential reactions at the contacts. The modified size of matrix A is (6 x 15) and that of matrix B is (15 x 1). The magnitude of the frictional reaction at each contact was adjusted by the linear programming algorithm to achieve the optimum value of W " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002450_iet-map.2012.0077-Figure9-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002450_iet-map.2012.0077-Figure9-1.png", + "caption": "Fig. 9 Photograph of the switchable bandpass filter for WiFi and UMTS transmit standards", + "texts": [ + "955 GHz with a passband bandwidth of 715 & The Institution of Engineering and Technology 2012 140 MHz, and a second state having a centre frequency of 2.440 GHz, with a passband bandwidth of 80 MHz. These states correspond to the UMTS and WiFi transmit standards respectively, the filter must also be able to handle maximum power levels of 16 and 21 dBm for the WiFi and UMTS states, respectively, without generation of third-order inter-modulation distortion. The filter was designed using a Rogers 1.524 mm thick substrate (1r \u00bc 3.55, d \u00bc 0.0021) and HPND-4028 Avago Technologies PIN diodes. The fabricated device is shown in Fig. 9. The layout including the DC bias lines was patterned on the substrate using standard photolithographic techniques. The Bias network consisted of a choke inductor to isolate DC bias lines and circuitry from the microwave circuit [10]. The current on each diode was limited to 10 mA by placing a 1 kV series resistor in the forward bias state; a voltage of 210 V was supplied in the reverse bias state. Lumped element models for the PIN diodes and choke inductors were calculated for both forward and reverse bias states" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002376_j.acme.2017.11.002-Figure10-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002376_j.acme.2017.11.002-Figure10-1.png", + "caption": "Fig. 10 \u2013 The control points for phase volume fractions analysis.", + "texts": [ + " The simulation conditions such as rolling schedule or cooling strategies were the dual stresses in strip for late cooling mode. same as in the rolling experiment. In the calculations, heat generation during plastic deformation and heat of phase transformations were taken into account. The analysis of residual stresses was carried out for the average cross-section along the length of the strip. The control points (P1, P2) in which the change in volume fractions over time was analyzed are shown in Fig. 10. The modeling results of volume fractions of bainite and pearlite in the coils after cooling process and in control points P1 and P2 during cooling are shown in Figs. 11 and 12. In turn, Figs. 13 and 14 illustrate the distribution of residual stresses across the width for both investigated steels. As was mentioned in Section 4, the experimental validation of the developed model of residual stresses was carried out for the steel S235. Late and early cooling mode was adopted for laminar cooling system" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003494_iet-map.2010.0218-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003494_iet-map.2010.0218-Figure1-1.png", + "caption": "Fig. 1 Adaptive surface patch meshing used for antenna modelling", + "texts": [ + " For the electromagnetic surface patch model used, the antenna geometry is adaptively divided into optimum numbers of trilateral and quadrilateral polygons, each polygon node being specified by its x, y and z co-ordinates, subject to the defined antenna parameters. These polygonal surfaces are then optimally subdivided into a set of rectangular and triangular surface patches, constrained to be small compared with the operating wavelength, and then the basis functions are adaptively generated over these patches using a designated algorithm as shown in Fig. 1. The design of coaxially-fed air-dielectric microstrip harmonic-rejecting patch antennas for 2.4 GHz was investigated, enforcing suppression of the first two harmonic frequencies, using a GA. The designs included patch antennas with a shorted wall, as first presented in [24] and an extended work on a new design with a folded wall is also presented in this paper. Simple coaxially-fed air-dielectric patch antennas with shorted and folded walls, mounted on an infinite ground IET Microw. Antennas Propag" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001395_ijvp.2016.075351-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001395_ijvp.2016.075351-Figure1-1.png", + "caption": "Figure 1 The SIGMA/SAMS tracked vehicle model (see online version for colours)", + "texts": [ + " He is currently a Professor at the University of Illinois at Chicago and is a Fellow of the American Society of Mechanical Engineers. Accurate prediction of the dynamic stresses in multibody systems (MBS) simulations is crucial in the design of many engineering systems. One system that is commonly modelled and tested because of its use in tough terrains and because of the high manufacturing and maintenance cost is the tracked vehicle. Examples of tracked vehicles include bulldozers, military battle tanks, and armoured personnel carriers such as the one shown in Figure 1. One of the most complex and difficult to model sections of these vehicles is the track chain, which consists of multiple track links interconnected by revolute joints. These joints can be modelled using kinematic constraints or bushing elements to eliminate or constrain the necessary degrees of freedom and allow for a rotation about a single axis. Because of the difficulties of developing flexible link chain tracked vehicle models, the stress analysis of such systems has been performed by obtaining forces from a rigid body analysis of the vehicle", + " j jl l T sU EA x EI x U\u03b5 \u03ba= + = \u2212 \u2202 \u2202\u00f1 \u00f1 Q e The axial strain at an arbitrary point on the Euler-Bernoulli beam element can then be defined using the gradient vector evaluated using the element assumed displacement filed as 1 111 ( 1) / 2.T x x\u03b5 = \u2212r r The evaluation of this axial strain using ANCF finite elements in a coupled rigid body/deformation analysis is straight forward and does not require the complexity required by the less accurate post-processing stress analysis that ignores the effect of the elastic deformations on the rigid body motion. The tracked vehicle model shown in Figure 1 will be used in the following section to compare between post-processing decoupled rigid body/deformation stress analysis based on the simplified FFR equations and the ANCF fully coupled stress analysis that takes into account the effect of the deformation on the rigid body motion. The tracked vehicle modelled is an armoured personnel carrier consisting of a chassis, idler, sprocket, 5 road-wheels, and 64 track links on each track side (right and left). Figure 6 further shows the contact engaged between track links and other components such as the sprocket, road wheels, and ground" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002988_0094-5765(76)90054-0-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002988_0094-5765(76)90054-0-Figure1-1.png", + "caption": "Fig. 1. Principle of the accelerometer (1 axis)--(l) towards the position detecting electrode; (2) towards the action electrodes; (3) position detector; (4) correcting circuit; (5) amplifier; (6) telemetry transmitter.", + "texts": [], + "surrounding_texts": [ + "In view of its measuring range, the Cactus accelerometer can be tested in normal operat ion only by weightlessness simulation; the testing facilities available at O N E R A (42-m free fall in vacuum) authorize only a test duration lower than 3 sec; so it is not possible, on the ground, to proceed to a low level direct calibration and to a precise verification of the instrument systematic error. The calibration coefficient is determined on the ground by calculation, f rom measurements of electric capaci tances and of distance. The instrument systematic error, which results f rom the effect on the proof mass of the spurious forces (magnetic effect, electric charges, thermal effects, d i ssymmetry of gravitational at traction forces); it has been est imated by theoretical means. Apart f rom the verification of the instrument correct operat ion in space conditions, qualification aims at checking the calibration coefficients and the measurement of systematic errors. To this end, the satellite was equipped with three mobile masses (1 kg--free course +-50 ram) allowing the displacement of the satellite inertial centre along the three reference axes and with two fly-wheels that can provide it with a rotation velocity of about 5\u00b0/sec. With these remotely controlled devices and instruments for measuring the satellite attitude (Sun trackers and magnetometer) it is possible to know precisely the acceleration due to the rotational inertial forces applied to the accelerometer in a given configuration. The instrument systematic error is measured after a centering ensured within a few microns and a reduction of the rotation velocity to a value small enough for the inertial effects to be negligible." + ] + }, + { + "image_filename": "designv6_24_0003921_el:19900445-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003921_el:19900445-Figure1-1.png", + "caption": "Fig. 1 Cross-section of GaAs FET showing regions contributing to R, and R,", + "texts": [ + " The question is, are such negative values erroneous, as they are mostly taken to be, or do they in fact have a physical basis? The purpose of this letter is to give a brief account of one of several mechanisms operating in GaAs FETs and HEMTs that may require R, to be negative, and to illustrate the basic principle that R, cannot be treated as a simple biasindependent resistance. Bias variation of R,: The equivalent circuit is an abstraction that is intended to represent the physical processes occurring in the real device in all respects. Fig. 1 shows an FET in cross-section in which the region between the extreme righthand end of the channel depletion and the external drain contact is represented by the circuit element R,. When determining an empirical value for R,, DC or \u2018cold FET\u2019 measurements are often used. R, is subsequently treated as bias-independent, i.e. the same fixed value is used in equivalent circuit fits regardless of the bias applied. In the real physical device, however, when either the gate bias or the drain bias is changed, the space-charge layer extension into the gate-drain region is also changed", + " Fibre-optic polarisation beam splitters/combiners, have been intensively studied and presented in several different schemes.\u201d\u2019 The polarisation beam extinction ratio in these schemes has been reported to be around 20dB.1-3.5.6.9 In Reference 7, a polarisation extinction ratio (PER) as high as 25 dB is reported.\u2019 With the recent development of polished plasmon FPBS devices even better performance has been obtained.\u201d,\u201d It is found that simple FPBS devices with very good performance can be constructed with the twin-elliptic core fibre. The FPBS is shown in Fig. 1. For an elliptic core single-mode fibre, the deformation from circular symmetry 682 introduces form birefringence. If we define x and y as the transverse axes of the fibre cores, as shown in Fig. 1, then x-polarised and y-polarised modes have different coupling coefficients. The birefringence in each single-core, together with the asymmetry caused by the noncircular cross-section, causes the modal patterns bj (j = x, y), the propagation constants and the coupling coefficients C j of the two polarisations to differ.\u201d Our twin-elliptic-core fibre has an relative index difference A - 0.0035, core separation d - 9pm, core short radius p, - 1.4pm and long radius pL - 3.5pm. The coupling coefficients C , and Cy are measured to be about 1", + " Another stringent requirement is that two cores should be perfectly identical. This involves the design and preparation of twin-core preform. The mismatch between the two cores will prevent energy from one core from being completely coupled to the other. The residual energy presumably limits the obtainable PER. Fig. 2 shows the experimental measurement of the outputs at core 1 and core 2 against the azimuth angle of a polariser put at the output end of the fibre. The zero degree position of the output polariser is adjusted to be consistent with the y-axis in Fig. 1. From this result we can see that the x-polarised mode emerges in the cross core and the ypolarised mode in the straight-through core. The polarisation beam extinction ratios from the measurement are 26.4dB and 26.2 dB for straight-through and cross core outputs, respectively. The asymmetry in the measured curves (the flattened peak in the straight-through output and the sharpened peak in the cross output) is because of the polarisation selective transmission of the polariser and the angle prism reflector used in the measurement" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000715_062035-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000715_062035-Figure4-1.png", + "caption": "Figure 4(a). Heat flux contours for Figure 4(b). Temperature contours for structural steel structural steel", + "texts": [], + "surrounding_texts": [ + "The boundary conditions followed in the present work is mentioned below and these points are considered in the design and analysis of the proposed work. (i) The model was meshed using tetrahedron method. (ii) The element was meshed with coarse type and medium smoothing. (iii) Number of nodes = 3986 (iv) Number of Elements = 2003 (v) After this, the teeth section of the rack was refined further to provide a more accurate simulation model and to give more accurate results. (vi) Number of nodes (after refinement) = 16380 (vii) Number of elements (after refinement) = 8266 To set-up the thermal-electric simulation, the number of time steps over which the calculation is done is take as 10 (not the default 1) time steps with 1 second per step and an update interval of 2.5 seconds kept as default. This is done so as to provide ample number of iterations and at the same time, to provide a 5 IOP Conf. Series: Materials Science and Engineering 263 (2017) 062035 doi:10.1088/1757-899X/263/6/062035 congruence point in a relatively quicker time [10-12]. Also, the radiosity controls that measure the relation between the radiation convergence and the surface parameters has been kept on with a convergence value of 0.0001 W/m2. Smooth contours have been set to provide details about the variations in the input and output results. The meshed model of the designed structure is shown Figure 1. The solver used here was of gauss-seidel iterative type. These are the FEM equations needed and used by the solver while analyzing the data given and these are the main equations used for determining the results. The resources for this data were taken from the help section in the project window main screen. Biot-Savart Law for finding the magnetic vector potential Gauss Law for determining flux density Faraday\u2019s law for calculating the electric field intensity. Ampere\u2019s Law for current density (not compulsory). Equation for determining Joule Heat Generated. Taking into account, the real-time conditions present, an input voltage of 525 volts was applied in increasing steps throughout each of the ten time steps with phase voltage set to zero. To improve the accuracy of the readings, the voltage was applied across the 28 teeth only, but the output was generalized throughout the entire surface of the rack. Also as part of the thermal input, the room temperature was maintained at 30 \u00b0C and the gear was initially placed at the room temperature. As the voltage was applied, the rack was simultaneously cooled by the surrounding air and the coolant. For this, the convection factor was considered between the outer surface and water. Also, the heat flow was assumed to flow through the rack body and between the rack and the surrounding medium and so the total heat flux was evaluated as part of the results. 3. Results and Discussion 6 IOP Conf. Series: Materials Science and Engineering 263 (2017) 062035 doi:10.1088/1757-899X/263/6/062035 In the figure 2(a) given above the step increase of the input voltage provided to heat the rack is given in a graphical format to a maximum value of 525 V. Figure 2(b) shows the variation of the Convection Coefficient of quench water with respect to the temperature. Maximum Value Over Time Electric Voltage Joule Heat Total Electric Field Intensity Minimum Value Over Time Minimum 100. V 8.0742e-015 W/m\u00b3 1.0846e-011 V/m Maximum 525. V 2.3646e-013 W/m\u00b3 7.8596e-011 V/m Maximum Value Over Time Minimum 100. V 0.56343 W/m\u00b3 1.7167e-004 V/m Maximum 525. V 15.531 W/m\u00b3 9.013e-004 V/m 7 IOP Conf. Series: Materials Science and Engineering 263 (2017) 062035 doi:10.1088/1757-899X/263/6/062035 Minimum Value Over Time Minimum 100. V 4.4055e-015 W/m\u00b3 1.0291e-011 V/m Maximum 525. V 1.1376e-013 W/m\u00b3 8.415e-011 V/m Maximum Value Over Time Temperature Total Heat Flux Minimum Value Over Time Minimum 265.38 \u00b0C 8.2511 W/m\u00b2 Maximum 589.72 \u00b0C 1266.3 W/m\u00b2 Maximum Value Over Time Minimum 421.04 \u00b0C 4.9745e+005 W/m\u00b2 8 IOP Conf. Series: Materials Science and Engineering 263 (2017) 062035 doi:10.1088/1757-899X/263/6/062035 Maximum 611.41 \u00b0C 9.3048e+005 W/m\u00b2 The tables given above 5 to 7, denotes the maximum and minimum value of the output parameters for all of the materials based on the iterations conducted for all the time steps. These values are taken as comparison against the standard values of EN8C. As It is noticeable, for the given input voltage the Joule heat value is the maximum for Cast Iron > EN8C > Structural Steel. This means that maximum heat is produced in Cast Iron for the same voltage. However, the electric field intensity is the maximum for Structural Steel > EN8C > Cast Iron. For Heat Flux Structural Steel > EN8C > Cast Iron. This implies that the rate of heat energy transfer is maximum in Structural Steel and minimum in Cast Iron. 9 IOP Conf. Series: Materials Science and Engineering 263 (2017) 062035 doi:10.1088/1757-899X/263/6/062035 Figure 3(a) shown above, is a close up view of the heat flux variation in the teeth section in the rack. As seen, it is not even around the teeth gaps and varies individually. The figures given above depict the contours obtained for the parameters Heat Flux and Temperature for all of the simulated materials. These are based on the iterative values whose boundary limits are mentioned in the tables 5-7. As seen from figures 3(b), 4(a), 5(a) we notice that the Heat Flux is maximum in the teeth region of the rack gear and starts decreasing as we move across the length of the rod. This is due to the dispersion by conduction and convection to the quench water. The same can be derived for the Temperature contours in 3(c), 4(b), 5(b). Here the temperature drop is noticeable right from the edge of the teeth section throughout the length of the rack. 1234567890 IOP Conf. Series: Materials Science and Engineering 263 (2017) 062035 doi:10.1088/1757-899X/263/6/062035 The Figure 6(a) provides the graphical representation of Joule Heat variation for 525 V. It is deduced that up to the 5th time step the values are almost similar but after this, the curve for Cast Iron rises almost exponentially when compared to the other materials. This is shown as the error percentage is between 2% and 7%. To test if this variation was constant, the input voltage was raised 615 V from the standard 525 V as shown in Figure 6(b). As expected, the values of Joule Heat for Cast Iron showed a similar rise after the 1234567890 IOP Conf. Series: Materials Science and Engineering 263 (2017) 062035 doi:10.1088/1757-899X/263/6/062035 4th time step while the values for the other materials remained constant with a small error percentage throughout the range of the time steps. This led to the conclusion that based purely on the results of this analysis, Structural Steel might be used as a substitute for EN8C. From the graph in Figure 6(c), we can determine that as the input voltage is lowered, the values of Cast Iron and Structural Steel obtain a similar pattern of results while those of EN8C show a considerable difference of almost around 2-4 W/m3. This voltage however does not provide a suitable threshold for the molecular structure to change to a martensitic structure and moreover shows us that the substitute materials used do not follow the same trends as the base material. From Figure 7(a) given above, we can determine that the rate of temperature drop is slightly more for Structural Steel = Cast Iron > EN8C. Upon comparison of the rate of change of heat fluxin Figure 7(b), it was found out that all the materials displayed a rise in the values of heat flux until the first-time step before falling steeply. Here also, the values for Cast Iron and Structural Steel were almost constant but the least heat flux drop was in EN8C. 1234567890 IOP Conf. Series: Materials Science and Engineering 263 (2017) 062035 doi:10.1088/1757-899X/263/6/062035 With slight variations to the voltage and electric field, the output values can be different for the materials. To further observe the changes in this process, the variables constituting the non-linear analysis pattern in the analysis section must be changed to include other analyses such as impedance analysis with an increasing accuracy of the finite element analysis itself. It was also found out that with a decrease in the input voltage, the case depth of the material decreased from the front (teeth) side and remained almost constant from the back side. This means that any change in the voltage should be undertaken only upon 1234567890 IOP Conf. Series: Materials Science and Engineering 263 (2017) 062035 doi:10.1088/1757-899X/263/6/062035 the specifications for the surface hardness and case depth being explicitly mentioned and tested after the process is completed. This matches with the inference made by Castallenos [11] \u201cThe overall efficiency of these systems is about 94% (near the resonance). The losses in the rectification stage can be assumed as 1% of the total efficiency. While remaining 5% may be related to inverter stage. In extreme cases, losses can be associated with an additional 5%\u201d. Furlani [12] has also verified these claims \u201cmaximum temperature values achieved in some nodes along a radius of the specimen and their correlation with the corresponding experimental final structures. In spite of the high heating rate at the surface of the specimen, the very high temperatures reached (about 1250 \u00b0C) in this zone are able to completely transform the external layer in austenite. The subsequent cooling phase induces in the material a complete martensitic micro-structure At intermediate zones, the maximum temperatures reached are in 600-700 \u00b0C interval, so that the transformation from perlite to austenite is almost complete (as well as the consequent austenite to martensite transformation), while the ferritic phase cannot transform itself into austenite.\u201d 4.Conclusions With increasing differences in the shape of the surface that has to be hardened due to the wide variety of gears used in the market today, a multiple frequency pattern of induction hardening would be considered more productive as the output time can be reduced thus pushing out more components per batch. This would mean the incorporation of more number of coils in the heating process with each coil being supplied a set voltage in a different frequency and used to heat a different section of the component. This would enable the different sections of the component to come out with different hardness values based on the demand of the customer. Also, due to the constraints of having properties similar to the standard material, only two alternative materials have been analyzed through this simulation. However, any other material with similar composition can be included in this analysis. To determine the viability of the substitute to the standard material, further testing has to be conducted and various other parameters have to be verified to put this substitute into large scale production. Some advances in IH process as mentioned by O. Lucia [13] such as \u201cOne of the issues for the future of IH is the load adaptive capabilities and some solutions have been proposed. An adaptive simmering control of the temperature for a domestic induction cooker is required. Parameters are updated online, depending on the estimates provided by a multiple-model reset observer (MMReO). This observer consists of a reinitialized reset observer and of multiple fixed identification models.\u201d" + ] + }, + { + "image_filename": "designv6_24_0001669_icmech.2019.8722911-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001669_icmech.2019.8722911-Figure1-1.png", + "caption": "Fig. 1. Schematic views of example systems. Left: Two coupled masses. Right: Planar four-joint manipulator where the end-effector is actuated w.r.t. the reference frame. Note that for display-reasons a) in the depicted configuration joint angles \u03b82 and \u03b84 take negative values and b) link lengths l1, . . . , l4 are not shown.", + "texts": [ + " However, in some cases the conclusion can be drawn without any calculation: Corollary 8: For systems with m \u2265 n\u2212 1 and for systems with B(\u03b8) = const there always exists a separated coordinate representation of the form (21). Proof: In the case of m \u2265 n \u2212 1 = k the wedge product on the left hand side of (19) is at least a (n + 1)-form in n variables and thus vanishes. Independently of m it is clear that d\u03c9i \u2261 0 for all i = 1, . . . ,m if B does not depend on \u03b8 (which includes the case of linear systems). Hence, in that case (19) also holds trivially and thus a separated coordinate representation does exist. 2 For illustration we consider one the simple mechanical systems, depicted in Fig. 1 (left). Its equations of motion read( m1+m2 m2 m2 m2 ) \u03b8\u0308 + ( 0 c\u00b7\u03b82 ) \u2212 ( 0 d\u00b7\u03b8\u03072 ) = ( 1 1 ) \u03c4. (33) Due to B = const Corollary 8 applies and the existence of a separated coordinate representation is granted. To calculate the transformation a good starting point is (31). Because here m = 1 the regular m\u00d7m matrix S(\u03b8) needs not to be considered and we have \u03a6\u2032(\u03b8) = ( \u2217 \u22171 1 ) , which is easy to complete to get a regular Jacobian of the transformation: e. g. \u03a6\u2032(\u03b8) = ( 1 0 1 1 ) . From this we obtain the new (separated) choice of coordinates \u03be = ( pq ) = ( \u03b81 \u03b81+\u03b82 ) . Underactuated manipulators are an important subclass of underactuated systems, e. g. see [4] and the references therein. Typically, they are composed of active and passive joints, which directly lead to a separated coordinate representation if relative joint angles are used. Here, we consider a planar manipulator with four passive joints, where the two external forces \u03c4t and \u03c4n are applied along the last joint in tangential and normal directions, respectively, see Fig. 1 (right). It is assumed that these forces act w.r.t. the reference frame. From the equations of motion (3) for this system only the right hand side is necessary to evaluate both Theorem 6 and Corollary 7. For the considered system we have B(\u03b8) = ( l1s234+l2s34+l3s4 l1c234+l2c34+l3c4+l4 l2s34+l3s4 l2c34+l3c4+l4 l3s4 l3c4+l4 0 l4 ) , (34) with sijk := sin(\u03b8i + \u03b8j + \u03b8k), cijk := cos(\u03b8i + \u03b8j + \u03b8k) and analogous abbreviations for the other c- and s-terms. For illustration, we apply both the 1-form and the vector field version of the presented result" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002388_2015-01-0090-Figure9-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002388_2015-01-0090-Figure9-1.png", + "caption": "Figure 9. As viewed from side of the truck, the illustration shows start and", + "texts": [ + " Figure 8 shows strain responses obtained from the test of front left wheel on the ramp. The truck was driven up several times before the front left wheel could reach the highest position of a ramp, which was 300 mm from level road. The considered responses were between 280s-300s, in which the truck was in static condition and engine was not operated. In this case, a front left spring experienced its maximum bump travel observed by a zero gap between bump stopper, attached on the axle, and stopping plate, welded to the side rail lower flange. In Figure 9, vertical distances between the wheel centre and level road of the front left wheel at start position (yi) and stop position (yf) were measured. truck (Location 3, lateral direction), in case of Front-Left wheel on ramp stop positions of the acting wheel. The distances from wheel centre to level road at start position (yi) and stop position (yf) were measured. In the simulation, vertical displacement of yf\u2212 yi (Figure 9) was assigned to the left end of the front axle. Figure 10 is a contour plot of equivalent strains on the truck frame predicted by the simulation. For convenience, lateral and longitudinal strains were defined to comprehend our discussion later on. The lateral strain was normal strain measured in a direction along the length of cross members near their ends as showed in Figure 11. The longitudinal strain is normal strain located on the parallel flanges of side rails (U-shaped channel), Figure 12" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003657_ccece.2013.6567714-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003657_ccece.2013.6567714-Figure2-1.png", + "caption": "Fig. 2. Nonlinear Bicycle Model where \u03b8 = 0, \u03c8 = 0, \u03b4 = 0, and point b corresponds to the origin. Here Gx and Gz represent global Cartesian axes.", + "texts": [ + " Additionally, validity of the linearized version of this model has been experimentally verified for some velocities by Kooijman [9]. Key assumptions of the model are that the bicycle consists of two wheels, a fork, and a rigid rider attached to the rear frame. It is also assumed that the travelling surface is flat, that the wheels travel without slipping, and there is no friction within the bicycle. The model in [5] is derived using Lagrangian mechanics, and lends itself well to control due to the selection of generalized coordinates; the important ones are shown in Figure 2. From a control perspective, the most important variables are the lean angle (\u03c8) and the steering angle (\u03b4); the angle (\u03b3b), not shown in the figure, also plays an important role in path following. The two geometric constants \u00b5 and w are also significant. The full model is quite complicated: there are eight nonlinear differential equations and five constraint equations. The solution of these equations requires solving differential algebraic equations. Additionally, the terms in the equations are very complicated; including the full model here would consume dozens of pages", + " We first introduce integrators (with state denoted z= [z\u03c8 z\u03b4] T ) to guarantee perfect asymptotic tracking (where r\u03c8 and r\u03b4 are the reference values of lean and torque), and compute the control gains (denoted F1 and F2 in Figure 1) by minimizing J = \u222b \u221e 0 ( xT Qx+uT Ru ) dt where x= [xT zT ]T and where the LQR matrices are chosen to be Q = 1 0 0 0 0 0 0 0 0 0 0 0 0 0 1 0 0 0 0 0 0 0 0 0 0 0 0 0 100 0 0 0 0 0 0 100 , R = [ 10\u22125 0 0 10\u22124 ] . (4) The resulting gains correspond to a high-performance controller and lead to large values of T\u03c8; if the values are too large in practice, the (1,1) entry in R can be increased and new gains computed. The objective of the outer-loop controller in Figure 1 is to set the reference signals r\u03c8 and r\u03b4 such that the the rearwheel contact point b of the bicycle (see Figure 2) tracks a given path. To achieve this for paths other than straight lines, it is necessary to relate the steering angle to the yaw rate of the bicycle. For this we use the linear approximation derived by Meijaard [4] which is \u03b3\u0307b = v\u03b4+a\u03b4\u0307 w cos\u03bb, (5) where a is the distance the front wheel contact point trails the point where the steering axis intersects the ground plane. When treating \u03b3\u0307b as the input and \u03b4 as the output the transfer function is \u03b4(s) \u03b3\u0307b(s) = w (v+as)cos\u03bb . (6) We note that the value of a is typically small and when tracking a circle or straight line \u03b3\u0307b is constant in steady state" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001960_tec.2016.2609338-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001960_tec.2016.2609338-Figure1-1.png", + "caption": "Fig. 1. Cross section of MGs. (a) Proposed CPM-MG. (b) Conventional SPM-MG.", + "texts": [ + " In Section III, the proposed MG is analyzed with emphasis on air-gap field distribution and torque-angle characteristic. Some critical parameters of the proposed CPM-MG are optimized and discussed in Section IV. Then three MGs with different rotor structures and magnet materials are designed and compared, with emphasis on the end effect, losses and cost effectiveness. Finally, test results are given in Section V, and conclusions are then drawn in Section VI. The configuration of the proposed coaxial MG is shown in Fig. 1(a). Likewise, the conventional coaxial MG is shown in Fig. 1(b), which is taken as a benchmark for comparison. Important geometric parameters of the MGs are labeled in both figures. As can be seen, both MGs consist of three components, viz., the outer rotor, the magnetic modulation ring, and the inner rotor. In many applications, the magnetic modulation ring is stationary, the rotor with more poles (i.e., the outer rotor in the two figures) rotates at low speed, and the rotor with fewer poles (i.e., the inner rotor in the figures) spins at high speed. A Coaxial Magnetic Gear with Consequent-Pole Rotors Jian-Xin Shen, Senior Member, IEEE, Hua-Yang Li, He Hao, and Meng-Jia Jin M 0885-8969 (c) 2016 IEEE", + " To guarantee the maximum torque density of the optimum MG, critical parameters including the outer rotor yoke thickness, and the geometries of the outer rotor magnets, pole-pieces on modulation ring and inner rotor magnets are all optimized both sequentially and globally. The influence of some key design parameters is discussed below. A. Geometry of Ferromagnetic Pole-Pieces on Modulation Ring Ferromagnetic pole-pieces on the modulation ring are of great significance in the performance optimization, and are scaled by the width ratio s and the thickness hs which have been illustrated in Fig. 1. Take the pull-out torque of outer rotor as the main concern, Fig. 9 shows the variation of toque with the two dimension parameters. Note, in the analysis, when adjusting hs, the outer rotor dimensions and the two air-gap thickness are unchanged. Therefore, for example, when hs increases, the outer diameter of the inner rotor reduces, and then the inner rotor magnets thickness hmi slightly increases so as to keep the magnets volume Vi and pole-arc coefficient i unchanged. The pole-pieces and adjacent air space provide specific magnetic reluctance distribution, thus modulating the field produced by both rotors" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001223_1.c032150-Figure12-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001223_1.c032150-Figure12-1.png", + "caption": "Fig. 12 Front view showing aerostructural deflections for M1 condition (left) and C1 condition (right).", + "texts": [ + " The TOGWoptimum has a stiffer structure, and the raked wing tip, because of the lower lift-curve slope, is able to produce the beneficial inboard load shifting for the maneuver conditions. In the fuel-burn case, however, even though the wing is heavier, the increased span leads to larger deflections and amore consistent gradual twisting near the tip. A similar aeroelastic effect of reducing the tip load is achieved naturally without raking the tip. A comparison of the deflections of the initial and optimized designs as well as the jig shape is given in Fig. 12. An examination of the chordwise Cp distributions and airfoil shapes can explain the aerostructural tradeoffs that the optimizer made between the five cruise-design points and the two maneuver conditions. The cross-sectional data are extracted from the 3-D geometry using a cut plane orientated with the x-z plane. Figure 13 shows the airfoil shape and Cp distribution for the initial and optimized designs for each operating condition at the 66% semispan location. Cruise conditions C1, C4, and C5 are all at the design cruise Mach number of 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003197_iecon.2016.7793899-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003197_iecon.2016.7793899-Figure2-1.png", + "caption": "Fig. 2. Stator current vector in PMS motors.", + "texts": [ + " In order to implement PMS motor current regulation, stator current reference should be determined based on the PMS motor mode of operation. PMS motor can be controlled based on different modes including maximum torque per ampere (MTPA), unity power factor, loss minimization, etc. After selecting the mode of operation, the reference of stator current vector can be calculated in the rotating rotor frame. The MTPA control is used here. The developed torque by PMS motor can be written as below in which si and are magnitude and angle of stator current in the rotor reference frame as shown in Fig. 2 [13]. 23 . 2 2 m s L sT p i sin i sin , (13) 2 q d L L L . (14) According to the above equation, PMS motors can produce a specific torque with different stator current magnitudes and angles. The optimal values under MTPA control can be found by solving: 0 T . (15) Substituting the torque equation of (13) into the above condition yields: cos 2 cos 0m s d q i L L . (16) Solving the above equation results in: III. PROPOSED CONTROL METHOD Assuming that the neutral point of stator phase windings is not grounded, the sum of the three machine currents is zero" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002691_2017-01-1767-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002691_2017-01-1767-Figure5-1.png", + "caption": "Fig. 5. Time-varying bearing stiffness (Ref. [19]): (a) Stiffness variation, (b) Bearing roller orbit positions", + "texts": [ + " Therefore, the load displacement relation of the combined shaft-bearing system can be written in matrix form as: (21) The 5 by 5 matrix [\u0394R] stands for the displacement at reference point in five DOFs except the torsional displacement around rotation axis. The 5 DOF matrix [\u0394S] stands for the displacement at other nodes. The external force expressed as [FR] is acting on the reference point. For calculation simplicity, the force on reference point is set to be unit load. Therefore, the equivalent time-varying lumped stiffness matrix can be written as: (22) The stiffness is calculated at each mesh frequency and specific roller orbit position. An example of x-direction translational stiffness of lumped support on pinion is illustrated in Fig. 5. The stiffness value in Fig. 5 (a) is dependent on different position of rolling elements as shown in Fig. 5 (b). To investigate the relationship between vibration mode and nonlinear effect on dynamic response, linear geared rotor system is also used in this study by ignoring gear backlash and bearing radial clearance. Based on modal superposition method and by solving eigenvalue problem, the dynamic compliance matrix of geared rotor system can be expressed as: (23) where \u03c9r is natural frequency, \u03a6r is the corresponding frequency mode and \u03b6r is the modal damping ratio written as: (24) The linear dynamic mesh force along effective LOA can be formulated as: (25) In current study, 80 Nm is applied at driving end of the model" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000327_s1068366618050112-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000327_s1068366618050112-Figure1-1.png", + "caption": "Fig. 1. Design diagram of contact problem for hub\u2013plunger frictional pair.", + "texts": [], + "surrounding_texts": [ + "The effectiveness of the operation of many types of frictional pairs largely depends on temperature. During a long period of operation the plunger with a large force rubs against the hub surface. The surface layer of the hub\u2019s metal is intensively heated. Because of the large heat conductivity, the heated surface layer of the hub is very quickly cooled and is hardened. Such a sharp hardening of a thin layer promotes the appearance of cracks on the hub. It is shown in [1\u20133] that at the spots of that actually touch a powerful heating occurs in the thin surface layers. This causes the appearance of regions of crack formation. In the case of powerful heating under the action of a temperature spike, the cracks originate in the near-surface layer. The destruction is caused by the heat production during friction. Thus, it can be assumed that each material has a limit (allowable) temperature , exceeding which leads to the formation of burns and regions of microcracks in the material of frictional pairs. Consequently, the magnitude of the maximum temperature reached in the material can be considered the cause of the thermal destruction of the materials of the frictional pair. Thermal failure in the development of friction nodes can be controlled by design and technological methods, in particular the geometry of the tribological conjunction, at the design stage. Solutions of the mechanics of problems by constructing such a geometry of the elements of the surface of the frictional pair, which would promote the reduction of the thermal stress and wear, with the exception of [4\u20136], are not known. The solution of this problem will make it possible to increase the working capacity of tribological conjunctions." + ] + }, + { + "image_filename": "designv6_24_0002995_s1759078716000775-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002995_s1759078716000775-Figure5-1.png", + "caption": "Fig. 5. Layout of: (a) the reference UWB antenna and (b) the proposed dual band-notched UWB antenna (unit, mm).", + "texts": [ + " contrast, the currents for the passband frequencies flow smoothly along the transmission line from input to output as shown in Figs 4(a), 4(c), and 4(e). I I I . D U A L B A N D - N O T C H E D U W B A N T E N N A In order to validate the design idea of the proposed resonator, two UWB antennas with/without dual notched bands are designed on the RT/Duorid 4350 substrate with 1r \u00bc 3.48 and the thickness of 0.508 mm. The proposed UWB antenna without dual notched bands as the reference antenna consists of a microstrip feed line and a staircase-like tapered semi-circular patch on a truncated ground plane as shown in Fig. 5(a). A rectangle slot is etched on the bottom ground to improve impedance matching of the antenna. Figure 5(b) displays the proposed dual band-notched UWB monopole antenna with the above-mentioned bandstop filtering element beside the feed line. The outer highimpedance line is modified into a meandering shape to achieve size reduction. The dimensions of the proposed dual band-notched ultra-wideband (uwb) monopole antenna 5 resonator were initially calculated by equation (1), and then HFSS was employed to optimize and determine the final dimensions of the whole structure based on the analysis as shown in Fig. 3. After simulated optimization, the dual bandnotched UWB antenna and the reference antenna as its counterpart are both fabricated whose parameters can be seen in Fig. 5. The photographs of the fabricated dual band-notched UWB antenna and its reference antenna are displayed in Fig. 6. Compared with most previous published work where multiple resonators are required to create multiple notched bands for an UWB antenna [6\u20138], the proposed design greatly simplify the design process since only one multi-mode resonator is needed. The simulated and measured reflection coefficients of the two UWB antennas are illustrated in Fig. 7, where they are all below 210 dB from 3" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000083_ir-03-2018-0042-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000083_ir-03-2018-0042-Figure8-1.png", + "caption": "Figure 8 Rear wheel mechanism description", + "texts": [ + " Sayed et al. Industrial Robot: the international journal of robotics research and application Volume 46 \u00b7 Number 6 \u00b7 2019 \u00b7 740\u2013762 The rear wheels are attached to the trunkmechanism by using a movable link which is actuated to assure that the center of gravity of the wearable vehicle stays within the support polygon. The range of motion of the rear wheels mechanism is restricted mechanically. It can move 30\u00b0 from the vertical line. These links are moved by using two electrical motors as shown in Figure 8. Motors are fixed in the trunk through four bolts. Two bearings are also fixed on the trunk. Lockers are used to avoid moving the motor shaft on the other side. The proposed design of the wearable vehicle system also allows for significantly more maneuverability on uneven terrain. One fashion, in which the suspension deals with off-camber surfaces, is through independentmovement of the wheel while moving on a surface, where the right shock compresses and the left relaxes. Then the wearable vehicle is able to level itself, offering more stability to the user" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002553_j.ymssp.2018.03.037-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002553_j.ymssp.2018.03.037-Figure2-1.png", + "caption": "Fig. 2. Cantilever beam model of meshing tooth.", + "texts": [ + " An empirical formula has been proposed to conveniently calculate the Hertzian contact stiffness expressed as follow [32]: kh \u00bc E0:9b0:8F0:1 1:275 \u00f02:5\u00de where E is Young\u2019s modulus; b is the tooth width. It is worth noting that similar formula was also adopted in Refs. [21,24], of which the Hertzian contact stiffness was assumed to be load-independent as: kh \u00bc pEb 4\u00f01 m2\u00de \u00f02:6\u00de where m is the Poisson\u2019s ratio. In this paper, Eq. (2.5) is adopted for its higher accuracy. Based on cantilever beam theory (see Fig. 2), ks, ka and kb can be expressed as follows [13,14]: 1 ks \u00bc Z d 0 1:2 cos2 am GAx dx \u00f02:7\u00de 1 ka \u00bc Z d 0 sin2 am EAx dx \u00f02:8\u00de 1 kb \u00bc Z d 0 \u00bdcosam\u00f0d x\u00de sinamh 2 EIx dx \u00f02:9\u00de where G is shear modulus; Ax is the area of cross section; Ix is the area moment of inertia of the cross section; For the calculation of fillet foundation stiffness (see Fig. 3), Sainsot et al. [15] proposed an analytical formula expressed as: 1 kf \u00bc cos2 am Eb L uf Sf 2 \u00feM uf Sf \u00fe P \u00f01\u00fe Q tan2 am\u00de ( ) \u00f02:10\u00de where the coefficients L ;M ; P ;Q are constants for a given mesh point One can find more details about the variables in Eq" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001153_amr.189-193.1882-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001153_amr.189-193.1882-Figure1-1.png", + "caption": "Figure 1", + "texts": [ + "what\u2019s more ,it will try to make up for all the current shortage of dump trucks. Here the designed high-dump truck with the scissor lifting mechanism can lift the compartment horizontally.The design of link length can lift the compartment horizontally at the same time move back the compartment.The flip machanism adopts the common cylinder direct-drive design. The automtic opening and closing machanism of the compartment door adopts the common four-bar linkage.Its mechanical principle is shown in figure 1. The virtual prototype model of the high-dump truck Click the icon \u201d extrude\u201d and set it as a new part.Set \u201dProfile\u201d as \u201dpoints\u201d and choose \u201dclosed\u201d.Set \u201dpath\u201d as \u201dabout center\u201d and \u201dlength\u201d =2000.Next choose the following points one by one (-300, 100, 0) , (4000, 100, 0) , (4000, 0, 0) , (-300, 0, 0) (-300, -100, 0) , (-2000, -100, 0) (-2000, 400, 0) , (-1100, 700, 0) , (-800,1200,0),(-300,1200,0).Change the grid and view to YOX plane.Draw 4 cylinders by clicking the icon .Define the place of 4 wheels by booleans \u201dsubtract\u201d" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000083_ir-03-2018-0042-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000083_ir-03-2018-0042-Figure4-1.png", + "caption": "Figure 4 Hip joint description", + "texts": [ + " The coupling has two bearings which are perpendicular to the flexion/extension movement to allow abduction/adduction movement. Flexible material is used to augment the wearer while walking. The lower side of the coupling is the base of rotational joint, while in between, there is a thrust bearing to support and absorb the axial force. The telescopic bar is used to adjust the link of the thigh link of the wearer which is fastened by bolts on both sides. The detailed design of the hip joint is shown in Figure 4. The knee joint is a key point in the functioning of the wearer while walking. It connects the thigh link with the shank link. Two ball bearings are fixed on the bearing mount, while the motor shaft passes through bearing mount. This shaft is attached to the motor to transmit the angular movement to the shank link. The motor is fastened on the other side by using four bolts and spacer, whereas a locker pin is used on the other side to prevent axial motion of the shaft. The wearer is attached to the wearable vehicle through a fixable tie" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003646_2017-01-1732-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003646_2017-01-1732-Figure1-1.png", + "caption": "Figure 1. In-line 4 cylinder engine consisting of IC unit and EMI reciprocating piston unit made of conducting material. Parts from left to right-Wheel, Transmission, 2nd Generator Starter Motor, Valves, Cooling Jacket, Solenoid, 1st Generator (Coils to produce electricity when piston moves up and down)", + "texts": [ + " The dimensions of the IC engine pistons are set according to EMI pistons and the power produced by total number of pistons with fixed dimensions is equated to the maximum power produced by all the EMI pistons. It is connected to crankshaft through connecting rod in usual way. The cooling system used to cool these pistons makes use of voids which has a running coolant to carry away excess heat. The reciprocating motion of piston can be achieved by two ways- 1. Piston made of permanent (Refer figure 2). 2. Piston made of conductive material (Refer figure 1). In case of piston made with permanent magnet, the coil or coils can exert both repulsive and attractive force on the piston. If there are two coils they will be wound and connected so that their like poles face each other, so that when (for example) the poles facing the piston are both negative, one pole will attract the piston south pole while the other will repel its north pole. When the armature reaches the extreme of its movement, polarity to the coils is reversed. In case of piston made of conductive material, eddy currents are produced on top surface of piston which has polarity opposite to that of coils in cylinder head thus pushing the cylinder away, in these high repulsive forces are induced", + " Main Generator: It consists of an electric motor which is coupled to crank shaft and generate electricity while the vehicle is moving and produced electricity is stored in battery. The main generator is also used as a starter motor. To crank the engine. The main purpose of this generator is that it can be instantly used to generate power and store it in battery. 2. Secondary Coil generator: The coil winding is done perpendicular to the motion of the piston surrounding the cylinder of EMI piston(refer figure 1).This generator acts as an auxiliary to Main generator by producing electricity when EMI piston is shut and can also produce electricity when EMI piston is working. This system makes the EMI engine even more efficient. This two way charging mechanism ensures instant charging of battery. This mechanism makes use of the basic concepts and hence it is used to supplement the additional electrical power. Incorporating Regenerative Braking System to this further more increasing the efficiency. EMI pistons operate due to Electromagnetic repulsion produced by solenoid" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002162_icecs.2008.4675079-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002162_icecs.2008.4675079-Figure3-1.png", + "caption": "Fig. 3. System level model of the typical SC integrator of Fig. 1", + "texts": [ + " During a first phase, \u03c61, a voltage v1(t) is sampled onto capacitor Cin. During a second phase, \u03c62, the charge on capacitor Cin is transferred to the feedback capacitor (CFB) where it adds to the charge already present on CFB . With the charge conservation rule we find: 978-1-4244-2182-4/08/$25.00 \u00a92008 IEEE. 1221 Cinv1(kT ) + CFBvout(kT ) = Cinv2((k + 1)T ) + CFBvout((k + 1)T ) (1) Or, in the z-domain, we derive: Vout(z) = Cin CFB 1 1 \u2212 z\u22121 ( z\u22121V1(z) \u2212 V2(z) ) (2) Hence, from a system level point of view, a SC-integrator can be modeled like shown in Fig. 3, with a = Cin CF B . This corresponds to the rectangular part of Fig. 2. Dielectric relaxation on a single capacitor can be modeled in several ways, but for the simulation of our test case the best approach is to replace all capacitors by the model of Fig. 4 [2]. In this model to each ideal capacitor some parallel branches consisting of a resistor (Ri) in series with a capacitor (Ci) are added. The ratio Ci/Ri is a good measure for the severity of the dielectric relaxation phenomenon, the higher Ci/Ri the worse the dielectric relaxation" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000206_tmag.2013.2242087-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000206_tmag.2013.2242087-Figure2-1.png", + "caption": "Fig. 2 shows the illustrative representation of these three types of magnetization patterns [4]. The radial and tangential components of the magnetic flux density in the air-gap/winding region can be, respectively, expressed as [4]", + "texts": [], + "surrounding_texts": [ + "The 2-D analytical magnetic field distribution of slotless brushless PM machines due to PM and AR is first presented. Based on the analytical magnetic field expressions, the vital quantities required for the analysis of brushless PM machines such as electromagnetic torque, reluctance torque, back-EMF, and inductances are calculated. Fig. 1 shows the key geometric parameters of a brushless slotless PM machine with surface inset magnets. The assumptions made in this study are similar to those listed in [4] and [5]." + ] + }, + { + "image_filename": "designv6_24_0003929_rast.2013.6581342-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003929_rast.2013.6581342-Figure5-1.png", + "caption": "Fig. 5. Structure for Satellite Quetzal V.I", + "texts": [ + " The satellite bus consists of the following subsystems, which will be developed under requirements and constraints of the selected payload: navigation, attitude determination and control subsystem (NADCS), power supply subsystem (PSS), structural subsystem (SS), thermal control subsystems (TCS), telecommunications subsystem (TS), data handling and processing subsystem (DHPS), telemetry subsystem (TMS) and propulsion subsystem (PS). For the correct integration stage, it has been considered the interface subsystem. Also will be performed the reliability and electromagnetic compatibility analysis, and as part of our recent research field, the sustainability analysis. Preliminary work on the Quetzal satellite is shown in the following figures (Fig. 5, Fig. 6 and Fig. 7). The first proposals are currently used for mechanical, thermal and vibration analysis, which will be verified later on mock ups, while other systems are developed in a engineering model stage with COTS and low cost elements by the student groups at UNAM. Once the electronic design and software has the first round of test, we will evaluate if groups from other institutions will be invited for the integration of the full engineering model, and definition would be made for the flight model based on reliability, cost and execution" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000083_ir-03-2018-0042-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000083_ir-03-2018-0042-Figure3-1.png", + "caption": "Figure 3 Trunk mechanism description", + "texts": [ + " Lastly, the seating mechanism enables the wearer to remain in a comfortable seating position, while the transferring process can be carried automatically between the twomodes. 3.1Walkingmode The wearable vehicle is designed to support a human while walking. It has three active joints along a sagittal plane which are hip flexion/extension, knee flexion and ankle flexion/ extension. Moreover, the other joints have elastic material to add force to the wearer while walking. In this section, one can describe the mechanical components that are used in this system as follows. The trunk mechanism is designed to support all sub-systems as shown in Figure 3. It is attached to the hip joint by flexion/ extension joint and rear wheels mechanism on one end while the other end is a revolute joint. This revolute joint is actuated to assure that the center of gravity stays within a certain polygon. The two powered anthropomorphic legs have ankle, knee, and hip joints similar to human legs and can be adjusted to accommodate all thigh and shank heights. The wearable vehicle is rigidly attached to the operator at the feet via custom specific boots, and bindings are provided at the torso via a custom vest" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002704_imece2016-65509-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002704_imece2016-65509-Figure1-1.png", + "caption": "Figure 1 Geometry of a journal bearing", + "texts": [ + " The radial clearance of the wave bearing is defined as the deviation between the mean circle radius and the journal radius. The clearance is on the order of one thousandth of the shaft radius, and the wave amplitude is nominally up to one-half the clearance. The position of the feeding groove is generally be located at the wave peak which is the furthest from the geometric center of the bearing house. 1 Copyright \u00a9 2016 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/conferences/asmep/90982/ on 03/13/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use Figure 1 shows the sketch of a three wave bearing. The wave bearing performance depends on the position of the wave relative to the direction of the applied load W. this position can be defined by the wave position angle, which is the angle between the starting point of the waves and the line of center of the bearing and the journal. The wave amplitude, the number of waves, as well as the wave position angle, are the basic design parameters of the wave journal bearing. Dimofte [1-3] has done lots of theoretical and experimental research on the hydrodynamic lubrication wave bearing", + "org/pdfaccess.ashx?url=/data/conferences/asmep/90982/ on 03/13/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use which one can determine a velocity and temperature profile. Many factors play a role in the dependent parameters of oil pressure and temperature in bearings, such as rotor weight, forces generated during operation, bearing geometry, fluid properties, rotational speed (RPM), and bearing type. All these elements must be considered when designing bearings for any application. Fig.1 shows a schematic view of a cylindrical bearing. The journal spins with angular speed \u03c9 and its center Oj, due to dynamic loads, also describes translational motions within the bearing clearance. The bearing or housing is stationary in most applications with the center of Ob. The hydrodynamic journal bearing consists of a continuous specific profile with a radius of curvature \u03c1b on the stationary part of the bearing. In Fig. 1, the clearance between the journal and the sleeve is greatly exaggerated to better visualize the geometry. Actually, the clearance is around a thousandth of the diameter. The hydrodynamic film pressure builds up within the small gap between the rotating shaft and the bearing house. The hydrodynamic journal bearing performance depends on the shape of bearing sleeve and the position of the journal center Oj to the direction of the applied load W. The radius of curvature of the bearing house is obviously a periodic function of \u03d5 with the period 2\u03c0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001536_0167-9260(96)00003-x-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001536_0167-9260(96)00003-x-Figure4-1.png", + "caption": "Fig. 4. Rerouting a projected dependence.", + "texts": [ + " Next, we give a few concepts concerning the inclusion of topological constraints in the overall mapping. During the projection from dependence graph to virtual array, the choice of the projection vector can be restricted, based on the desired interconnection pattern. Two are the issues involved. On the one hand, it is required that data can be rerouted by using the desired interconnections only. On the other hand, one has to make sure that the timing correctly reflects the ordering among operations. These conditions can be imposed on the projection and scheduling matrix accordingly. Fig. 4 shows the simple case of a dependence, D, projected onto D', which needs to be rerouted along the \"admitted\" interconnections given by the dashed arrows. It is equally viable to constrain the projection on the basis of the desired location of I/O operations. For instance, it is very often convenient to have input/output operations performed only by the processors at the edges of the array. This constraint translates into a requirement on the projection matrix U. In Fig. 5, the requirement of having I/O ports on the edges of the array translates into a constraint on the projection vector u" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003975_apmc.2012.6421484-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003975_apmc.2012.6421484-Figure4-1.png", + "caption": "Fig. 4 The photograph of the developed Wilkinson balun.", + "texts": [ + " As shown on Fig. 2, considering a 180\u00b0\u00b13\u00b0 phase difference, the proposed balun has a 72.2 % bandwidth, but the conventional one only has a 49 % bandwidth. The developed Wilkinson balun is designed at the center frequency, 1.5 GHz, and fabricated on a FR4 substrate with the thickness of 0.8 mm, the dielectric constant of 4.4, and the loss tangent of 0.02. The MM TL is realized by Murata 0603 chip inductors and capacitors connected with a section of microstrip line. The photograph of the circuit is shown in Fig. 4. In this paper, the measured results are carried out by Agilent E5071C four-port vector network analyzer, while the simulated results are obtained by Agilent ADS and Momentum. Fig. 5 shows the measured and simulated Sparameters of the developed Wilkinson balun. The measured |S11| is better than -10 dB from 0.68 GHz to 2.52 GHz, while the isolation |S23| is larger the 10 dB over the frequency range of 0.5-3 GHz. At 1.5 GHz, the |S21| and |S31| are -3.3 dB and - 3.7 dB, respectively. Considering the 180\u00b0\u00b13\u00b0 output phase difference, the frequency range covers from 1" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003189_978-3-030-49916-7_102-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003189_978-3-030-49916-7_102-Figure5-1.png", + "caption": "Fig. 5. Sectional diagram showing possible reinforcement solutions (a) BJ formwork; (b) fully 3DCP segment reinforced with 3DCP mesh and post-tensioning cables; (c) 3DCP reinforced with CFRP mesh and integrated formwork for main beams, cast in-situ with reinforced concrete. 1-3. 3DCP layers of concrete, with 1.5% CAC, 5% CAC and 10% CAC; 4. CFRP reinforcement mesh; 5. Post tensioning tendons; 6. Brass L and I profiles for joint detailing; 7. Reinforced concrete cast in-situ.", + "texts": [ + " Nevertheless, proper structural dimensioning and structural testing are essential and will be an integral part in future investigations. Moreover, the ribbed ornamented pattern can be integrated as part of the structural concept and dimensioned accordingly. 3D printing the beams as solid elements was not equally successful (Fig. 3d and 5b). Even if 3DCP is fast, it cannot compete with the fabrication speed and structural performance of cast concrete. This is why, for future implementations, only the outer contour of the ribs can be 3D-printed, as a formwork instead of printing the full rib (Fig. 5c). Consequently, the system will become lighter during transportation and will work similarly to a Filigree Wideslab method, in which the beams are cast in-situ with conventional reinforced concrete. This fabrication method allows for extensive morphological variation. The process is based on using different fabrication resolutions. First, 3D printing of high-resolution formwork shapes the most detailed part of the slab that is apparent in the inhabited space. Even if the formwork can be reused, the strength of the process lies in the possibility to customize and replace the formwork as the design requires" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001792_ipemc.2006.4778133-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001792_ipemc.2006.4778133-Figure1-1.png", + "caption": "Figure 1. SR-PM motor structure", + "texts": [ + "00 \u00a92006 IEEE IPEMC 2006 I. INTRODUCTION A synchronous reluctance motor with permanent magnet assisted (SR-PM) is a combination of a synchronous reluctance motor (SynRM) and a permanent magnet synchronous motor (PMSM). Because a SR-PM motor has many good features, such as high power density and power factor, excellent efficiency and a wide constant-power speed range, it has been a focus of research recently. The rotor structure of this motor has many different types. The rotor can radial-laminated or axially-laminated. Fig.1 shows a SR-PM motor [1] with a axially-laminated rotor with the permanent material added in between and mounted on the steel bracket. The stator is the same as that of a common induction motor (IM). Since SR-PM machines share both the characteristics of SynRM and PMSM, the research on SynRM and PMSM is a good start [2][3][4][5][6][7][13][12]. For example the analysis of the axially-laminated SynRM by Staton D.A [3], the analysis and design of axially-laminated interior permanent magnet motor with the shape of sandwich by Soong W" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000862_optronix.2019.8862342-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000862_optronix.2019.8862342-Figure1-1.png", + "caption": "Fig. 1.Schematic representation of HMSIW based semi- circular antenna excited with coaxial feed.", + "texts": [ + " Semi-circular based SIW cavity backed hybrid antennas as well as HMSIW semi-circular antenna have already been designed, fabricated and reported in literature [10-12].A proposed antenna which is SIW based and a slit has been introduced to miniaturize the same by the authors in this paper. A. Half Mode Substrate Integrated Waveguide SemiCircular Antenna Semi-circular antenna which is Half Mode Substrate Integrated Waveguide based uses Arlon AD270 as the substrate with the thickness of h = 0.79mm, the dielectric constant is \u03b5r = 2.7 and tan\u03b4 = 0.002 is its loss tangent. It is excited with coaxial feeding as shown in Fig. 1. The antenna of radius is r = 14.5 mm (diameter = 29 mm) with the allinclusive measurement as 60 mm \u00d7 60 mm. The circumference of the semi-circular antenna is fitted with vias to make it a PEC wall. The side having no vias acts as PMC wall. The vias have a diameter of d1=1mm with distance between the centres d2=1.5mm. The fundamental TM010 mode is excited through coaxial feeding from point M at g = 10.5 mm as mentioned in the below given Fig. 1. The various measurements of the mentioned antenna are showcased in Table I. L W r h l1 l2 G 60 60 14.5 0.79 39 15.52 10.5 The optimized return loss after simulation turns out to be - 25.5 dB at 5.22 GHz while its value on measurement is -33 dB at 5.17 GHz as given in Fig. 2. The bandwidth of the impedance varies from 5.19\u20135.25 GHz covering 60 MHz .The radiation pattern obtained after simulation is shown in Fig. 3 where in the co-polarized gain turns out to be 6.5 dBi of the simulated E-plane while it is measured to be of a value of 6" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000277_iros.2011.6048434-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000277_iros.2011.6048434-Figure1-1.png", + "caption": "Fig. 1: Continuum arm schematics", + "texts": [ + " Then, the length of the any actuator at any time is LiO + lij (t) where LiO is the original length of actuators and t is time (Fig. la). Length variables, defined as lij(t) E [li,min, li,max], can be positive or negative to represent extension and contraction where i E [1, N] is the section number, j E [1,3] is the actuator number, li,max E jR+ is the mean maximum extension for extending actuators, and li,min E jR- is the mean maximum contraction for contracting actuators. Following a similar procedure as in [15], as shown in Fig. 1 b, the spatial orientation of section i upon actuation can be defined by {Ai, \u00a2ii, Bd with respect to the local coordinate system (coordinate set 0iXiYiZi in Fig. la) as given in (2) where Ri is the section radius. This approach can be applied to any geometrically constrained variable length continuum arm structures, planar or spatial, with a differing number of actuators per section by deriving {Ai, \u00a2ii, Bd (in planar case, {Ai, \u00a2id) in variable actuator lengths. By employing standard homogeneous rotational and trans lation matrices, the homogeneous transformation matrix (HTM), T, for points along the neutral axis of the ith section is derived as (3) where Ai = Ai (qi) , qJi = \u00a2i (qi) , and ()i = ()i (qi) \u00b7 Rz, Ry are rotational matrices about the Z and Y axes", + " Therefore both the forward and inverse kinematics can be easily computed without the need for separate singularity resolving methods such as those used in [9], [12]. Further, in this method, without having multiple MSFs to cover the workspace, the errors, eij, of T and T ep elements are simply calculated as eij = tij - tepij. Therefore, this approach eliminates (i) the requirement for mode switching schemes, (ii) mapping complications (\" .. it is solved directly in joint space), (iii) possible modal singularities [12], and (iv) complicated error models. Having obtained the MTM for any ith section, the result can now be extended to multiple arm sections. Fig. 1 b depicts a general multi section arm schematic. By employing classical coordinate transformation techniques, the MTM for any point along the neutral axis of the full continuum arm is derived with respect to global coordinates OXY Z (Fig. Ib), to which the base section of the continuum arm is attached, as N rr N {k } oTcp(e,q) = k=l k-ITcp(\ufffdk,qk) Tk = [ Rcp (e,q) Pcp (e,q) ] Olx3 1 (6) where q = {[ qi qf qrv V : q E R.3Nx I} is the joint space vector that uniquely describe configuration of the entire arm and Tk is any joint transformation present at the section joints", + " This is an appealing feature for computer implementations because MSFs can be defined as functions that take {qi' LiD, Ri} as the input for the ith section. As previously mentioned, this means that only 9 such modular functions (6 from CPR and 3 from cPp) are required to completely define the position and orientation of any geometrically constrained variable length multisection continuum robotic arm with any number of sections. Due to the geometrically constrained design, our model is only capable of rotations about the x' and y' axes of the neu tral axis reference frame 0' x' y' z ' (Fig. 1 b), and incapable of torsional (rotations about the z ' axis) movements. Defining the rotating Euler angle vector 'P = [a (3]T, where a and (3 are the rotations about the X and Y axes at e respectively, the following relationship can be derived. (7) Orientation angles a and j3 are easily computed as a(e, q) = arctan2(r32' r22) j3(e, q) = arctan2(r13' -rl1) (8) where rij are elements ofR and {a, j3} E [-7r, 7r]. In the case of torsional movements, this is easily accommodated using an additional Euler angle parameter \" to represent rotation about the Z axis" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003283_870182-Figure11-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003283_870182-Figure11-1.png", + "caption": "Figure 11", + "texts": [], + "surrounding_texts": [ + "870182 5\n3.2 lABORATORY BURST TEST\nCharts 1 and 2 depIct the test results obtained from burst tests of two IlHSK\" assembled radIators and two tabbed header assembled radiators, respectively. Each plot represents a separate burst deflection test on an IndivJdual radiator.\nTable 4 shows that the \"HSK\" assembled radlators '111 thstood higher press ure be fore I eaks occurred, as well as hIgher maxlmum pressure to ultImate fal1ure\nthan the TAB method.\nRad I ators us 1ng the \"HSK\" tank attachment method fa i I by deforming the header as schematically described In Flgure 10. The radlators with the TAB tank attachment method falled because the tabs simply 11fted or deflected, completely releasing the tank. Modfne Is continuIng burst test studies as well as startIng pressure cycle and vlbratlon tests. A combined test\nto determlne how pressure cycle loadi ng affects burst pressures will also be conducted. Results of these SUbsequent tests were not avallable at the tJme of this paper.\nIt should be noted that the radlators using the tabbed\nheader tanks \"bl ew off\" the headers. With the \"HSK\" radiators, the plastic tanks remalned seCured to the header, and allowed the pressure to escape gradually. This fact must be considered heavlly with respect to quality and rellabillty concerns of today.\nI I\nRadialor - #1 & #3 Header Mall. - .040 Method - HSK & TAS\nKey I RADNG.:! HSK RAONQ.4 TAB\nChart 1\nT~k 61cwQfl\n/\nPRESSURE V. DISPLACEMENT\nPRESSURE V. DISPLACEMENT Radiator - 112 & #4 Header MatI. - 0.032 Method - HSK & TAB '00 ,,-,--,-,---,-,--,-,----,---,--,-,---T'T=--r-.e:::;.:-::+:; I I\n00\nno\n70 i5 ~ 60 \"2, 50 \u00a3 Leak I'rc~5ufe\n\"\nChart 2\n010 060 000 ,100 120 l~O ,160 .100 20ll 220 240260 ,280 ,JOG ,320 .340 Di~plocemenl (lnchos)\n'0 ,-..,'r:':c\"'.;'C:\"::::\"::;\"'::::'-,!\"\",,/\u00b7417'c..'-t--+---+_H ~ ---I-+--+-+---I--I--+---i r V' ~ '0 1--t---1J\"'1- 13 ~ 1+---11-+-+-+1 \u00a7 -+-+-+--+-+---+-1---1 to ~i/~,~''''-I--+I ~_ ~ -+--+-+---+-+-+-1---1 o 020\nThe charts presented provide the followIng Information:\n- Measured deflection versus pressure\n- Start of leaks (deflection and pressure) - Point at whTch tank \"blew off n headers or aIr\npressure could no longer be increased due to leakage\nThe actua I data po I nts used to deve I op the prevIous I y ment i oned charts SCCI'om, - Eac~ 2 Ir>c~~' long IOf 4 1~:llC~ TOlall\nGASKET COMPRESSION\n/~ ;0',\n010 f---+--t---+--i>-\u00ab:.--+---j---+---j---+--I /f-\"\n050f--+--+---,~j-=-+--+---t--+--+---t---I1/ -w\n0'\" f-_-+_~j_I'---'-;'j_; __-j-_-+__-j-_-+__+-_-+_-If- ,\", ~ (MO f----j,!:::::-\"I-.,---t--\n\"0 j-...,/4---+_-+-\n/", + "870182\ndoes not al low for hlghly accurate gasket compressJon, due to part tolerance stack up as well as tab spring back that may occur during manufacturing and actual use. The \"HSKIl plercings are gauged from the bottom of each Individual header. and thereby accurately fix the des 1red amount of gasket compress Ion. The added 3dvan'tage of applying uniform gasket compression offers Improved tank to header joint quality, and therefore. radiator durabJ I ltV.\n3.6 MANUfACTlJRING CONSIOEllATIONS\nThe tabbed header plastlc tank attachment method ut 1\\ i zes a header wi th notches bl anked out at the 5 ide walls to form tabs. These notches require a more\ncampl leated tool for header manufacture than 1t the side wail could remaln solid as It Is wIth the \"HSK\" method. Tabbed headers a I so requ I re speci a II zed tool ing for each tank and header model in order to bend the tabs over durlng the tank assembly operatIon. This tooling is made in such a manner that will allow It to accommodate fIttIngs, outlets, and ti ller necks on the plastic tanks. The nHSKII method on the other hand, uses an assembly tlxture that can be easIly adapted to many dl fferent tank and header confIguratIons without the expense of new toolIng. Figures 13 - 15 Illustrate a sImple nHSKlI assembly fIxture.\nSIMPLIFIED HSK MANUFACTURING CONCEPT\u00b7 STEP 1\nFigure 13\n7\nSIMPLIFIED HSK MANUFACTURING CONCEPT\u00b7 STEP 2\n,ISK PunCMf' (MOCll,ln,cnily AClUJIOdl\nFigure 14\nSIMPLIFIED HSK MANUFACTURING CONCEPT\u00b7 STEP 3\nIt shou I d a I so be noted that w1 th the IIHSKII assemb I y method, no dlrect pressure 1s appl1ed to the plast1c tank!s toot, whereas the tabbed header assembly method al lows the tabs to compress this foot during assembly, lnvariably inducing hJgh enough stresses to crack the plastic tank dur1ng assembly, unless extreme care 1s taken.\nIt appears quite possible, therefore, that the \"HSK\" assembl ed rad lator will reduce the amount of complicated tooling usually required to manufacture a plastIc tank radiator, thereby reducl ng the cost. In addltlon, It Is believed that the \"HSK\" method will contrIbute sIgnificantly to the overall QualIty and rei Jabil Ity of the radIator." + ] + }, + { + "image_filename": "designv6_24_0001638_tim.2015.2477160-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001638_tim.2015.2477160-Figure5-1.png", + "caption": "Fig. 5. Geometry between the robot and target in w-space.", + "texts": [ + " Finally, the target can be expressed in w-space as \u23a1 \u23a2 \u23a3 zl yl sin \u03b8l cos \u03b8l \u23a4 \u23a5 \u23a6 = \u23a1 \u23a2 \u23a2 \u23a2 \u23a3 ha + bs2w2 + l 2 s1w3 l 2 s2w4 \u2212 bs1w1 \u2212s1w1 s2w2 \u23a4 \u23a5 \u23a5 \u23a5 \u23a6 = \u23a1 \u23a2 \u23a3 ha 0 0 0 \u23a4 \u23a5 \u23a6 + A \u23a1 \u23a2 \u23a3 w1 w2 w3 w4 \u23a4 \u23a5 \u23a6 (14) where A = \u23a1 \u23a2 \u23a2 \u23a2 \u23a3 0 bs2 ls1 2 0 \u2212bs1 0 0 ls2 2\u2212s1 0 0 0 0 s2 0 0 \u23a4 \u23a5 \u23a5 \u23a5 \u23a6 (15) and a constraint is established in the last two terms of (14) as (s1 2w1 2 + s2 2w2 2 = 1). As a result, the w-space is defined in a 3-D space where w \u2208 R 3. Applying the time derivate in (14), the relative velocity between the robot and the target is obtained as \u23a1 \u23a2 \u23a3 z\u0307l y\u0307l \u03b8\u0307l \u23a4 \u23a5 \u23a6 = \u23a1 \u23a2 \u23a2 \u23a2 \u23a3 0 bs2 ls1 2 0 \u2212bs1 0 0 ls2 2\u2212s1s2w2 s1s2w1 0 0 \u23a4 \u23a5 \u23a5 \u23a5 \u23a6 \u23a1 \u23a2 \u23a2 \u23a2 \u23a3 w\u03071 w\u03072 w\u03073 w\u03074 \u23a4 \u23a5 \u23a5 \u23a5 \u23a6 . (16) To estimate the velocity between the robot and the target, a geometry in w-space needs to be established. As shown in Fig. 5, the robot and the target are defined as (zr , yr , \u03b8r ) and (zt , yt , \u03b8t ), respectively. Following the development in Section III-C, the relative configuration between the target and robot frame is given as (zl, yl , \u03b8l): \u23a1 \u23a3 zl yl \u03b8l \u23a4 \u23a6 = \u23a1 \u23a3 cos \u03b8r sin \u03b8r 0 \u2212 sin \u03b8r cos \u03b8r 0 0 0 1 \u23a4 \u23a6 \u23a1 \u23a3 zt \u2212 zr yt \u2212 yr \u03b8t \u2212 \u03b8r \u23a4 \u23a6. (17) And the time derivative of (17) is calculated as \u23a1 \u23a3 z\u0307l y\u0307l \u03b8\u0307l \u23a4 \u23a6 = \u23a1 \u23a3 cos \u03b8l 0 \u2212 sin \u03b8l 0 0 1 \u23a4 \u23a6 [ vt \u03c9t ] + \u23a1 \u23a3 \u22121 yl 0 \u2212xl 0 \u22121 \u23a4 \u23a6 [ vr \u03c9r ] . (18) From the relative kinematics shown in (18), the linear and angular velocity of the target in w-space can be estimated as vt = (z\u0307l + vr \u2212 yl\u03c9t ) cos \u03b8l + (y\u0307l + zl\u03c9r ) sin \u03b8l (19) and \u03c9t = \u03b8\u0307l\u03c9r " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003772_313-FigureI-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003772_313-FigureI-1.png", + "caption": "Figure I. A", + "texts": [ + " In particular, a type of elastic hysteresis loop which has not previously been noticed has been observed to occur in certain conditions, while in other circumstances lead has a definite elastic range in which Hooke's law is obeyed. 5 2. E X P E R I M E N T A L M E T H O D The apparatus consists essentially of two parts, of which the respective functions are to apply a known stress to the specimen and to measure the strain produced. The main frame of the apparatus consists of two vertical brass plates, each about 25 cm. square. held parallel to each other at a distance apart of about 8 cm. by horizontal brass bars screwed to the plates. The disposition of the cross-bars is shown in section in figure I (A, F , 2, 2). Application and measurement of the stress. The specimen S, figure I a, is soldered at its lower end to the centre of the cross-bar A, and its upper end is soldered to the end of the brass rod B, so that S and B are collinear. The upper end of B is fixed to a cross bar C through which an upward force is applied to B and S. Near each end of C, 29 cm. from B, is a screw held vertically by a screw thread in C and a lock nut, 3 54 Bruce Chalmers and from the lower end of each of these screws a steel gramophone needle point projects. These points rest in punch marks on the two bars D, which are rigidly connected to form the lever through which the stress is applied", + " In order to determine the strain of the specimen corresponding to any stress, it is necessary to measure changes in the distance between two points on the specimen. These measurements were made by an optical interference method as follows. Two screws K, 4 cm. apart as measured along the bars 2, 2, each carry a gramophone needle, point upwards. On these points is supported a brass frame L, to which are rigidly attached a flat piece of glass M , and an arrangement N through which contact is made with the specimen S. N consists of two similar brass plates, shown in figure I , between which are held two halves of a razor-blade of the three-hole type. The position of the razor-blades, which lie in a horizontal plane, is indicated by the dotted lines in figure I b . The distance apart of the razor-blades is adjusted so that they just cut the surface of the specimen when it is inserted in the position S. The second frame P of the interferometer, figure I a, holds a second glass plate Q and a second arrangement R similar to that shown in figure I b . The frame P is supported on L at two points by means of two screws T , the actual points of support being gramophone needle points projecting downwards from the screws T. The distance and angle between the glass plates Q and M can be adjusted by means of the screws T. The frames L and Pa re each counterbalanced by the adjustable weights U and v, so that no stress is applied to the specimen S through R or N . The interferometer is illuminated from above by means of light from a mercury arc passing through a water cell and a monochromatic green filter, and reflected downwards from the glass slip X . The fringes produced by the interference between the light reflected upwards from the lower surface of Q and from the upper surface of M are viewed through the microscope W. The direction and spacing of the fringes, which are the loci of points of equal separation of the plates M and Q, can be adjusted by means of the Screws T. This adjustment is made SO that the fringes form a series of parallel lines ,vhose direction is perpendicular to the plane of figure Ia, i.e., parallel to the line TT of the fulcrum of the interferometer unit. When the distance RN alters, the fringes move in a direction perpendicular to their length. The plates Q and M were pieces of good plate glass mirror, and it was found that the fringes obtained when there was no silvering on either plate were sufficiently sharp and sufficiently straight for satisfactory measurements to be made. Precautions taken to avoid disturbances. In order to avoid disturbances due to vibrations of the bench on which the apparatus stood, the two frames P and L of the Interferometer were held together by two steel springs in the plane of the two Screws T", + " The critical stress is 9.0 kgm./cm? This point is also the lower limit of strains which result in permanent set observable by the present method. A third characteristic of this critical strain is that when it is exceeded a change takes place in the type of elastic after-effect obtained. The method of investigation of the after-effect was to change the tension on the specimen by definite steps, usually either about 1500 gm./cm? or 3000 gm./cmi corresponding to changes of water-level in the tube H , figure I (a), of I cm. and 2 cm. respectively. These changes of tension were made in the direction of either increasing or decreasing stress, and after each change the movement of the fringes was observed until it was less than 0-01 fringe per minute, which corresponds to a change of length of the specimen of about IO-' cm./cm. per minute. The type of curve obtained when the tension did not exceed the elastic limit was always that shown in figure 3 ( U ) in which several curves, selected at random, for these changes of stress are shown", + " Various experiments have been made with a view to eliminating this possibility. Any cause which would prevent the upper plate of the interferometer from moving freely, so that it lagged behind its true position, would cause a loop such as was found to appear. Two possible ways in which this might happen suggest themselves, the first being that some friction might prevent free movement of the upper plate, and that the upper interferometer mirror might not move in proportion to the movement of R (figure I) until R had moved by a certain amount. This would cause the length to appear too small when it was increasing and too large when decreasing. A similar effect on the lower plate would give the reverse effect, and a loop of apparently negative area would result. The effect of friction is usually to prevent all movement until a limiting force is reached, when movement suddenly starts; this is not what happens with the A n interference extensometer and observations on elasticity of lead 365 hysteresis loops, since the ends of the loops are curved and not flat" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001387_acdt.2018.8592940-Figure10-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001387_acdt.2018.8592940-Figure10-1.png", + "caption": "Fig. 10. Axial velocity of the flow at 1R after propeller disk of the actual model.", + "texts": [], + "surrounding_texts": [ + "The author would like to thank the aeronautical engineering research team at the Defence Technology Institute for their helpful discussions around the topic of the paper." + ] + }, + { + "image_filename": "designv6_24_0000159_j.geothermics.2016.02.002-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000159_j.geothermics.2016.02.002-Figure2-1.png", + "caption": "Fig. 2. Axisymmetric nonlinear FEM model of the IDDP-1 well. The anchoring effect of the couplings in the concrete is included in the model by using bonded contact at the coupling locations. A simplified wellhead is also included to account for wellhead pressure and the interaction between the production casing and the wellhead. The casing depths are shown to the right by scaling the x-axis by 1000 to 1.", + "texts": [ + " ydril/Tenaris 563 connections were selected for the production asing and anchor casing and the top part of the anchor casing as designed for creep and rupture conditions (Th\u00f3rhallsson et al., 014). The anchor casing therefore consists of two different thickesses and steel grades. The top 290 m, consist of 15.9 mm thick PI grade T95 steel casing with a diameter of 13 5/8\u2032\u2032 and the lower ection, below 290 m, consist of 13.1 mm thick API grade K55 steel asing with a diameter of 13 3/8\u2032\u2032. A nonlinear finite-element model (FEM) created in Ansys as used to model the well. The model, see Fig. 2, is twoimensional and axially symmetric around the center of the well. he stress\u2013strain curves of the steel grades that are implemented n the model were obtained from tensile tests by Karlsdottir and horbjornsson (2009). The curves are converted to true stress\u2013true train before implementation. Strength reduction at elevated temeratures is accounted for by scaling the curves according to uidelines in the recently updated New Zealand standard NZS 403:2015, \u201cCode of Practice for Deep Geothermal Wells\u201d" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002170_1754337120980627-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002170_1754337120980627-Figure5-1.png", + "caption": "Figure 5. Diagrams of whip test apparatus from initial swing to impact (1\u20133).", + "texts": [ + " Dimensions and masses of human body segments found from anthropometric databases for approximate estimation of characteristics of 5% female white non-Hispanic.21. Segment Upper arm Fore arm Hand Length (L) 30.33 cm (11.94 in) 23.80 cm (9.37 in) 17.60 cm (6.93 in) Mass (M) 1.56 kg (3.45 lb.) 0.90 kg (2.00 lb.) 0.33 kg (0.73 lb.) M/L ratio (kg/cm) 0.051 0.038 0.019 calculated from the assumption of a homogeneous material with the mass and dimensions of the hand shown in Table 2. Figure 4 shows photos of the actual model with whip attached. The torsion springs and potentiometers can also be observed. Figure 5 shows diagrams of the whip test apparatus in three positions: start of swing (1), mid-swing (2) and impact (3). Elastic modulus of the target impacted by whip end To compare the test apparatus to the performance of an actual whip used in Thoroughbred horse racing, it was necessary to match the stiffness (modulus) of the target, as well as the impact portion of the whip. The whip functions as the projectile in the impact event and the rise time is dependent on the padding of the whip, as well as the modulus of the target" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001154_0954407011525557-Figure12-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001154_0954407011525557-Figure12-1.png", + "caption": "Fig. 12 A new design of the upper bracket", + "texts": [ + "from the upper bracket fastening points to the joint between the upper tube and the upper bracket. The shortened path (capsule\u2013bracket plate\u2013locking bolt\u2013 3 DISCUSSIONback plate\u2013upper tube) would further increase the overall stiVness of the upper bracket. Modal analysis reveals that the shape of the back plate is also important. To Because of the bending exibility of bellows the rst natural frequency of a bellows-type steering column isreduce local bending, the shape of the back plate can be further modi ed as shown in Fig. 12. The locking bolt signi cantly reduced relative to the conventional design. It is obvious that the rst natural frequency has to beand channel has been moved closer to the tube. The back plate has been modi ed such that local bending is pre- increased by modi cations to parts other than the bellows. In many cases the rst natural mode of vibrationvented along the weld lines. In the second sensitivity analysis, the lower bracket parameter has been replaced is a vertical bending mode with its stationary nodes at the upper and lower bracket fastening points", + " 10 An upper bracket with ribs added D12099 \u00a9 IMechE 2001 Proc Instn Mech Engrs Vol 215 Part D at Fachhochschule Osnabr\u00fcck on April 19, 2015pid.sagepub.comDownloaded from Table 5 An L8(27 ) orthogonal array with analysis results for the bellows-type steering columns with respect to the second set of design parameters C B First natural Run A F A6F (mm) A6C F6C (mm) frequency (Hz) A, bracket plate ribs (A1, no ribs; A2, ribs added; see Fig. 10); B, bracket plate thickness (mm); C, capsule length (mm); F, back plate and locking bolt assembly (F1, no modi cation; F2, modi ed design; see Fig. 12). The major part of this study has been done in the framework of a research project from Hyundai Motor Company. The authors would like to oVer special thanks to the Vehicle Development and Analysis Team of Hyundai Motor Company. 1 Kim, H. Study on the design of automotive steering column with metallic bellows. Master\u2019s thesis, Korea University, 1997. 2 Chiang, S. L. Using experimental medal modeling techniques to investigate steering column vibration and idle shake of a passenger car. SAE paper 850996, 1985" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001397_icitacee.2016.7892413-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001397_icitacee.2016.7892413-Figure6-1.png", + "caption": "Fig. 6 Mechanical design", + "texts": [ + " There is also a separate parameter used to limit the Integral parameter\u2019s scope of what it can do so as not to overreact. III. SOFT CONTACT LENS INDEXER DESIGN The design of the system was developed as Fig. 4. The smart motor as an actuator was controlled by computer using RS 232 cable data. Driver motor and rotary encoder, as a sensor, installed in smart motor block system. Fig. 4 Soft contact lens indexer design Mechanical design of the machine was designed to use two motors indexer. There were motor rotary axis (R-axis) and the linear motor (Y-Axis) as shown in Figure 6. The design was based on the condition of the factory and the ergonomics aspect. The red line is area of contact lens and moveable equipment. The flow chart of the system can be seen in Figure 7. The rotary index table movement depends on the rotary encoder that was mounted on the gear box. Rotary encoder relative (incremental encoder) was obtain to reach the desired ICITACEE 2016 70 values of radial velocity on the wheel. The direction of the index rotary tables was determined by the values that was read by sensor as function of the speed motor" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002471_1.365458-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002471_1.365458-Figure3-1.png", + "caption": "FIG. 3. The magnetic flux distribution between the stator core and the rotor magnet B .", + "texts": [ + "1\u20133 The cogging torque of the aquatic ac synchronous motor is decreased quickly as the trough width of the stator core is reduced. For the trough width of 11 mm with the rotor magnet B , the maximum cogging torque of the motor is 179.54 g\u2022cm, while the maximum cogging torque for the trough width of 13 and 15 mm are 231.44 and 282.65 g\u2022cm, respectively. This phenomenon can be explained by analyzing the magnetic flux distribution between the stator core and the rotor magnet. The magnetic flux distribution with the trough width of 11 and 15 mm as shown in Fig. 3. The magnetic flux distribution of the trough width of 11 mm is smoother than 15 mm. The magnetic profile of the rotor magnet can affect the cogging torque of the motor markedly. The different magnetic profiles of the rotor magnets were made by using different magnetizing fixtures. Figure 4 shows the magnetic field distribution of the two-poles rotor magnet at a distance of 0.3 mm over the magnet. The material character of the stator core and rotor magnet as shown in Fig. 5. The maxi- mum magnetic field strength is about 1232" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000912_mwsym.2002.1011896-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000912_mwsym.2002.1011896-Figure1-1.png", + "caption": "Figure 1 shows the circuit structure of the proposed frequency converter consisting of sub-harmonic direct quadrature conversion mixers integrated with a planar antenna at a 40 GHz RF carrier. A direct-coupled rectangular microstrip patch antenna is designed and integrated. The inset feed and transformer to the antenna are used to match impedance.", + "texts": [], + "surrounding_texts": [ + "IF-WE-33 Millimeter Wave Direct Quadrature Converter Integrated with\nAntenna for Broad-Band Wireless Communications Ji-Yong Park, Seong-Sik Jeon, Yuanxun Wang, and Tatsuo Itoh\nDepartment of Electrical Engineering, University of Califomia, Los Angeles, 405 Hilgard Ave., Los Angeles, CA 90095-1594, U.S.A.\nE-mail : jypark@ee.ucla.edu\nAbsiraci - A compact quadrature APDP direct conversion mixer design, the need for costly millimeter wave LO sources is eliminated as well as the\nThis paper is organized as follows: The overview ofthe\nbuilding block of the system including the antenna and the quadrature modulator/demodulator is introduced and the perfoFance is validated with measured results. Finally, a communication link is set up by using a pair of integrated modulator/demodulator integrated with a 40 Gna patch antenna for millimeter wave wireless applications is additional IF mixers and filters, proposed. Anti-parallel diode sub-harmonic mixers are constructed for broad-band direct quadrature conversion. Overall phase and amplitude imbalance between the mixer I f r 0 n t - d system is first given. Then the design of each and Q output channels are less than 1.2 0 and 1 dB respectively. An average conversion loss of mixers of -14.6 dB is achieved in the frequency range from 39.75 GHz to 40.25 GHz. A communication link is built based on a pair of the proposed front-ends. Data transmission at up to 1 Gh/s data rate for QPSK modulation Is successfully demonstrated.\nfront-ends to examine the data mo&laGon, transmitting, and demodulation function, The link shows\nI. INTRODUCTION\nThe limited available spectrum at the lower end of the microwave bands is due to the recent dramatic growth in demand for wireless communications. Many services requiring high data rate such as LMDS, WLL, wireless CATV, point-to-point radio are tuming to millimeter frequencies for more available spectrum and broad-band Capability. Compact and low cost designs of millimeter wave communication front-ends thus become necessary for extensive applications.\nIn this paper, a direct quadrature frequency converter integrated with an antenna is proposed. A 40 GHz patch antenna is designed to have 10 dB bandwidth of 1.64 GHz for broad-band communication. The integrated antenna\nsuccessful transmission of QPSK data at up to 1 Gbls data rate.\n11. FRONT-END CIRCUIT OVERVIEW\ndesign reduces the interconnection loss that-is a significant issue at millimeter frequencies [ I ] . The quadrature mixer provides the direct conversion capability for digitally modulated signals. More importantly, the existence of I/Q\noutput channels can greatly reduce. the post-stage signal f ~ - z O c ~ - :cm\" - processing load and increase the system throughput when an advanced system such as an adaptive beamforming a i p Reillor a i p a p o i t o r array is under consideration [2]-[3].\nI Chsnnr1 Sub-harmonic mixers with anti-parallel diode pairs design. They\nhave heen widely studied because of the advantage of low LO frequency, LO noise suppression and no bias circuit [4]. Many applications of sub-harmonic mixers with APDPs have been proposed for millimeter wave and submillimeter wave mixers [5]-[8], modulators [9]-[10] as well as direct conversion mixers [11]-[12]. By using an\nare for the quadrature Fig. I . The circuit architecture of the direct quadrature converter integrated with a planner antenna ( circuit size: I x 0.73 inches ).\nA sub-harmonic direct quadrature conversion mixer is composed of two pairs of anti-parallel diodes, open and short stubs, 45 phase delay line at 20 GHz, low pass\n- _.\n~\n1217\n0-7803-7239-5/02/$10.00 0 2002 IEEE 2002 IEEE MTT-S Digest", + "filters and Wilkinson power dividers for a RF of 40 GHz and a LO of 20 GHz, respectively. In order to terminate the RF and LO leakage, open and short stubs are optimized to a quarter-wave length at 20 GHz and a halfwave length at 40 GHz. Bandpass filters and capacitors work for IF decoupling. A phase delay line of 45 at 20 GHz is inserted in one of the LO power split paths after the LO Wilkinson power divider. This line will be 90\u00b0 long at 40 GHz ( even harmonic of 20 GHz ), and the two mixers can each generate in-phase and quadrature IF mixing signal. Two identical direct quadrature converters, which are integrated with an antenna are fabricated. One acts as a modulator and the other as a demodulator. The proposed modulator and demodulator are fabricated on RTiDuroid 5880 with dielectric constant of 2.2 and substrate thickness of 10 mil. The Agilent HSCH-9251 GaAs Schottky bamer anti-parallel diodes are used for the suh-harmonic mixers. Agilent ADS 2001 is utilized to predict harmonic mixer performance and antenna characteristics.\nIn order to measure RF performance of the direct quadrature converter integrated with an antenna, the\ntransmitter system is located in the distance of 47 cm for far field measurement. Figure 3 shows the average conversion loss for I and Q channels as a function of LO power. The conversion loss is defined by the ratio of the RF power right before the microstrip patch antenna to the IF power. The conversion loss is lower than -15 dB from 10 dBm to 14 dBni of LO power. Figure 4 compares I with Q channel waveforms at an IF of I O MHz. It shows less than 1.2 O phase difference between I and Q channels in terms of a quadrature phase.\nFigure 5 shows the measured conversion losses as a iimction of the RF signal with LO power of 11.8 dBm. Within the frequency range of 39.75 GHz to 40.25 GHz, it has a 1 dB power imbalance between I and Q channels.\nMeasured phase comparison of I and Q channels at IF" + ] + }, + { + "image_filename": "designv6_24_0002572_978-3-319-76276-0_5-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002572_978-3-319-76276-0_5-Figure2-1.png", + "caption": "Fig. 2 Schematic representation of fatigue fracture surface", + "texts": [ + " The progression marks shows up which is because of the varieties in the load that brought about relating varieties in the crack growth rate. Eventually, the crack reaches the point where the remaining material gets overstressed, and the overload zone results. The progression marks indicate how the crack has developed and are just present in fractures where there have been generous varieties in the component stress as the crack develops over the piece [7]. The schematic representation of fatigue fracture surface showing crack origin and its progression is shown in Fig. 2. This section discusses the finite element analysis of helical suspension spring of rail vehicle using a numerical tool ANSYS to find its fatigue life. Analytically, the forces acting on suspension springs and shear stresses induced are calculated for static condition and continued to fatigue analysis for displacement variation of 6\u2013 8 mm as per the observation of rubbing marks over dampers. Fatigue analysis has been carried in ANSYS considering the load ratio, ultimate and endurance shear limit for chrome vanadium material" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002824_ias.2007.4347776-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002824_ias.2007.4347776-Figure1-1.png", + "caption": "Fig. 1. The structure of IPM rotor.", + "texts": [ + " For a precise computation of cogging torque by taking account of a magnetic saturation in stator teeth and rotor flux barrier, it is inevitable to employ numerical methods such as FEM and boundary element method (BEM). In this paper, however, the method of cogging torque is used to the space harmonic field analysis which is Fourier analysis of magnetization distribution of PM\u2019s and teeth distribution in stator core. By using the suggested method, this paper presents a cogging torque reduction method which is alteration of magnetization distribution. Fig 1 shows the structure of an IPM rotor, and an equivalent magnetization distribution of IPM motor is shown in Fig. 2, when the fringing and leakage effects of magnetic flux are considered. In the IPM motor, the effective magnetic pole angle is increased but residual flux density is decreased by leakage flux in rotor core. P U.S. Government work not protected by U.S. copyright 119 The energy function of magnetic field at airgap, when the stator does not have any slot, can be expressed as follows [4]: 2 0 10 ( )( ) sin ( ) 2 2k k BF X X kP\u03c6 \u03c0\u03c6 \u03c6 \u03bc \u221e = = = + + (1) { }0 2 4( ) sin( ) sin ( ) sin( ) ( ) eMB an n a b n bn \u03b8 \u03b8 \u03c0 \u2212 = \u2212 + (2) 2 0 0 0 ( ) (2 ) BX a\u03c0 \u03c0\u03bc = \u22c5 \u2212 (3) 2 0 0 ( sin ) ( )k BX ka k\u03c0\u03bc = \u22c5 \u2212 (4) where P is the number of poles" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001675_pemwa.2012.6316381-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001675_pemwa.2012.6316381-Figure2-1.png", + "caption": "Fig. 2. Magnetic flux density and flux density distribution of the PMSM.", + "texts": [ + " As a result of FEM, 72 slotted motors with winding structure PMSM and SPSM and same output power and speed also had almost the same shape and same dimensions of the stator slots, but had different number of turns as determined. A special form is given to the top of the SPSM\u2019s rotor pole to be able to obtain the sinus form in the air-gap. The geometrical parameters of rotor and stator are given in Table 4. The two models are compared regarding the flux density and its distribution under maximum torque conditions. Fig. 1 depicts the magnetic flux lines and flux density for the PM machine while Fig. 2 provides the same result for the salient rotor machine. The comparison of the field calculation shows that both models are in quite good agreement and the main difference is that in the PM machine, the flux density distribution is less affected by machine load. The flux distribution of the rotor completes its path properly (without breaking in the air-gap) through the stator and air-gap. This means that the air-gap spaces of the designed machines are in appropriate values. Also, from the performance curves and tables of the machines it can be deduced that the air-gap values are in an acceptable range" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003283_870182-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003283_870182-Figure1-1.png", + "caption": "Figure 1", + "texts": [ + " It performs as we II or better than other systems and Is also less complicated to manufacture. ThIs system, named the '1HSI<\" method, can be used on copper/brass cores as well as alumInum cores. This paper will descr 1be the funct i ona 1 and rna nu factur Tog advantages of the \"HSKII plastic tank attachment method. Bent Tab Crimp Strip Slotted/Dimpled Ne:rflx 1. BACKGROUND The tank to header joInt Is one of the most crItical joints on a plastic tank radiator (either copper/brass or alumInum). Figure 1 illustrates the most commonly used method tor makl ng th 1s jo i nt, the Tabbed Header. Figure 2 TABBED HEADER Plastic Tank Attachment Method Other methods cu rrentl y be I ng used for the tank to header joint are: separate crlmp strlps, Nertex strips, and slotted/dimpled confIgurations (see Flgure 2). AII of these methods re I y on the tabs or flange All the structural loads at the tank to header jolnt are resIsted by the crlmped or bent connectlng sectlon. The loads are caused by lnternal fluld and gasket pressures, thermal expanslon/contractlon, and vlbratlon" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000083_ir-03-2018-0042-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000083_ir-03-2018-0042-Figure2-1.png", + "caption": "Figure 2 3 D assembly drawing of the wearable vehicle in fast motion mode and walking mode", + "texts": [ + " When facing obstacles such as rough terrains or stairs, a wearer simply stands up and walks as shown in Figure 1B. The exoskeleton, in this walking mode, carries the loads needed to be transported, as well as the motorized wheels. The seat and the free wheels are set in folding configurations. When the human reaches rough surface or new flat free ground, they sit down again and repeat the fast transportationmode. The wearable vehicle has two modes of motion which are fast motion mode and walking mode as shown in Figure 2. Walking mode comprises a trunk mechanism, backpack mechanism, two powered anthropomorphic legs which have hip, knee and ankle joints. Fast motion mode operates via rear wheels mechanism, front mechanism and foldable seat. In fast motion mode, the subsystems of the walking mode are act as the frame of the vehicle. The hip joint has three DOFs and the knee joint has one DOFs, while the ankle joint has three DOFs. Three motors are used to generate the provided motion in the sagittal plane. The range of joint angles is mechanically restricted to avoid any potential danger to the user due to system failure as given in Table I" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002367_iros.2009.5354127-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002367_iros.2009.5354127-Figure1-1.png", + "caption": "Fig. 1 Structure of 3-DOF active rotational ball joint", + "texts": [ + " In this research, the objective is the development of a structurally simple and compact 3-DOF active rotational ball joint, which can realize a rotation around an any-direction rotational axis with a well manipulability and can change its direction of rotational axis smoothly and arbitrarily when it is on rotating. We present and elucidate the construction principles behind the 3-DOF active rotational ball joint, and we demonstrate the effectiveness of the proposed principles by means of motion analysis and practical experiments. A. Construction principles Fig.1 shows a model of the construction of joint device proposed in this research, which is a 3-DOF active rotational ball joint between a pair of links. The construction principles of the device arrange the rotational axes of each of the three hollow shaft motors supplying the drive power orthogonally for active rotations. The axes of active rotation are denoted as Xm, Ym and Zm 978-1-4244-3804-4/09/$25.00 \u00a92009 IEEE 5153 which are at right angles one another. The power output for the Xm axis, the Ym axis and the Zm axis are compounded, and aligned at the centre of the joint in the manner of a ball joint" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002010_jae-2008-999-Figure18-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002010_jae-2008-999-Figure18-1.png", + "caption": "Fig. 18. Prototyped vibrator of the proposed RUSM.", + "texts": [ + " As shown in Fig. 16, when speed arrived at the steady state, the speed of motor was 816[rpm]. This result also validated the proposed characteristic analysis and design method. Important to note in transient analysis result is that only it takes under 1[msec] to arrive at the steady state. Figure 17 indicates the speed and the efficiency of the proposed 8.5[mm] diameter RUSM about the varying torque of the motor when 14.14[Vrms] was applied. The characteristics of this motor are tabulated in Table 3. Figure 18 shows the bottom side of the prototyped vibrator that displays the electrode pattern of the prototyped piezoelectric ceramic of the 8.5[mm] diameter RUSM. The piezoelectric ceramic was made from the Dong-Il and the comb-tooth structure was made from the phosphor bronze. The phosphor bronze had good properties for the vibrator of a RUSM, such as superior anti-abrasion, nearly uniform young\u2019s modulus to temperature change. Material coefficients are in the Appendix. Figure 19 indicates the prototyped comb-tooth structures" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000664_9781119238386.ch5-Figure5.18-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000664_9781119238386.ch5-Figure5.18-1.png", + "caption": "Figure 5.18 (a) Physical layout. (b) Top/bottom-view photographs of the proposed filter. (W12 = 0.5mm,L12 = 7.44mm,L12 = 7.76mm, S12 = 1.68mm,W1 = 0.45mm, S1 = 0.23mm, L1 = 8.4mm,WS = 3.75mm,LS = 7.8mm,W2 = 10mm,W2 = 4.35mm, L2 = 16.3mm,L2 = 10.59mm,WA= 0.3mm, LA1 = 1.85mm, LA2 =LA3 =LA4 = 1.7mm,LA5 = 0.85mm,WB = 2.25mm,LB = 5.46mm). Reprinted with permission from Ref. [8]; copyright 2016 IEEE.", + "texts": [ + "57 GHz, respectively. So simulated frequency ratios with respect to 6.35 GHz are 0.581/1.419 and 0.493/1.507, which are equal to the computed ones and indicate the controllability of CM TZs. Therefore, with appropriate impedance in internal cross-coupling and 3\u03bb/4 SIRs, enhanced CM rejection with 5 TZs is finally gained. For demonstration, this balanced filter is optimized and fabricated on a substrate with a relative dielectric constant of 2.55, a thickness of 0.8 mm, and a loss tangent of 0.0029. Figure 5.18 shows the physical layout with all dimensions provided and photograph of the fabricated filter. The comparisons between the results of EM simulation and the measurement are shown in Figure 5.19, where solid line and dot line represent the EM simulated and measured results, respectively. In EM simulated (measured) result, a wide DM passband with 3 dB fractional bandwidth of 58% (59%) and a wide CM stopband with fractional bandwidth of 123% (118%) more than 15.5 dB (18.2 dB) are simultaneously attained" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001397_icitacee.2016.7892413-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001397_icitacee.2016.7892413-Figure3-1.png", + "caption": "Fig. 3 Linear ball screw calculation", + "texts": [ + " There are several factors that determine measures on gears can be show in Table 1 as follows: Linear ball screw actuator is a mechanical equipment which uses the rotational motion of the moving objects with minimal friction. Screw on the shaft serves as groove ball bearing so that the removal can be done with precision / high-precision position that can be shown in Figure 2. The Figure 2 shows that the linear ball screw was installed directly with Smartmotor with specific gear box. The design linear lead screw can be seen in Figure 3. The calculation for the soft contact lens inspection can be determined by (1)-(5). SmartMotor is a servo motor in which it is equipped by a servo control system. In a Smartmotor already include servo motor controller, amplifier and encoder. Things that are required to operate a Smartmotor is the power supply, internal programming and serial communication command from the outside (or both). The control of the motor use specific software that was dedicated for the motor. The Proportional, Integral, and Derivative (PID) controller was be used in the system" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002870_1.5122106-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002870_1.5122106-Figure1-1.png", + "caption": "FIGURE 1. Principle construction schemes of ball valve: (a) turned ball scheme; (b) stamped ball scheme", + "texts": [ + " Therefore, the main tasks of the research are: the creation of a mathematical apparatus that describes the flow of liquid and gas in the working cavity of ball valves of stamped and turned type; computational research of fluid flow at various angles of the ball rotation and determining the dependence of flow resistance on the rotation angle; comparison of flow resistances of ball valves with turned and stamped balls as well as with experimental data for the fluid flow in similar structures. According to the tasks set, mathematical modeling of the fluid flow in ball valve was carried out for two principle construction schemes shown in Fig. 1, where D0 is the internal diameter of the pipeline, \u03b4 is the rotation angle of the ball. In scheme No. 1 (Fig. 1, a) a ball valve with a cylindrical opening area (turned ball) is considered, in scheme No. 2 (Fig. 1, b) a more manufacturable design with a spherical opening area (stamped ball) is presented. In order to create the most general mathematical model, the gas medium is considered as a fluid. Mathematical model is based on the Navier-Stokes system of equations. The model is taken as the turbulence model. It is based on the turbulent viscosity conception. The fluid is the ideal gas. Below is the complete system of equations implemented in the ANSYS CFX software package [16]. 1. Continuity equation: 0U (1) where is the mixture density; ( ) is the velocity vector in the Cartesian coordinate system; is the Hamiltonian vector divergence operator", + " The values of and are determined directly from the differential transfer equations for the turbulent kinetic energy and the dissipation of the turbulence velocity: \u00a0t k k k U k k P t (5) 1 2\u2022 \u00a0t kU C P C t k (6) where , , and are the turbulence model constants [16]; is the turbulence generation due to viscous and Archimedes buoyant forces, 2 \u2022 \u2022 3 3 T k t t kbP U U U U U k P (7) where is the generation of turbulence, for the calculation of which the Boussinesq model is chosen: t kbP g T (8) where is the constant [16]; is the thermal expansion coefficient. The problem is considered in the stationary setting. It is assumed that heat sources due to viscous dissipation and radiation in the flow are significantly less than convective and diffusion components (effective turbulent numbers ). The internal part of the pipeline, shown in Fig. 1 (a) and 1 (b), is taken as the region of computation. The developed mathematical model is complemented by boundary conditions. When simulating workflows in a ball valve, the static pressure was set in the inlet cross-section, the mass flow rate of gas \u0307 was set in the output cross-section, the wall was assumed to be absolutely smooth, the gas velocity at the wall m/s. The computational analysis is carried out on the basis of the finite element method after creation of the mesh for the model. When creating the mesh in the region of computation, the tetrahedron was chosen as the finite element, 030056-3 the minimum and maximum block sizes are 0.1 mm and 10 mm, respectively, the maximum block size on the surface is taken to be 10 mm. At that, for scheme No. 1 (Fig. 1, (a)) the total number of elements was about 600 thousand; for scheme No. 2 (Fig. 1, (b)), the total number of elements was about 650 thousand. A numerical study of the working processes in a ball valve was carried out for a full-bore valve used in the chemical industry with a high gas flow: 0.2 m, \u0307 5 kg/s with 1.5 and 2.5 MPa in the angular range 10\u00b0\u202675\u00b0. During computational research of the working process in a ball valve, the gas pressure and velocity fields were obtained for two computational schemes (Fig. 2 and 3). Let us consider the results of numerical simulation in more detail" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000941_j.aasci.2017.12.010-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000941_j.aasci.2017.12.010-Figure1-1.png", + "caption": "Fig. 1. Diagram of the interaction of a driving wheel with the soil.", + "texts": [ + " on behalf of Agricultural Unive traction dynamics, maneuverability and possibility of the existing and newly designed tractors and other wheeled power machines needs a further development of the theory of interaction of the tractor driving wheels with the soil by using the fourth approach, particularly by considering the rheological properties of the soils. When considering the interaction of an elastic driving wheel with the soil, an assumption implying that under a steady-state regime, a tangential force Fk of the wheel traction equals to the sum of the soil tangential reactions directed towards themovement (Fig. 1). In order to draft an equivalent design model, we proceed from the consideration that during the interaction of a driving wheel with the soil, there act the traction forces between the support surface of the tire and soil, in particular, the forces arising at cutting the ground brick with the side edges of the lugs. The work [2] proves that for loose grounds the value of shear and cutting forces increases and is frequently a determining one. As a driving wheel moves, its lugs shift and cut the ground in the direction opposite to their movement. The support of the lugs in the ground, shift and cutting of bricks, constrained between them are possible only if the traction forces are thoroughly used, i.e. in case of a wheel ship. It is considered [2] that in case of steady motion of a wheel rsity of Georgia. This is an open access article under the CC BY-NC-ND license (http:// (V \u00bc const), a shift and cut of ground bricks take place basically in the periods of the egress of the last lug of the wheel support surface out of the ground (Fig. 1), and at this moment the load is distributed over the remained lugs, which are geared. All lugs move and cut the ground by equal values Di, while the first lug shifts the ground by Di, the second lug shifts the ground by 2Di, the third lug shifts the ground by 3Di And so on. As the first lug passes all stages of gearing from the ingress into the ground to the egress from the ground, the maximum shift and cutting of the ground at the egress from the gearing Dmax \u00bc nDi (here n is the number of lugs in the gearing of the support surface with the ground)", + " It is proved [2] that the distribution of deformation of shear and cutting of ground bricks in contact with the support surface of a wheel with the ground may be presented as a triangle, and its maximum value may be presented as the product of the slipping coefficient d and length of the support surface of a wheel L, i.e. Dmax \u00bc dL. Under the impact of the lugs, there arise shear stresses tshi, which increase at the beginning and reach their maximum value tmax at shifting the ground by D0, then they decrease and reach a constant value tcut at full cutting of the ground brick. The ground deformation at distance x from the start of gearing equals to Dmax \u00bc dx (Fig. 1). The tangential traction force necessary to overcome the ground shift is as follows: FKcut \u00bc ZL 0 t\u0445dA ZL 0 bt\u0445d\u0445; (1) here dA is an elementary area of the support surface of a wheel, equaling to dA \u00bc bdx; b is the width of the wheel lug; b \u00bc 2l sin b (here l is the length of the side edge of a lug; b is the angle characterizing the position of a lug on the wheel (Fig. 1); at b \u00bc 900, b \u00bc 2l); dx is the length of the elementary area. In a general case, tx depends on normal pressure Px of the ground deformation, its physical-mechanical properties and wheel parameters, i.e. tx \u00bc f \u00f0PxDx\u00de. In order to establish the given dependence, different empirical and semi-empirical formulae are used, including the functional dependence given by V.V. Katsigin [3]. In the given work, by using the methodology [2] of deducing a design equation of the tangential traction force, a more general approach is assumed", + " At very slow processes of deformation, equation (2) with the speeds of dt dt and dg dt can be neglected in relation to the values of t and g, and thenwe arrive at a commonHooke's lawwith a longitudinal shear modulus G\u221eg \u00bc t, and on the contrary e in case of very rapid processes of deformation, the speeds of deformation and stresses are very great and the deformations and stresses may be neglected in comparison with them. At the same time, we arrive at Hooke's Law again, but it is differentiated in time and has an instantaneous modulus of shear G0 dg dt \u00bc dt dt. We used the dependence [2] to deduce the formulae of the tangential force of wheel traction at the ground shift and traction coefficient. In order to reduce the derivation, as it is commonly accepted, let us change the support surface with a big curvature radius (within the limits of the contact with the ground) by a horizontal surface (Fig. 1) with the length of contact L. At the same time, let us assume that the shear reactions are parallel to the given plane and the wheel moves in the steady-state regime. Let us determine the relative shear deformation and speed of the distribution of shift deformations, included in the original equation (2). If under the impact of lugs the soil is shifted by valueD the relative shear resistance will be g \u00bc D H0 (whereH0 Is the depth of deformation, m). Fig. 1 shows that at distance x, the shear resistance is Dx \u00bc dx (with d as a slipping coefficient). Then, at distance x, we will have g \u00bc dx H0 . Let us express x by an actual speed x \u00bc ydt \u00bc yT \u00f01 d\u00det. Then we have g \u00bc dydt H0 ; dg dt \u00bc dy H0 . Based on the foregoing, the original equation (2) will be as follows: t\u00fe Tr dt dt \u00bc G0Tr dyd H0 \u00fe G\u221e dydt H0 : (3) And its solution will be as follows: t \u00bc G\u221e dydt H0 \u00fe t0e t Tr \u00fe Tr\u00f0G0 G\u221e\u00de dyd H0 Tr\u00f0G0 G\u221e\u00de dyd H0 e t Tr ; (4) where t0 Is the value of the initial tangential stress" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000485_icorr.2013.6650440-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000485_icorr.2013.6650440-Figure2-1.png", + "caption": "Fig. 2. Illustrates how the Scotch-Yoke mechanism moves hip from fully extended (1) to fully flexed (3). See Fig. 1 for alignment with gait cycle.", + "texts": [ + " To realize a variable transmission that changes in a sinusoidal manner the scotch-yoke mechanism was used. Scotch-yoke mechanisms convert continuous rotary motion into reciprocating linear motion and have been used in valve control applications, internal combustion and steam engines [10-11]. These mechanisms are similar to a crank and slider mechanism in that the linear output moves in a sinusoidal pattern except scotch-yoke mechanisms have fewer moving parts and are capable of higher torque output. Figure 2 shows a full cycle of this mechanism and it starts with the slider all the way to the right. As the wheel rotates counterclockwise the roller attached to the wheel pushes the slider to the left. After the wheel has rotated 180 degrees the slider reaches the end of its range of motion and as the wheel continues to rotate the slider begins to move toward the right. This reciprocating motion continues as long as the wheel continues to turn. When the wheel rotates at a constant velocity the slider reaches its maximum velocities when the wheel is at top dead center and bottom dead center" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000811_tpel.2007.909208-Figure12-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000811_tpel.2007.909208-Figure12-1.png", + "caption": "Fig. 12. Synthesis of the required output voltage vector in sector 1.", + "texts": [ + " The specification of 1 kW IPM motor under test is presented in Table I. Real time control software was coded using C language. The voltage space vector output from variable structure controller is realized through symmetric space vector PWM with lower THD. The voltage vector is synthesized by the two neighbouring active vectors in the same sector where it sits. Space vector PWM signals are generated on the DS1102 board through the TMS320P14 slave DSP. It can be noted that the maximum range of the VSI inverter output is defined by the hexagon as shown in Fig. 12. However, the requirement of sinusoidal output voltages restricts the voltages within the inscribed circle of the hexagon. The chattering free torque and flux dynamics are illustrated in Fig. 13 when the toque command reverses between 2 N.m. The corresponding trajectories of speed, current and flux are in Figs. 14 and 15. Under the same conditions, the steady state performance of the VS\u2013DTC and PI\u2013DTC has been examined. Comparison of steady state performance between PI\u2013DTC and VS\u2013DTC proves that VS\u2013DTC has higher performance" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001212_6.2008-5955-Figure17-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001212_6.2008-5955-Figure17-1.png", + "caption": "Figure 17. Regional Jet Parametric Geometry", + "texts": [ + " The vehicle was chosen to perform the same mission and have the same configuration characteristics as the Boeing 747-100 with a mission profile as shown on Figure 16. Thrust loading and wing loading of the vehicle are set to equal 0.25 and 132 lb/ft2 respectively. 17 of 23 As described before, aircraft geometric information is hierarchically modelled with support from the parametric geometry modeller that provides a unified geometric description to all disciplines. The aircraft geometry used in the validation analyses as developed by the parametric modeller is shown on Figure 17, Figure 18 and Figure 19 respectively. Comparison of the primary sizing results from the new design environment and the real aircraft data for the evaluated examples is shown on Table 1, Table 2 and Table 3. The implemented design environment correlated well with key aircraft parameters, not only in terms of weight but also in estimated performance such as takeoff and landing field lengths. 18 of 23 19 of 23 Note as well in Table 3 how the different aerodynamic methods used for drag buildup provide good correlation with published aerodynamic data" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000370_ijmee.27.1.1-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000370_ijmee.27.1.1-Figure1-1.png", + "caption": "Fig. 1. Simplified model of a gear transmission system. (a) Gears in forward contact. (b) Gears separated (no contact). (c) Gears in reverse contact.", + "texts": [ + " The journal collars were prepared at CARLETON UNIV on June 26, 2015ijj.sagepub.comDownloaded from International Journal of Mechanical Engineering Education Vol 27 No 1 individually and appropriate machining, cleaning and polishing procedures were used to secure the desired surface properties. The roughness of the test bearing and journal sleeves were monitored with a profilometer both in the normal and parallel directions of sliding. All test materials were cleaned with soap, water and alcohol after each experiment to secure that all loose material was removed. Fig. 1 depicts a lump mass model of a machine of a typical single-stage spur gear physical transmission system, consisting of a motor connected by a flexible shaft to a pair of involute spur gears, which are connected to a load by a second flexible shaft. The whole system is supported on four bearings. The mathematical model may be represented by masses, springs, and dampers as shown in Fig. 2. The equations of motion for the system are expressed as follows: J C K W R1 1 1 1 2 1 1 2 1 1 \u02d9\u0307 \u02d9 \u02d9 \u02d9\u03b8 \u03b8 \u03b8 \u03b8 \u03b8 \u00b5 \u03b8+ \u2212( ) + \u2212( ) = ( ) (1) J C K P f m W R c 2 2 1 1 2 1 1 2 2 1 2 1\u02d9\u0307 \u02d9 \u02d9 ( ) \u02d9\u03b8 \u03b8 \u03b8 \u03b8 \u03b8 \u00b5 \u03b8+ \u2212( ) + \u2212( ) + + = \u2212 ( ) (2) J C K P f m W R c 3 3 3 3 4 3 3 4 3 3 3 1\u02d9\u0307 \u02d9 \u02d9 ( ) \u02d9\u03b8 \u03b8 \u03b8 \u03b8 \u03b8 \u00b5 \u03b8+ \u2212( ) + \u2212( ) + + = \u2212 ( ) (3) J C K W R4 4 3 3 4 3 3 4 4 4 \u02d9\u0307 \u02d9 \u02d9 \u02d9\u03b8 \u03b8 \u03b8 \u03b8 \u03b8 \u00b5 \u03b8+ \u2212( ) + \u2212( ) = ( ) (4) where J1, J2, J3, J4 are the mass moments of inertia of the load, the gears and the motor under consideration" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002862_detc2007-34856-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002862_detc2007-34856-Figure3-1.png", + "caption": "Figure 3: Metal Belt [3]", + "texts": [ + " While the CVT is designed such that the belt should not slip, some slipping will occur, especially at higher torques. This tendency to slip at higher torques greatly limits the ability of this type of CVT to be adapted to higher torque applications. 1 Copyright \u00a9 2007 by ASME erms of Use: http://www.asme.org/about-asme/terms-of-use D In response to the need for higher torque transfer in automotive CVT applications, companies such as Nissan (see Fig. 2), Honda and Toyota have developed v-belt CVT\u2019s utilizing metal belts, as shown in Fig. 3. The metal belt designs can transmit more torque than the composite belts, but are likewise limited in the amount of torque they can transfer before slipping occurs. The need for higher torque capabilities coupled with the desire for the efficiency gains capable with CVT\u2019s has prompted the desire to develop a CVT that does not rely on friction. ownloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 02/04/2016 Te Current development of CVT\u2019s can be classified in the following families: hydraulic systems, systems based on belt transmissions (Transmatic Van Doorne system, Fouillarion\u2019s system, the Kumm mechanism, the PIV-Reimers, Variomatic), systems based on wheels in contact (Nu-Vinci, Vadetec NTXA2, Hayes CVT toroidal, Torotrak), and oscillating systems [4,5]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001187_intmag.2003.1230623-FigureI-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001187_intmag.2003.1230623-FigureI-1.png", + "caption": "Fig. I , 4-pole, 1 phase slot-less permanent (magnet motor topologies: (a)Radial magneliied rotor (b)Halbnch mamcvzed rulur.", + "texts": [ + "xternal forces, manufacturing imprecision and bearing effects induce rotor eccentricity and this causes vihratinn and noise of motor 1 IJ121. Using the linile element analysis, this paper analyLes unbalance force according to rotor eccentricity and deals with the comparisons of unbalance force of two types, radial and Halbach magnetized rolors of high spccd slo~kss P M machinc. Figure I shows a 4-pole, 3-phase slot-less permanent magnet motor, with (a) conventional radial magnetized permanen1 magnets. and (h) a multi-pole Halhilch inagnetired miigneti. 'The lopologicr were designed as I kW permanent m~~giiet molor/gencrator wilh Ihc raled speed 0140.000 rpnl. Figure 2 shows the geometric configuration of permanent magnet machine with rotor eccentricity and unbalance force fluctuation to be analyzed. When pernunent magnet rotor rotates around the axis through 0, at the angular velocity m,, Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002138_energycon.2010.5771763-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002138_energycon.2010.5771763-Figure8-1.png", + "caption": "Fig. 8: Control of the SCs discharging", + "texts": [ + " First of all, the voltage value of the DC stage must be of 150.5V; a hysteretic regulator has been used to verify this condition. with: I2MIN=22 A and I2MAX=23 A. Then, the SCs are charged under a constant current (22.5A) ensured also by a hysteresis regulator with 1A of bandwidth. In order to stop the charge of the SCs when its voltage reached 10V, we added a condition on the control of the SCs voltage. The SC current reference is equal to zero when the SC voltage is greater than VSCmax. The Fig. 8 chows the control scheme of SCs discharging. From the power needed by the load and dividing by the SCs voltage, the discharging SCs reference is generated. The simulations have been achieved by using the Saber software. In this section we present the simulation results highlighting the different functionalities of the studied converter. We are interested in the current and voltage waveforms in the most important points of the converter. The Fig. 9 proofs that the proposed converter really absorbs a sinusoidal current from the AC network" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003025_pi-b-1.1957.0157-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003025_pi-b-1.1957.0157-Figure6-1.png", + "caption": "Fig. 6.\u2014Planar diode: hole distribution in the /i-region at the end of the period of constant current.", + "texts": [ + " 5 shows how V \u2014 Vf changes with time; the curves are similar to some given by Kingston. During Tx (the normalized period for which / = Jr), the polarity of the voltage across the diode may not have reversed although a reverse current \u2014ir is flowing. Eqn. (13) applies until px^0->0, at 7\", (= /,/TP) where Tx is given by At T\\ a new set of boundary conditions applies one of which is that/?x-0 = 0 for all subsequent times. Another, the distribution of holes at Tlt is given by substituting Tt for T in eqn. (13). The curves labelled PTx ( = p/p0) of Fig. 6 show the distributions for two representative values of Tx. There is a simple relationship between a chosen value of Tx and the slope, at the origin, of the curve appropriate to it, because Jr \\Jr\\Jf)if \u201e o,T~Tl~eDp~ AQeDp and Tx and JrjJf are related as in eqn. (14). The solution of the continuity equation with the new boundary conditions is given in the Appendix (Section 8), together with comments on an earlier, less rigid solution. For large values of T (measured from the end of the period 71,), it is shown that J -7,X - (r, + \u2022 ' + ", + " He assumed that / decays, subsequent to Tx, as if it had started from infinity (i.e. Ro had been zero, and hence the conditions and analysis of Section 2.1.1.1 applied) and had reached Jr at some time V measured from the time of instantaneous reversal of the diode voltage which would have to be (Jx - T') after the beginning of Tx. He assumes, therefore, that not only can \\ 1 = \\(*L\\ ) A>o_k, T, L A * * ' * - o, r but also that implying that the distributions PT>T,, and PT>T- are equal at and after Tx and 7\", respectively, for all values of X. The two curves marked PT> in Fig. 6 show, by comparison with the more accurate curves marked PTl, the extent of his approximation for the two values of Tx chosen earlier to illustrate PTl. Although the two quantities P and dP/dX are equal for the two curves of each pair at X'\u2022= 0, they differ for all the other relevant values of X; the differences are reflected in different values of / at all subsequent times, as will be shown by the following analysis. The boundary conditions to the continuity equation are now = e-x _ from eqn. (13), p = 0 at X = 0 for all values of T (now measured from T, as origin) p = 0 at X = oo for all values of T" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000855_0954406220936732-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000855_0954406220936732-Figure2-1.png", + "caption": "Figure 2. Coordinates of the flexible link.", + "texts": [ + " Stiffness modeling of the hybrid limbs parallel force sensor Stiffness model of flexible link element and coordinate transformation The stiffness of the flexible link element can refer to the form of cantilever beam with rectangular cross section and the flexibility matrix determined by the parameters of the flexible beam is obtained20,21 where L and A are the length and cross-sectional area of the flexible links, E and G are the Young\u2019s modulus and shear modulus of materials; Ip is the polar moment of inertia of the flexible beam, Iy and Iz are the moment of inertia of flexible beam, which can be Cg\u00bc L=EA 0 0 0 0 0 0 L3= 3EIz\u00f0 \u00de 0 0 0 L2= 2EIz\u00f0 \u00de 0 0 L3= 3EIy 0 L2= 2EIy 0 0 0 0 L= GIp 0 0 0 0 L2= 2EIy 0 L= EIy 0 0 L2= 2EIz\u00f0 \u00de 0 0 0 L= EIz\u00f0 \u00de 2 6666664 3 7777775 \u00f01\u00de calculated as follows, respectively Iz \u00bc bh3 12 , Iy \u00bc b3h 12 \u00f02\u00de The flexibility matrix of flexible link is established in its local coordinate system {Og} of which the xg direction along the axis of the link and its direction follows Figure 2. However, the flexibility matrix needs to be expressed in another reference coordinates {Op} whose x-axis coincides with that of moving platform. The reference coordinate system {Op} is established at the geometric center of the free end of the link. The xp-axis is perpendicular to the upper end face of the link. The deformation and force transformation relation of flexible link in reference coordinate system and local coordinate system can be derived from the following formulae Xp \u00bc JXg \u00f03\u00de Fg \u00bc JTFp \u00f04\u00de where J is the pose transformation matrix as follows J \u00bc p gR O3 3 O3 3 p gR \" # \u00bc Rot\u00f0zp, \u2019\u00de O3 3 O3 3 Rot\u00f0zp, \u2019\u00de \u00f05\u00de Based on equations (1) and (5), the flexibility matrix of the flexible link can be obtained at the reference coordinate system {Op}" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000124_aps.2016.7696396-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000124_aps.2016.7696396-Figure2-1.png", + "caption": "Fig. 2, The Radar Instrument Structure (RIS) with the six L-Band Tiles on the top deck. The radomes of the L-Band tiles in this view are semi-transparent. Also visible is the base of the reflector boom at one end of the array. The TRMs are located on both sides of the RIS.", + "texts": [ + " A stripline feeding network made in Rogers 6002 provides the proper power splitting and phasing to feed the patch pairs. The patches, organized in a stacked patch configuration to widen the bandwidth, are supported by the feed probes plus a center post for structural rigidity. There is no dielectric under the patches. The lower patch is capacitively fed, which enables proper tuning of the patch. The top patch is directly soldered to the probes. A center post adds structural rigidity, and has the side benefit of naturally rejecting the second harmonic that is generated by the Transmit/Receive Modules (TRMs). Fig. 2 shows the main body of the Radar Instrument Structure (RIS), including the top deck with both the L-Band and the S-Band feed arrays. The L-Band and S-Band TRMs are located both inside and outside the RIS. Each L-Band tile is covered by a radome. The radome shell is made of a 1mm thick Astroquartz layer and is painted for thermal control and protection from atomic oxygen corrosion. Inside the radome shell, a foam layer 1379978-1-5090-2886-3/16/$31.00 \u00a92016 IEEE AP-S 2016 with cut-outs for the patches provides additional thermal insulation" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003858_tmag.2006.872493-Figure6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003858_tmag.2006.872493-Figure6-1.png", + "caption": "Fig. 6. Model of analysis. (a) Whole view. (b) Analyzed region.", + "texts": [ + " When all sheets are subdivided into a fine mesh in a model of more laminations ( mm), the CPU time and the number of unknown increase greatly. When four sheets in the surface side are subdivided into a fine mesh the increase in CPU time and the number of unknowns is not remarkable. Therefore, the proposed modeling is very effective in an actual model composing many sheets in terms of CPU time and the number of unknowns. The validity of the proposed modeling method is examined by comparing the calculated results and measurements of local loss on the surface of core. Fig. 6 shows the model reactor [6] used for experiment and analysis. Fig. 6(b) shows the analyzed region ( of the whole region). It is a serial reactor for capacitor facilities for 3-phase, 60 Hz and 100 kVAR. The core is made of grain-oriented silicon steel 35 G165. Since the rating capacity is small, there is just one air gap at the center of the core leg. The air gap length between core blocks is equal to 6.7 mm. The average flux density of the core block is assumed to be 1.10 T. A thermocouple method is employed to measure the local loss on the surface of the core. Thermocouples are installed in five points at intervals of 1 mm along the z-axis from the gap ( mm, mm)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002921_iecon.2011.6119447-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002921_iecon.2011.6119447-Figure5-1.png", + "caption": "Fig. 5. Rotor of a squirrel-cage induction motor under test \u2013 thermocouples mapping.", + "texts": [ + "2 mm and are fixated in the rotor with cement glue. The calibration of the sensors together with the instrumentation electronics is carried out at 50 \u00b0C using a high precision environmental chamber. The thermal sensors embedded into the stators winding are of type KTY/130. Both motor are driven by PWM controlled inverter. The heat-up tests are carried out under nominal load. A. Induction Motor The parameters of the induction motor are listed in Table I. The thermocouples mapping in the rotor of the induction motor is depicted in Fig. 5 and summarized as follows: Thermocouple 1 (tc1) and 2 (tc2) are radial surfaced mounted on one end of the rotor. Thermocouple 3 (tc3) and 4 (tc4) are radial surfaced mounted on the other end of the rotor Thermocouple 5 (tc5) is buried in a hole of 2 mm diameter drilled in the rotor with a depth of 5 mm Thermocouple 6 (tc6) is buried in a hole of 2 mm diameter drilled in the rotor with a depth of 15 mm The motor is operated at nominal speed and cooled by a fan. After approximately 1.5 hours the motor is switched off together with the fan" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002565_iecon.2016.7793740-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002565_iecon.2016.7793740-Figure1-1.png", + "caption": "Fig. 1. Dq1dq3 phasor graph of five phase PMSM", + "texts": [ + " Similar to Park transformation in three-phase systems, it is possible to transfer the electrical parameters of a five phase motor into two rotating reference frames (namely 11qd and 33qd ) and one homopolar axis [7]. Derivation from five phases can be found in [8]. Magnetizing inductances are a subject for commissioning or they can be evaluated by analytical expressions, reluctance network or finite element (FEM) methods. A quick and precise numerical procedure which can estimate the 11qd and 33qd inductances by winding function is reported in [9]. 978-1-5090-3474-1/16/$31.00 \u00a92016 IEEE 2904 The phasor graph of five phase PMSM is depicted in figure 1. The stator voltage equations are given by dt di LiRU iL dt di LiRU iL dt di LiRU iL dt di LiRU iL dt di LiRU s PMddr q qqsq qqr d ddsd PMddr q qqsq qqr d ddsd 0 000 333 3 333 33 3 333 111 1 111 11 1 111 3 3 where sR is a stator resistance, 3311 ,,, qdqd iiii are stator currents of the first and third harmonics, 3311 ,,, qdqd LLLL are inductances of the corresponding axes and 00 , LR are resistance and inductance of zero sequence component. Permanent magnet components are denoted as 1PM and 3PM " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001401_detc2017-68370-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001401_detc2017-68370-Figure7-1.png", + "caption": "FIGURE 7: ANSYS model of a 2 x 10 Outside LET Joint array.", + "texts": [ + " The resistance of the tab or end piece is: Rtab = \u03c1Ltab wtabt (15) The resistance of one of the torsional legs is: Rt = \u03c1Lt wtt (16) The resistance of one of the bending legs is: Rb = \u03c1Lb wbt (17) The total resistance for a single joint, ignoring the upper tab, gives: Req = Rtab + 2Rt +Rb 2 (18) The resistance of the total array will then be: Rarray = nReq +Rtab m (19) If the flexible material used is an insulator and there is a conductive material placed along the surface, the thickness in these equations can be switched to the thickness of the conductive material. Equation 19 shows that the resistance of an array is easily related to the resistance of a single surrogate fold. Equations 1- 8 give an example of calculating the resistance of a single joint, while Equations 15-19 show how to calculate the resistance of an array based off the number of folds in parallel and in series. To validate the assumptions in the analytical models, FEA models were created in ANSYS (see Fig. 7). The arrays were modeled as Solid186 elements with Beam188 elements being coupled to the array in order to apply the desired rotation of 180 degrees. This large deflection necessitated using a nonlinear solver in order to achieve accurate results. The von Mises stress was retrieved to compare to the analytical models. The margin of error between the FEA and analytical models for different sized arrays was minimal (see Fig. 8). As can be seen, the error for all the joints was less than 3%, which is acceptable when compared to the error introduced through the variability in manufacturing and the materials" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000223_1.4030653-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000223_1.4030653-Figure7-1.png", + "caption": "Fig. 7 (a) Fixed-guided bending of PCCP and (b) PRB model of PCCP", + "texts": [ + " The arc length difference between upper and lower plate during deformation allows PCCP exoskeleton to provide increased and customized force profile to the user by combining with PEB spring which is introduced in the PEB Spring section. Upper and lower plates of PCCP are fixed to sliding and fixed portion of PEB spring separately as shown in Fig. 3. Hence, the arc length difference provides the PEB spring extension which is then translated into a corresponding force profile. Fixed-Guided Bending Mode. The stiffness of PCCP increases discontinuously and drastically when the sliding pins are limited by sliding slots as shown in Fig. 7. This fixedguided mode can be modeled by a PRB model which one end of beam is fixed while the other is guided with fixed angle. To maintain the constant angle, a resultant moment has to be present at the end. The length of unit segment of PCCP prototype is 33 mm and overall assembly consists of eight unit segments. Those are distributed symmetrically from the center of PCCP as shown in Fig. 3. For each of two springs, the torsional spring constant, K, is twice as stiff as for the case of cantilever beam and there are two springs for the PRB model" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001290_j.1460-2695.1995.tb00839.x-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001290_j.1460-2695.1995.tb00839.x-Figure2-1.png", + "caption": "Fig. 2. Finite element models used in the analysis of homogeneous and cladded specimens.", + "texts": [ + " The materials were assumed to be elastic-plastic with a piece-wise linear approximation of the hardening behaviour taken from the uniaxial test results. For the base material it was assumed that the material obeyed the von Mises flow criterion with its associated flow rule and isotropic hardening behaviour. For the cladding material, because of the observed anisotropy, the Hill flow rule with its strictly proportional hardening was assumed. Finite strain formulation was employed in all finite element calculations. Typical FE models used for the numerical analyses are shown in Fig. 2. Eight-noded plane and twenty-noded solid isoparametric elements were used for 2D- and 3D-analysis, respectively. Due to symmetry in geometry and load, only one quarter of the specimens needed to be modelled in the 3D analysis. For 3D modelling of the homogeneous specimens the mesh consisted of 720 twenty-noded isoparametric solid elements (1250 elements for a finer mesh in a verification study), Constraint effects on ductile crack growth in cladded components subjected to uniaxial loading 1057 while for the cladded specimens 2360 elements were used", + " However, one needs to perform a FE analysis with high degree of refinement to resolve the SSY fields. The solution is obtained by imposing a K-field on the remote boundary of a standard boundary-layer model (a semicracked annulus). In practice, the radius R of the boundary layer model needs to be about 100 times the plastic zone size developed due to the imposed K - field. The FE model used for evaluation of the SSY solution consisted of 840 eight-noded elements comprised in 40 rings focused toward the crack tip, Fig. 2. The crack tip was modelled by a notch with a radius of the order 10T6R. To ensure a self-similar crack-tip field, the applied K was chosen to result in a CTOD-value of at least 10 times the initial notch opening. Simulation of crack growth Stress fields ahead of a growing crack differ from those of a stationary one. For a growing crack unloading takes place near the crack tip when crack growth occurs and thus J becomes pathdependent. The order of singularity for a growing crack is such that an integration path taken very close to the tip yields a zero value for the J-integral" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000718_bmei.2010.5639976-Figure9-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000718_bmei.2010.5639976-Figure9-1.png", + "caption": "Figure 9. Force anlysis after bent in the first articulation.", + "texts": [ + " bF is the force to bend the needle, and rF is the force to resist the insertion of the needle. As shown in Fig. 8, the relationship among F , bF and rF are cos sin b r F F F F \u03b8 \u03b8 =\u23a7 \u23a8 =\u23a9 . (1) It is shown that smaller tip angle \u03b8 generates larger bending force. The value of \u03b8 is determined by both needle bending requirement and needle\u2019s using purpose. While inserting, bending in the first articulation happens firstly as it is more flexible than the head. After the head is bent an angle I\u03b2 , as shown in Fig. 9, the thrusting force IF can be divided a force IbF and a force ItF on the first articulation: sin cos Ib I I It I I F F F F \u03b2 \u03b2 =\u23a7 \u23a8 =\u23a9 . (2) IbF is parallel to the bending force bF . Although in opposite direction of bF , IbF accelerates the turning of the head as it acts on the opposite end of the head to bF . In this way, the head turned quickly while inserting the needle. ItF is parallel to the direction of the head, which thrusts the tissue of human body to move the needle forward. When the head turns, there is a force rhF acting on the head as shown in Fig. 9. rhF is the reacting force generated by the tissue of human body. It resists the turning of the head. Assume that the lengths of the head and the first section are equal. After the head is bent an angle I\u03b2 and the first section is bent an angle II\u03b2 , as shown in Fig. 10, the thrusting force IIF generates a force IIbF on the second articulation: sin cos IIb II II IIt II II F F F F \u03b2 \u03b2 =\u23a7 \u23a8 =\u23a9 . (3) IIbF is in the opposite direction to the reacting force of IbF . They turn the first section together" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002102_iecon43393.2020.9255013-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002102_iecon43393.2020.9255013-Figure2-1.png", + "caption": "Fig. 2. (a) A 17 kN MLS prototype in [13]; (b) Discrete PM with a bulge in [14].", + "texts": [ + " In this paper, a novel single-helix MLS topology is proposed, which provides an additional choice for the application of MLS in WEC system. And a simple assembly method of the discretized helical magnets is also introduced. In this paper, Section II includes the structure and working principle of the proposed MLS. Section III presents the simulation results and performance comparisons with conventional MLSs. Section IV deals with the design aspects of the magnets. The final section is the conclusion. The topology of the proposed single-helix MLS is shown in Fig.2. Similar with the traditional MLS, the proposed single-helix MLS also consists of two main parts: rotor and translator. Both the rotor and translator are spirally grooved. The rotor-nut is a hollow cylinder ferromagnetic iron with helical grooves on its inner surface, and the translator is a cylindrical ferromagnetic iron with helical grooves on its outer surface whose direction is the same as the rotor\u2019s grooves. The grooves are embedded with parallel magnetized PMs, so the rotor and the translator are coupled by the helical magnetic field generated by PMs. Compared with traditional MLS, the proposed single-helix MLS has magnets with only one polarity on the translator, which is shown by the green helixes in Fig. 2(a). Compared with the N-S alternating magnets (red helixes and blue helixes) MLS shown in Fig. 2(b), the usage of magnetic materials is reduced. From the figure, the effective working part, which is the strong coupling part of the rotor and translator magnetic fields, is the length of the rotor. However, the translator of the MLS should be long enough to meet the requirements of wave height. Consequently, most of the PMs on the translator of the traditional MLS are working in the inactive state most of the time. Also, the starting thrust force and the performance characteristic of the proposed MLS can be adjusted according to the wave state by the width of the magnets under a determined main size" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000718_bmei.2010.5639976-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000718_bmei.2010.5639976-Figure3-1.png", + "caption": "Figure 3. Proposed articulated needle", + "texts": [ + " On the other hand, a flexible needle is not easy to accurately control its trajectory, even to reach its target point. And with small-sized necks, a compliant needle is restricted for its usages. In this paper, a novel needle, namely one-dimension bendable articulated needle, is proposed for more easily and accurately steering and controlling. Its structure basic force transmission and basic steering are analyzed. II. STRUCTURE OF ARTICULATED NEEDLE The proposed needle is composed of a head, several sections and several articulations, as shown in Fig. 3, which is modified from a bevel-tip compliant needle [6]. The head is with a bevel-tip, which is used to lead the needle to be bent by its interacting with tissues of human body. The articulations are structured to be bendable in the direction along the bevel-tip. The parts of a needle between two articulations are called sections, which are more flexible than the other parts of the needle. The lengths of the head and sections are named as L0, L1, and L2 etc., respectively. The articulations are designed to be bendable only along the direction of the bevel-tip. The maximum angles of articulations are sequentially noted as 1m\u03b1 , 2m\u03b1 , and 3m\u03b1 etc. An enlarged articulation and its bent status are shown in Fig. 4. Combinations of different L and different m\u03b1 generate different needle bent shapes. Take the needle with two articulations, as shown in Fig. 3, as an example, and suppose the sections are rigid. 1) Different section lengths with equal articulation angles When L0, L1, and L2 are equal, 1m\u03b1 and 2m\u03b1 are both equal to 10 degree, the needle bent shape is given in Fig. 5(a). When L0, L1, and L2 are different, for example, L0 is half L1, and L1 is half L2, generated needle bent shape is given in Fig. 5(b). It can be seen that equal section lengths results in uniform curvature of bent needle. And different section lengths generate different curvatures" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000291_itoec49072.2020.9141670-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000291_itoec49072.2020.9141670-Figure3-1.png", + "caption": "Fig. 3 Antenna system structure", + "texts": [ + " As that ground plate of the whole antenna device, Shaped in a circular shape, 8 microstrip antenna elements are arranged circumferentially at 45 degree angle, The feed coaxial line of each antenna is linked to the beam switching antenna control module. Its function is to gate only one antenna at a time, and generate 8 directional beams by controlling 8 microstrip antennas. Each horizontal beam covers 45 degrees, and the gain within 45 degrees is greater than 8dB, thus meeting the demand of point-to-point long-distance broadband communication. The structure of the antenna array system is shown in FIG. 3. The antenna array consists of 8 microstrip antenna units arranged in sequence along the circumference, with the radiation patches facing outward and the floor facing inward, so that the communication station host and the other two band antennas can be placed in the middle part of the circumference. The 8 beam switching diagrams of the antenna array are shown in the following figure. By controlling the microwave gating array, the real-time switching of antennas 1 to 8 can be realized, and 8 beams in the horizontal direction can be realized" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001472_9781118354179.auto089-Figure12-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001472_9781118354179.auto089-Figure12-1.png", + "caption": "Figure 12. Hybrid clutch. (Reproduced with permission from ZF Friedrichshafen AG.)", + "texts": [ + " The pressure plate on the right pushes the second friction disc (on the right), splined to the hollow shaft, against the (heavy) clutch cover. Typically, there is a single clutch pedal, with the control levers and mechanism designed in such a manner that there is, as the pedal travel progresses, sequential power interruption (and establishment) to the two output shafts. With the introduction of engine start-stop technology and mild hybrid vehicles it has become viable to offer an \u201coff the shelf\u201d solution (by ZF) integrating the clutch and electric motor, as shown in Figure 12. Such solution can work well even with manual transmission, with the clutch and motor functions complementing each other and reducing vehicle fuel consumption and emissions. This is the most commonly used design and the presentation here will be concentrating on the typical assembly and component designs and materials. Figure 5 shows a cross section of the typical diaphragm clutch design, with all components of the assembly. Figure 13a show a photograph (from ZF Sachs) from the \u201cengine side,\u201d including the friction disc" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001348_ursi-emts.2010.5637349-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001348_ursi-emts.2010.5637349-Figure7-1.png", + "caption": "Fig. 7 Analysis model of the 2x2 element array", + "texts": [ + "0 mm thickness for the feed aperture is not required for the antenna operation and it is just to connect a standard waveguide (WR15) by screws. The antenna is designed so that the aperture distribution becomes uniform to obtain the highest aperture efficiency. A wide bandwidth for reflection is required to achieve a wide bandwidth in gain. The wideband 2x2-element array is designed by HFSS in this section. The design frequency is 61.5 GHz. The slot spacing is fixed to be 0.86 wavelength (4.2mm) in the x and y directions. Fig.7 shows the analysis model of the 2x2-element array. Two sets of periodic boundary walls are placed in the external region to simulate the mutual coupling in an infinite two-dimensional array of the radiating slots. Fig.8 shows the frequency characteristics of the reflection at the input port. It has the double resonances. The lower resonant frequency f1 is related to the field confinement of the radiating slot itself, which is controlled by its size to keep unchanged for its aspect ratio. The higher resonant frequency f2 is related to the field confinement by the periodic arrangement of the elements, which is changed by the slot aspect ratio" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002547_optim.2014.6850882-Figure10-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002547_optim.2014.6850882-Figure10-1.png", + "caption": "Fig. 10. Circuit proposed for the rheostatic starting of the asynchronous motor.", + "texts": [ + " The restrictive conditions required concern the limitation of the starting current: Ipmin= 0,9*I2N \u2013the lower limit and Ipmax =1,2*I2N \u2013the upper limit. This limitation induces a torque large enough even for starting on load and a low dynamic torque to ensure a slow starting of the belt-type carrier which has a large inertia moment. There has been made this analysis at the customer\u2019s request, which received an offer for the modernization of the starting installation of the high power wound rotor asynchronous motors which drive long belt-type carriers. The offer (Fig. 10) consists in changing the three-phase starting 306978-1-4799-5183-3/14/$31.00 ' 2014 IEEE rheostat with a single-phase one and a double three-phase bridge. The switches which remove the steps of the rheostat are replaced by thyristors commanded for the current limit specified. Through the study performed in this paper it is possible to establish quantitatively the effects of the distorting regime, the increase of the cost of the operation energy losses over the rheostatic starting, due to the increase of the active, reactive and distorting power", + " Using the motion equation, the relation (13) results and we compute the operation time on the step of resistance Rpk, having the speed as a variable. ( )\u2211 \u2212= \u2212\u2212 \u2212 \u03c0= kn 1kni 1ii rik t k nn M)s(M J2Tp (13) The total duration of the starting process will be: \u2211 = = 14 1k kTpTp (14) The average value of the input active power when operating on the step Rpk is: \u2211 \u2212= = ks 1ksi k1 medk1 N )s(PP (15) N- is the number of points from the interval (sk-1,\u2026.sk). The cost of the energy during the starting will be: kr.elkmed1a.el 14 1k kmed1p Tp)cQcP(C +=\u2211 = (16) Rheostatic starting proposed (Fig.10) means a distorting regime all the time and all quantities are established by numerical methods. In figure 10 there are depicted the stator and rotor windings, the three-phase rectifying bridge, the single-phase rheostat with 14 steps used for the starting process and the thyristors which remove the steps of the resistance. When operating on the step of resistance Rppk, the corresponding slips sk-1 and sk (and the speeds nk-1, nk), the powers P1k, Q1k are established numerically by the relations (6-12), according to the facts presented before. The final results of the two rheostatic starting methods analyzed are filled in the table" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003114_s10544-009-9320-x-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003114_s10544-009-9320-x-Figure2-1.png", + "caption": "Fig. 2 Illustration showing transportation of droplet via square conducting electrodes", + "texts": [ + " 2007) fw \u00bc g cos q; \u00f01\u00de where \u03b3 is the surface tension of the droplet at the liquidgas interface and \u03b8 is the contact angle of the droplet in the absence of an applied potential. From Eq. (1), it can be shown that the capillary force acting on the droplet in direction x (unit vector i) is expressed as Fx \u00bc Z L g cos q dl~n ~i; \u00f02\u00de where dl is a unit element of the droplet contour line and n is the unit normal to the contour line. Integrating Eq. (2) to obtain the total force yields Fx \u00bc g cos q Z L dl~n ~i \u00bc ag cos q; \u00f03\u00de where a is the width of the electrode. Figure 2 illustrates the contact line of a droplet traveling across two adjacent square electrodes. From Eq. (3), it is apparent that the magnitude of the capillary force is determined only by the length of the actuated electrode in contact with the droplet, i.e. it is independent of the shape of the contact line. Therefore, in optimizing the performance of EWOD devices, it is essential that the electrode configuration be designed in such a way that the length of the contact line is maximized at all times" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003376_igarss.2014.6947267-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003376_igarss.2014.6947267-Figure1-1.png", + "caption": "Fig. 1. AeroCube-4, a modified 10-cm CubeSat with extended \u201cwings\u201d that provide additional power.", + "texts": [ + " The introduction of three-axis attitude control on CubeSats has made remote sensing and Earth imaging an attractive area of development, offering high-quality science data at a fraction of the cost of a large-scale space mission. Three CubeSats of the AeroCube-4 series built by The Aerospace Corporation (Aerospace) were launched in September 2012 aboard an Atlas V launch vehicle from Vandenberg Air Force Base and delivered to a roughly 480 x 780 km altitude orbit inclined at 65 deg. An image of AeroCube-4, including its extended \u201cwings\u201d that provide additional power, appears in Fig. 1. The primary mission of AeroCube-4 is to demonstrate several technologies on orbit in the CubeSat form factor, including three-axis attitude control, highprecision rate gyroscopes, and a GPS receiver. AeroCube-4 carries three 2-megapixel (1600 x 1200) cameras, each with a different lens: fisheye, medium field of view, and narrow field of view. The fisheye lens provides all-sky imaging, the medium field of view provides a ground resolution of approximately 500 m, and the narrow field of view offers a ground resolution of approximately 50 m" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002574_s0266-3538(02)00044-1-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002574_s0266-3538(02)00044-1-Figure1-1.png", + "caption": "Fig. 1. Sketch of the spinnaker pole multi-layer laminate.", + "texts": [ + " This machine set up allowed two layers of wrappings to be placed at the desired winding angle and distance from the mean diameter. Two couples of dies and mandrels, with diameters of 80/76 mm and 60/56 mm respectively, were used to manufacture the prototypes. Only carbon fibre roving was employed as reinforcement material. Table 1 reports the mechanical properties of the fibres, as provided by the supplier. The resin is a low viscosity, high reactivity epoxy-anhydride formulation, with small quantities of filler and other additives of proprietary composition. In Fig. 1 the laminate is schematically sketched. The total number of tows and the wrapping pitch was fixed by the manufacturer, mainly on the basis of processing requirements, such as the attainment of the desired compression within the die, for a good consolidation of the composite, and of a smooth surface finish, with a sustainable pull-force for the machine. For the two prototypes the number of tows in each layer, the tow angles and the layers\u2019 thickness, as experimentally measured with a reflection microscope on polished crosssection surfaces, are reported in Table 2", + " The relevant literature [15\u201317] offers two possible approaches to the mechanical problem: (i) a micro-mechanics approach which takes into account the microscopic nature of the material, looking at the properties of fibres and matrix as well as at the effective nature of the interface between them; (ii) a macro-mechanics approach which analyses the behaviour of the composite as a whole exhibiting, generally, an orthotropic behaviour. Hereafter a micromechanics approach is utilised to predict the moduli of the composite from those of the constituent materials by means of the r-o-m; while phenomena like interaction between contiguous layers are treated, for simplicity, at a macro-scale level. Thereafter we refer to the FE model of the 80 mm pole specimen used to calibrate the numerical simulation. As described in Section 2 (see also Fig. 1) the pole presents five different layers. Referring to Fig. 2, where a portion of the generic pole cross section is schematically depicted, the layers are numbered sequentially going from the outer layer (ply number 1) to the inner one (ply number 5). The plies #1, #3 and #5, with fibres oriented at W=0 with respect to the longitudinal axis of the prototype, are discretised with \u2018\u20183D-solid\u2019\u2019 isoparametric elements with 27 nodes per element, while plies #2 and #4, with wound fibres at W = 76 respectively, are modelled through 2D 9-node isoparametric \u2018\u2018shell elements\u2019\u2019" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002572_978-3-319-76276-0_5-Figure15-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002572_978-3-319-76276-0_5-Figure15-1.png", + "caption": "Fig. 15 Contour plot for fatigue life of primary suspension spring", + "texts": [ + " The shear stress and fatigue life for the modified suspension spring is given in Table 4 and contour plot for fatigue T ab le 4 C om pa ri so n of sh ea r st re ss es an d fa tig ue fo r ex is tin g an d pr op os ed su sp en si on sp ri ng fo r in cr ea se in m ea n di am et er Pa rt ic ul ar s U ni t E xi st in g pa ra m et er s Pr op os ed fo r W A G -9 M id dl e ax le E nd ax le sp ri ng M id dl e ax le E nd ax le sp ri ng O ut er sp ri ng In ne r sp ri ng O ut er sp ri ng In ne r sp ri ng Fo rc e (W ) N 31 26 8. 7 93 51 .4 6 40 45 1. 38 28 05 0. 3 79 36 .6 8 42 75 7. 68 D efl ec tio n (d ) m m 65 .6 64 .6 45 .6 69 .2 6 67 .2 6 49 .2 6 Sh ea r st re ss N /m m 2 58 3 59 8. 36 53 0. 98 54 1. 33 53 9. 96 56 1. 25 Fa tig ue lif e C yc le 3. 2 10 6 1. 62 10 4 4. 64 10 7 5. 48 10 7 5. 8 10 4 5. 3 10 7 life is shown in Fig. 15 which shows very little improvement in shear stress and fatigue life as compared to the original spring and hence, this modification is not advisable. The fatigue analysis reveals that the middle axle inner suspension spring has a finite life of 1.89 104 cycles which clearly indicates that the spring fails because of fatigue failure observed from the cross section of the failed spring with crack initiation at inside diameter. It has been also roughly estimated that the spring failure occurs within 90 days and it has been also confirmed from loco shed authorities" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003751_tencon.2015.7373193-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003751_tencon.2015.7373193-Figure1-1.png", + "caption": "Figure 1. Reference frames and force for schematic representation of a quadrotor", + "texts": [ + " The paper is composed of six sections followed by a conclusion. Section II presents the fundamental equations of motion of the quadrotor\u2019s dynamic model. Section III introduces the inner-loop controller design. Section IV describes the position and velocity features that are exploited in the outer-loop control law. Section V introduces the proposed outer-loop control strategy and its stability analysis. Section VI shows the experimental results obtained. The quadrotor vehicle is composed of a rigid frame equipped with four rotors shown in Fig.1. The system is modeled as a rigid body of mass m controlled through thrust and angular velocity inputs [3]\u2013[6]. 978-1-4799-8641-5/15/$31.00 c\u00a92015 IEEE 1 To describe the motion of the quadrotor, two reference frames are introduced: the inertial reference frame {I} fixed to the earth surface and the body-fixed frame {B} attached to the quadrotor\u2019s center of gravity. Let R = I BR \u2208 SO(3) denote the rotation matrix which transfers vectors quantities from the frame {B} to {I} and let \u03be \u2208 IR3 represents the position vector of the vehicle expressed in {I}" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001395_ijvp.2016.075351-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001395_ijvp.2016.075351-Figure4-1.png", + "caption": "Figure 4 Flexible body coordinates (see online version for colours)", + "texts": [ + " In the FE/FFR formulation, a deformable body can be divided into more than one finite element in order to obtain more accurate results. Using the element shape function matrix, the global position vector of an arbitrary point ijP on the finite element j of the deformable body i can be written as 0 ,ij ij ij ij f= + +r R u u where ijR is the position vector of a reference point ,ijO and 0 iju and ij fu are the undeformed and deformed local position of the arbitrary point ,ijP respectively, as shown in Figure 4. The local position vector of the arbitrary point in the FE/FFR formulation can be written as 0 ,ij ij ij f= +u u u or equivalently, 1 ,ij i ij ij ij ij ij i ij i n n= = =u A u C S C B q N q where iA is the rotation matrix that defines the orientation of the body reference, ijS is the shape function matrix, 1 ijB is a constant Boolean matrix describing the element connectivity conditions, i nq is the vector of nodal coordinates of body i, 1 ij ij ij ij=ijN C S C B is a space dependent matrix, and the element transformation matrices ijC and ijC are defined in the appendix (Shabana, 2014)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003247_9781119258827.ch7-Figure7.28-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003247_9781119258827.ch7-Figure7.28-1.png", + "caption": "Figure 7.28 Forces acting on the skid steered vehicle during a steady\u2010state turn. Only the tyre forces acting on the front outer and the rear inner tyres are shown.", + "texts": [ + " This modified process is shown to correlate better with some combined slip measurements for a car tyre, and is the method used here. The complete algorithms for both methods are given in the relevant papers [7.20, 7.21]. The combined slip properties for the tyres used in the vehicle models are shown in Figure 7.27a and b. A much simpler tyre model could be developed by assuming the longitudinal force\u2013 slip and lateral force\u2013slip properties and load dependencies are the same; appropriate values could be obtained by averaging them. Figure 7.28 shows the forces acting on the vehicle during a steady\u2010state turn. Only the tyre forces acting on the front outer and the rear inner tyres are shown. The basic dimensions (wheelbase and track) are the same as those for the AMX 10RC. The vehicle weight is taken as 180 kN with equal wheel loading of 30 kN per wheel in the static condition. The wheels on each side of the vehicle are assumed to rotate at the same speed (except for one neutral turn case). The equations of motion and their solutions are generally the same as those for tracked vehicles except, of course, there are no track effects" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003670_tap.2020.3000517-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003670_tap.2020.3000517-Figure4-1.png", + "caption": "Fig. 4. Configuration of the beam-steerable RA.", + "texts": [ + "96 bits at 10.0 GHz. The frequency range over which the phase resolution is equal to or better than 1.7 bits was found to be from 9.0 GHz to 11.5 GHz. However, the co-polarization reflection coefficient of Mode 3 becomes slightly lower than -1 dB above 11 GHz. Therefore, we determined the operating frequency range for the 2-bit phase shifter as 9.0-11.0 GHz. The proposed 2-bit phase shifters were used to construct a beam-steerable reflectarray having a circular aperture with a diameter of 30 cm. Fig. 4 shows the configuration of the reflectarray along with a feed horn antenna used to illuminate the array. The horn antenna has aperture dimensions of 4 \u00d7 4 cm2 and is placed at a focal distance of 25.4 cm. The feed antenna is oriented so that the polarization of the incident wave is along the u\u0302-axis, or parallel to a diagonal line of the square-shaped unit cells in the array. First, we simulated the feed horn antenna and extracted the amplitude Authorized licensed use limited to: University of Glasgow" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002190_iemdc.2011.5994637-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002190_iemdc.2011.5994637-Figure2-1.png", + "caption": "Fig. 2. FEMM Mesh of the Generator", + "texts": [ + " Since FEMM supports LUA scripting language, an extensive LUA scripted library is made to draw the layout of the generator in FEMM to ease us through the process of drawing and redrawing on every simulation made. All the parameter, materials and ratios stated in the section above is inserted into a LUA parameter file and the LUA modeler is used to model the entire generator. After modeling, meshing of finite elements is done over the whole cross section of the generator. The meshing yields approximately 20,000 nodes and 38,000 elements. To improve the estimation of cogging torque, the meshing around the arcs and corners have higher resolution as shown in Fig. 2. It is a well known fact that by increasing the meshing resolution, a higher accuracy in estimation can be reached but only to a certain extent. Where by the increase of meshing nodes of 20,000 to 140,000 produces an increase in accuracy of approximately 0.2% [6]. D. Simulation steps Simulation starts of with setting the LUA parameter script with the corresponding number of stator slots. The LUA modeler script is then executed to generate the layout for the generator for that corresponding number of stator slots" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002575_125906-Figure10-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002575_125906-Figure10-1.png", + "caption": "Figure 10. Hub-and-spoke strain transducer.", + "texts": [ + " DC power for the force sensor; 4. screw; 5. nut A; 6. force sensor; 7. nut B; 8. hinge-levers; 9. transducer of the force sensor; 10. Data Acquisition Card; 11. solid cylindrical GMM structure. Digital controlled stabilized current supply is adopted, with continuously adjustable output voltage 0\u201330 V, and output current \u221210 to 10 A. The pre-tightening force is measured by force sensor, which is a hub-and-spoke strain transducer, with the measurement range 0\u20131 T, and its structure is as shown in figure\u00a010. The relation between working current and pre-tightening force is as shown in figure\u00a011. In the measurement experiment, the current is uniform with step 0.1 A from \u22123 A to 3 A, and then from 3 A to \u22123 A. The max pre-tightening force is up to 1892.87 N when the working current is 3 A. The pre-tightening force Fp is measured by the force sensor, and then it is output in the form of voltages. The relation between the voltage U and Fp is as follows: F U2000 N .p ( )= (2) Usually, the value of Fp is determined by certain application requirement" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000489_tgrs.2021.3051727-Figure7-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000489_tgrs.2021.3051727-Figure7-1.png", + "caption": "Fig. 7. Schematic of the SP formation. The primary satellite and virtual reference satellite form a Dual-Helix formation. The same reference satellite and auxiliary satellites form a Pendulum formation.", + "texts": [ + " Thus, SP formation can reduce the cost of auxiliary satellites and has the potential to be applied to the spaceborne M-SAR mission. It is worth mentioning that the cost reduction of SP formation is mainly due to the use of microauxiliary satellites with only radar echo reception capability. Compared with the traditional SAR satellites, the auxiliary satellite has the characteristics of a small antenna and no large-capacity battery. The SP formation is designed based on the Dual-Helix and Pendulum formations, as shown in Fig. 7. The primary satellite and a virtual reference satellite form a Dual-Helix formation, and several auxiliary satellites and the same virtual reference satellite form a Pendulum formation. Here, the baselines are defined with respect to the primary satellite and auxiliary satellites. The baseline variations in the SP formation can be obtained, as shown in Fig. 4(d). In order to understand SP formation more intuitively, (8) is used to describe it. The LVLH frame is attached to the primary satellite" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002362_amm.364.285-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002362_amm.364.285-Figure5-1.png", + "caption": "Fig. 5 The operation process of the manipulator 1-body 2-feeder 3-beam 4-slipway 5-balance cylinder 6-servo motor 7-X axis transmission system", + "texts": [ + " In such a short period of time, the manipulator has to operate at a high speed of 200-250 m/min, so that the operations such as loading, moving and unloading can be fulfilled. And this has put forward higher requirements to the mechanical structure, material friction characteristics, and structural dynamic characteristics of the robot manipulator [8] . According to the composition of the stamping process sets of on-line system and the role of the high-speed plate carrying manipulator, the working cycle of a manipulator contains nine action processes, as shown in figure 5 Based on the above analysis, we have developed a three-dimensional model of the manipulator as shown in figure6, which can be used to analyze the dynamic characteristics of the manipulator. Modeling and simulation of stamping processing set of control line system based on ADAMS The analysis of the simulation process After setting the joints from all directions, we begin to set the parameters of the motor; we call this Motion\u2019s parameters setting. We need to add motion to Translation Joint, Revolute Joint, and Gear Joint, respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001032_milcom.1994.473843-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001032_milcom.1994.473843-Figure2-1.png", + "caption": "Figure 2. Two Dimensional Example with Complex Weights", + "texts": [ + " From this example, two items should be clear. First, for real beam weights, there is no impact on performance if all beam weight components are multiplied by -1. Second, an important step in selecting beamweights is to determine whether the positive or negative version of the optimum for each user terminal should be used as the basis for the rest of the algorithm. These concepts can be extended for complex weights as follows. Consider a problem with two complex beam weights and two user terminals as shown in Figure 2. Both the magnitude and angle of the weights must now be considered. In this example, individual optimum weights, separately determined, are shown in Figure 2 for terminals a and fl. In the figure, the optima are represented by the points A and B respectively in both the magnitude and angle spaces. In this case, equivalent gain magnitudes can be obtained with the phase of the beam weight components modified to any value, as long as all components are modified by the same delta. Therefore, sliding the angle of the A beam weight for the a terminal along the straight line toward the angle of B has no impact on the magnitude of the gain toward terminal a and makes the resulting beam weight much closer to the optimum for the terminal" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000590_710664-FigureA-4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000590_710664-FigureA-4-1.png", + "caption": "Fig. A-4. In a 200 cc single-cyl engine, the primary force is about 768", + "texts": [], + "surrounding_texts": [ + "2364 HARRY I. HAZZARD\nlatest techniques of flow patterns and computer analysis. The porting is ample and would permit higher speeds, if needed. The engine has exceeded 104 bmep at 6000 rpm. Pistons are of a slipper skirt type, with the pin boss spaced down from the top. The spacing and cooling passages are needed to maintain tolerable temperatures at the pin bearing. Connecting rods are conventional chainsaw design of forged steel with the caps broken off after hardening. The fracture structure around the screw provides the location or doweling to locate the cap while finish grinding the bore, and in service. The technique requires that the metal thickness around the screw be hardened deep enough to prevent plastic deformation at the core.\nCrankpin diameter is critical, as the speed is high and the crankshaft heavily loaded. A pressed-together crankshaft with caged crankpin bearings would operate at higher speeds than the full-complement roller bearing before burnout. How ever, these cages would reduce the number of rollers and thus could cause failure by shaft pitting due to high loading, unless the bearings were longer. If the bearings were longer, the shaft would be weak in the throws, unless the cylinder center distance were increased. Bearing burnout speed was estimated at 8200 rpm for the uncaged bearing, and the fatigue life is adequate, so it was considered the best compromise (Fig.9). The balanced power engines were planned to suit the ma chines in which they may be used. They can be installed in verted or laid flat. Cooling should be adequate, and the low shaking forces will simplify mountings. The inlet and exhaust systems are designed as a part of the engine to assure com patibility and optimum performance. The design is our best calculation of the proper compromise to provide a good power unit for recreation vehicles of the snowmobile or A.T.V. type.\nAPPENDIX\nBALANCED POWER BALANCE\nThe balance of the McCulloch balanced power (BP) engines is achieved by letting the shaking forces cancel themselves. They are not isolated or suppressed, but are directed against each other within the solid structure of the engine so that they are not manifest externally. The BP single-cyl engine has 11% of the vibration of a conventional twin of twice its displace ment. The BP twin has about 14.4% of the vibration of a con ventional twin of its size. Such figures demand an explanation which must begin with fundamentals (Fig. A-1). These figures are fundamental principles of shaking forces in a reciprocating engine type of mechanism. In a single-cyl en gine, all of the rotating mechanism, including the crankshaft end of the connecting rod, is balanced by a rotating counter balance weight. This weight is on the opposite side of the crankshaft and is divided with about half on each side of the cylinder. This division is required because the individual counterweights cannot be directly opposite the mass which they are to balance. If the counterbalance were on one side, it would be offset from the mass which it is to balance, thus creating a rocking couple. This dynamic unbalance would cause the engine to shake in opposite directions at its opposite ends, a condition like a playground teeter-totter or an auto", + "RECREATIONAL VEHICLE ENGINES 2365\nmobile with both front wheels out of balance by the same amount, but with one wheel half a revolution ahead of the other. If the out-of-balance and the counterbalance were sep arated by an appreciable distance, they would be viewed as separate disturbances. If they were closer together, they would show up as a rocking disturbance or a rocking couple. It is the product of the counterbalance force times the amount it is offset from the out-of-balance force which it is to balance out. This rocking couple factor is of basic importance as it appears so often and can spoil an otherwise perfectly statically balanced system. The rotating parts of a mechanism can usually be balanced, but the reciprocating weights in a single-cyl engine are quite different. The piston end of the rod and the piston assembly move up and down and produce sizable shaking forces length wise of the cylinder only. Any attempt to balance them with rotating weights introduces extra forces crosswise of the cyl inder. This is the basis for the 50% balance in 1-cyl engines where half of the reciprocating forces are balanced by a rotat ing mass. The fallacy of this is that the 50% reduction length wise of the cylinder shows up as an unbalance added crosswise of the cylinder. In a logger's chain saw, the cylinder is paral lel to the bar which gives higher inertia in this direction, so the reciprocating mass is 30% counterbalanced leaving 70% in the direction of high inertia and adding 30% crosswise. You do not get something for nothing, but the higher shaking forces can sometimes be directed where they are the least disturbing;\nFig. A-2 also shows this. Another way of handling reciprocating unbalance is to try\nto ignore it and add soft rubber vibration isolators. This can be made to reduce the shake of the handles or of other parts of the unit where it will be noticed during the operation. The shaking forces are actually undiminished and expend them selves on shaking the engine worse than if it were anchored to a larger mass. The chain or belt drive from the engine also takes a beating from the shaking engine. The reciprocating unbalance is made of forces at two dif\nferent frequencies. The primary is at engine speed, and the secondary is at twice engine speed. The secondary or double speed force is less than the primary force, but because of its high frequency, it is quite annoying and requires peculiar treatment. Also, if 50% of the primary force is counterbal\nanced as noted above, the secondary then becomes 50-60% of the remaining primary force. The origin of the secondary force at twice engine speed is in the crank-connecting rod relationship Fig. A-3). In the first quarter of the engine rev olution from top center, the crank end of the rod swings out sideways and pulls the piston down more than half of the stroke. In the second quarter of the revolution, the rod straightens up and therefore moves the piston down less than half the stroke. This irregular motion of the piston relative to the crankshaft rotation is completed in half a revolution or twice in each revolution. The forces required to produce this motion therefore occur twice per revolution and, hence, have a frequency twice that of the engine speed. The shaking force is that force required to get the piston started from zero velocity at the top of the stroke and then to stop it at the bottom of the stroke, from 0-3000 ft/min. and back to zero in 0.01 sec (6000 rpm). The magnitude of the secondary or twice engine speed component is 25-30% of the primary and for this discussion will be simplified to 30%. In a single cylinder of an engine, the reciprocating shaking\nforces are only lengthwise of the cylinder and by formula are\nThe 70,480 is a figure accounting for the force required to accelerate a mass and to make the result come out in pounds. TheCos? is a term from trigonometry representing the ratio of the sides of a right triangle, one of whose angles are ?. The ? is the angle of crank rotation from top center. The Cos ? times the factor outside the parentheses represents the primary shaking force. The Cos 2 ? times the factor outside the paren theses is the secondary shaking force. Note that 2 ? indicates that it is changing at twice crankshaft angle ? or engine speed. The S/2L is half of the stroke divided by the connecting rod length in inches. These shaking forces act along the cylinder axis only, with no crosswise factors; it is shown as a curve in", + "2366 HARRY I. HAZZARD\ncritical factor in shaking forces where a change from 6000 to 7000 rpm means a 36% increase in forces. A speed of 8000 rpm would raise it another 31%, or an 80% increase over 6000 rpm. The above forces are based on actual values from an en gine; 0.734 lb reciprocating weight, 2.048 in. stroke, and a 3.375 in. long rod. In combining the shaking forces in an engine, the position of the cylinders and the angles of the crankshaft are both im portant. With this background, the forces on several engine configurations will be analyzed. A conventional single-cyl engine (Fig. A-5, left) just shakes up and down the cylinder axis under the forces from the reciprocating parts. The above 200-cc cylinder would have a primary shake of 768 lb and 233 lb of secondary. The secondary at twice engine speed cannot be helped by weights on the crankshaft, but the primary at crankshaft speed can be reduced. By balancing out half of the reciprocating force with a rotating weight on the crankshaft, the shake lengthwise of the cylinder is reduced to 384 lb of primary and 233 lb of secondary. Of course, the weight on the shaft introduces a shaking force of 384 lb crosswise of the cylinder. These forces show up in full value on any engine mountings or 154 lb on each of four mountings. A conventional twin (Fig A-5, right) has one piston going up\nas the other goes down, so the primary shaking forces cause a rocking couple of 3840 lb-in. with a 5 in. cylinder center distance. This can be cut in half by a rocking couple from weights on the crankshaft at the cost of adding a horizontal rocking couple. This means if there are four mountings ar ranged in a 15 sq in. pattern, which is assumed to be a prac tical configuration, the load on each mount would be 1920 divided by 15 and by 2 or 64 lb, up on the mounts on one end and down on those at the other end. There is still the matter of the secondary forces which are at twice crank angle, and crankpins at 180 deg means that these forces add together in a twin. This will give a total of 466 lb of shaking force with no practical counterbalancing.\nA conventional 3-cyl engine (Fig. A-6) has a wide spacing between the end cylinders and has a primary rocking couple of 5760 in.-lb, which can be reduced to 2880 in.-lb by weights on the shaft or 96 lb on each of the four mounts spaced at 15 in. The secondary is also a rocking couple of 1748 in.-lb with no way to counterbalance it twice crank speed, so it adds 59 lb to the mounts. Note that these are rocking cou ples. By adding another three cyl on the end with the proper timing, the forces are all balanced in a six inline engine. A conventional 4-cyl (Fig. A-7), 4-cycle engine would be a\npair of twins and the primary rocking couple would cancel, but the secondaries still add to a total of 932 lb shaking force. A conventional 4-cyl (Fig. A-8), 2-cycle with cranks at 90 deg to give uniform firing impulses would be a pair of" + ] + }, + { + "image_filename": "designv6_24_0001924_sp.j.2096-5796.2019.0016-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001924_sp.j.2096-5796.2019.0016-Figure5-1.png", + "caption": "Figure 5 Non-radial three-beam based 6 DOF F/T sensor[29].", + "texts": [ + " It is an integral spoke structure with a flexible link between the crossbeam and flange, thus simplified mechanical model of the elastomer as a cantilever beam. Strain gauges attached on the four main beams compose 6 Wheatstone bridges for six-dimensional force / torque measurement. The SAFMS 6 DOF F / T sensors, developed in 1987 by Hefei institute of intelligent machinery, in collaboration with Southeast University, adopted this 2019\u5e74 \u7b2c1\u5377 \u7b2c2\u671f\uff1a121\u2014135\u865a\u62df\u73b0\u5b9e\u4e0e\u667a\u80fd\u786c\u4ef6 Virtual Reality & Intelligent Hardware mechanical structure[27], as seen in Figure 6. A sensor with a non-diametric three-beam centrosymmetric structure is shown in Figure 5. The inner edge and outer edge of the sensor are fixed on the arm and claw of a robot respectively, and the force exerted to the sensor is transmitted by three beams tangle to the inner edge. Two pair of strain gauges is attached to top, bottom, left, and right sides of each beam, so that six groups of half-bridges can be formed. Exact solution of the six-dimensional forces / torques can be obtained by decoupling the six groups of bridges. This kind of force sensor has high rigidity. It was first proposed by Carnegie Melon University, and also studied by Huazhong University of Science and Technology, that is HUST-FS6 6 DOF F/T sensor[28]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003710_s0141-6359(03)00042-4-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003710_s0141-6359(03)00042-4-Figure5-1.png", + "caption": "Fig. 5. Photograph of mesoscopic mechatronic polymer gripper system.", + "texts": [ + " For the minimization of information content of the gripper system, the planar gripper mechanism including structure frame is employed to align and install the actuator and sensor. The fabricated gripper mechanism and coil winding are assembled first by utilizing simple hand tools and with adhesive glue. Then, the permanent magnets are assembled on the mechanical frame by using holding plates and screws. Finally, an opto-electronic position sensor, KT3092, which is available from industrial products, is calibrated and assembled to give 5 m accuracy. The fabricated prototype mesoscopic mechatronic gripper system is shown in Fig. 5. The experimental setup is shown schematically in Fig. 6. The fabricated mesoscopic gripper system is mounted on a testing platform with a step driving motor for horizontal and vertical motion driven test. An opto-electronic position sensor obtains the dynamic response. The overall system under testing is controlled by utilizing a 286 PC. The dynamic transfer properties of the mesoscopic mechatronic gripper system are identified in the time and frequency domain. The mesoscopic mechatronic gripper system is tested and measured to obtain characteristic curves of sensing, actuation, and gripping operations" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002469_cdc.1990.203965-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002469_cdc.1990.203965-Figure1-1.png", + "caption": "Figure 1. A Proposed Configuration for the DIGITS System", + "texts": [ + " Finally, the torsional resistance capability of a power grasp is also studied to complete the basic power grasp stability analysis. The basic model used to study power grasps is taken from the DIGITS Grasping System which is described in the next section. I1 THE DIGITS SYSTEM The DIGITS (Dexterous Integrated Grasping with Intrinsic Tactile Sensing) System is a grasping system currently under development at OSU for the study of the grasping and manipulation problems [lo]. It consists of four %-link, 3-degree of-freedom fingers, each with a position, velocity and torque sensing capability (Fig. 1). Each joint is belt-driven by a closely-located brushless DC motor to obtain high velocities and large force levels at the fingertips. Under development is a six-axis force/torque sensor for the fingertips which is an integral part of the finger design. Closed-form algorithms have been developed which extract the location and force levels at contacts from the six-axis force data [11,12]. Some of the preliminary tests indicate continuous fingertip force levels of 4 lb., a maximum fingertip velocity in excess of 50 in/sec, and a 1 in" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000083_ir-03-2018-0042-Figure19-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000083_ir-03-2018-0042-Figure19-1.png", + "caption": "Figure 19 FEA of the wearable vehicle in walking mode Figure 20 FEA of the wearable vehicle in fast motion mode", + "texts": [ + " The proposed design is examined by using SolidWorks 2014 software, to check its safety in different modes of operation and to determine themaximum stress and deflection of the wearable vehicle. The wearable vehicle is examined in both walking mode and fast motion mode. The wearable vehicle is fixed on the feet. Force and torque are then applied on the backpack system. The wearer weight is 80kg and the weight of the backpack is 55kg including the weight of the batteries and electrical components. The simulation analysis of the walking mode is shown in Figure 19, while the simulation analysis of the fast mode is demonstrated in Figure 20. The simulation recorded a stress of 143.5MPa compared to the yield strength of the material 500MPa, which is safe. In addition, the maximum displacement is 19mm. For the fast motion mode, the wearable vehicle is fixed on the axis of the four wheels. The maximum loads are the same as in the previous state. In addition, the wearer\u2019s weight is acting on the seat. This seat is fixed on the trunk and the thigh link. The maximum stress is 141" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001305_024-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001305_024-Figure3-1.png", + "caption": "Figure 3. SEM photograph of the micromotor.", + "texts": [], + "surrounding_texts": [ + "The electrodes are excited with a sinusoidal signal frequency between 50 and 800 kHz and with an amplitude which can reach 50 V. Membrane deflection is converted to a voltage signal by a laser vibrometer controller (Polytec OFV 3001) [4]. XY cartography, controlled by the computer, allows resonant mode visualization (in amplitude and phase). Figure 4 shows an example of the B0,5 mode measurement. The characterization is done in atmospheric pressure. Only one electrode is excited with 40 V applied voltage and 150 kHz frequency. Travelling waves (i.e. with phase depending on spatial coordinates) were also observed on the stator membrane surface. In that case, the two sets of electrodes with a \u03bb/4 phase difference in position (\u03bb = wavelength of the B0,5 mode) are both excited. On the first set we applied the signal V0 sin(\u03c9t) and on the second set the signal V0 sin(\u03c9t \u00b1 \u03c0/2). 171 Figure 5 shows the B0,5 travelling wave measurement for the following conditions: atmospheric pressure, V0 = 10 V, bias voltage = 30 V, frequency \u2248 150 kHz, which fits the mode B0,5." + ] + }, + { + "image_filename": "designv6_24_0000234_icelmach.2016.7732762-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000234_icelmach.2016.7732762-Figure5-1.png", + "caption": "Fig. 5. No load flux lines and flux density of mcB: PM magnetization direction as in Fig. 4", + "texts": [ + " The measured temperature at steady state of the prototypes are reported for reference in Section III. B. mcB: for 2 and 4 poles mcB is suitable to be used in 2- and 4-pole machines. Its lamination is shown in Fig. 4 and its geometrical data are reported in Table III. The size of mcB is similar to the size of mcA, with the exception that the external diameter has been slightly increased (from 200 mm to 210 mm). Due to the higher PM flux of this rotor structure, the stator back iron flux density in the 2-pole machine would be higher than 1.8T . The mcB PM volume is the same of the mcA one. Fig. 5 shows the distribution of flux lines and the correspondent flux density at no load for the 2- and 4-pole machines. It can be noted that the stator back iron flux density is 1.6T in the 2-pole machine. As expected, in the 4-pole machine the stator back iron flux density is 0.8T that is, halved with respect the 2-pole machine one. The performance of the mcB are shown in Table IV as a function of the number of poles, together with the losses components. The rotor structure of mcB allows a high performance to be achieved both in 2- and 4-pole machines" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002843_robot.1987.1087785-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002843_robot.1987.1087785-Figure1-1.png", + "caption": "Figure 1: The experimental set-up", + "texts": [ + " Force Control of A One D.O.F. Direct-Drive Arm 3.1. Modeling and Identification In this section, some basic haracteristics of force control were investigated through experiments on a one degree-of-freedom direct-drive arm. The motor speed of response, sensor/environment stiffness and sensor sensitivity are identified to play important roles in the performance. Here we concentrate at how the properties associated with direct-drive enhance the performance of force control. The experiment set-up is illustrated in Figure 1. The arm, which rotates in the horizontal plane, is driven by a brushless permanent-magnet DC motor powered by a current controlled PWM amplifier. A force sensor is mounted at the tip of the arm. With the link on the motor, the frequency response test was done to determine the transfer function relating the angular velocity of the motor to the input voltage to the amplifier. The data is shown in Figure 2. These data show that the motor has a dominant pole at 13 Hz. At about 200 H z , a resonant peak is observed, which is due to the flexibilities of the link and the coupling between the link and the motor shaft" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000234_icelmach.2016.7732762-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000234_icelmach.2016.7732762-Figure3-1.png", + "caption": "Fig. 3. No load flux lines and flux density of mcA: PM magnetization direction as in Fig. 2", + "texts": [ + " It is a lamination suitable to be used in 2-, 4- and 6-pole LSSM, respectively, by changing the magnetization direction of the PMs in the rotor and the stator winding. The slots occupied by a phase of the stator winding are highlighted in Fig. 2 for the different number of poles. The stator lamination is a standard 4-pole IM one. The rotor exhibits symmetry over 60 mechanical degrees and there are 30 rotor slots. The considered PMs are high energy NdFeB. They are oriented so as to form 1, 2 and 3 pole pairs, respectively. The machines with different number of poles have been analyzed by means of FE analysis. Fig. 3 illustrates the no load flux lines and the flux density. In the 4-pole machine, 3 barriers over 6 contain PM with different magnetization direction. In order to avoid PM flux to be short-circuited between any other pole, the rotor bar at the center of the PM is deeper with respect the other bars, as in Fig. 2. As illustrated in Fig. 3(a), the 2-pole machine exhibits a high flux density in the back iron (1.8T ) and the associated iron losses will be relatively high. As reference, the flux density in the tooth and in the back iron at no load in the corresponding IM is 1.6T and 1.5T , respectively. The performance of mcA are shown in Table II as a function of the number of poles. The highest power at which the IE4 efficiency is reached is associated to the rated power. The losses components (Joule losses Pj and iron losses Piron) varying with the number of pole of the machine are highlighted" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003460_esars.2015.7101415-Figure14-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003460_esars.2015.7101415-Figure14-1.png", + "caption": "Fig. 14. 3D thermal Analysis of the machine for 10000 rpm and full load operation with cooling system", + "texts": [ + " High temperatures results in a change in the PM properties and an increase of demagnetization risk. It can also damage the wiring insulation. Therefore, a coolant system seems to be necessary. In this paper a frame made by aluminum which has a good thermal properties and a low mass density with water cooling channels is used. Fig.15 shows the cooling system of the proposed machine. The water with 80 centigrade degree flows in the cooling channels, and reduces the temperature in different parts of the machine as shown in Fig.14. Table III shows hot spot temperatures in different parts of the machine for both without and with cooling system It is understood by using a cooling system the temperature will be decreased significantly at critical points of the machine especially on permanent magnets which temperature reduces from 173o to 102oC degrees. The thermal characteristic of motor and dimensions of cooling system is presented in Table IV. Iron loss calculation of a Halbach array permanent magnet motor is presented. To reduce the iron losses calculation time six critical points are considered and iron losses are calculated in these points instead of all machine elements as it is done in pure finite element methods" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001900_6.2003-5815-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001900_6.2003-5815-Figure8-1.png", + "caption": "Figure 8 \u2013 Concept of the Wright Flyer Simulator. Note the large frames allowing transportability of the system.", + "texts": [ + " The mathematical model of the Wright Flyer (and other Wright aircraft 12) has been operating at the University of Liverpool in conjunction with their full flight research simulator. The model is based on the NASA Ames wind-tunnel data, obtained from testing with a full-scale replica of the 1903 Wright Flyer. As part of American Institute of Aeronautics and Astronautics 9 this project, some additional features (aeroelasticity and control system) were incorporated. The model operates under the ART real-time FLIGHTLAB environment. Figure 8 shows the concept design of the Wright Flyer Simulator. This is intended for demonstrations and educational applications. Transportability is made possible by mounting the system on a folding frame. The presence of the wing center section enhances the realism of the flying experience, and the appeal of the simulator in general. The design of the Wright Flyer simulator has brought many interesting simulation design issues forward. Although the project has not met its destiny, it is clear that the process applied in the design, and even the design itself, have demonstrated a systematic approach to simulator design" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000298_1.1807415-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000298_1.1807415-Figure8-1.png", + "caption": "Fig. 8 Pressure distribution of flow-constant hydrostatic bearing: \u201ea\u2026 hydrostatic-operating, \u201eb\u2026 hybrid-operating", + "texts": [ + " Due to the rotation of the journal both hydrodynamic and hydrostatic lubrication effects can play essential roles in this kind of journal bearing. Hence, an externally pressurized journal bearing is a hybrid-operating bearing if the journal rotation speed is high. rom: http://gasturbinespower.asmedigitalcollection.asme.org/ on 08/07/20 In addition, depending on the fluid supply system, there are two categories of hydrostatic bearing: flow-constant and pressureconstant. Flow-Constant Hydrostatic Bearing. A set of pressure profiles for a flow-constant hydrostatic journal bearing with four pockets can be seen in Fig. 8. It can be seen that there is an obvious hydrodynamic effect in this hybrid-operating condition due to the rotation of the journal. In contrast, the pressure profile of a hydrostatic-operating bearing is symmetric relative to the eccentricity direction. Pressure-Constant Hydrostatic Bearing. In pressureconstant hydrostatic bearings, a restrictor such as an orifice or a capillary, should be used to compensate the pocket pressures, to produce the load capacity of the bearing. The pressure-constant hydrostatic bearing is widely used in engineering applications" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001943_0890-6955(95)00104-2-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001943_0890-6955(95)00104-2-Figure1-1.png", + "caption": "Fig. 1. Variable pitch screw transmission mechanism.", + "texts": [ + " The position and orientation of the kth frame, (xyz)k, with respect to frame (xyz)j can be written as: k -(A_,= H i-J_A i (2) i=j+ I In homogeneous coordinate notation a point vector ka r = arxi + a ~ + a,-~k is written as a column matrix kar = [area,yar~l ]7. Vectors of the form [rlq.nq>11qzO] T are used to represent direction. Given a point kay, its transformation Ja_, is represented by the matrix product Jar~-J.A_k k.a_r (3) The Denavit-Hartenberg notation is an important mathematical tool in the kinematic analysis of open and closed-loop mechanisms [11]. It can also characterize the abilities of multi-axis machine tools and generate the desired cutter locations. The illustrative example shown in Fig. 1 is the variable pitch screw transmission mechanism (VPSTM) used in a shuttleless weaving loom [1]. It consists of a driving slidercrank mechanism and a driven variable pitch screw possessing four cylindrical meshing elements (Fig. 2). Elements 5a and 5b engage one thread on one side of the screw, while elements 5c and 5d (not shown in Fig. 2) engage the other thread on the other side. Here only the thread meshed with elements 5a and 5b is considered, since after this threaded profile is machined, the screw blank can be turned 180 \u00b0 and the same cutting path followed in machining the second thread" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003378_bf02765177-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003378_bf02765177-Figure2-1.png", + "caption": "Fig. 2. Diagram of separate shift of ceilings and footings.", + "texts": [ + " Tilting of the ceiling, independent of the footings, takes place because the jacks for shifting the support are not connected rigidly with the ceiling, but rather are connected through sliders that can move freely in guides that are rigidly attached to beams of the shield ceiling. The same as the ShchR.P, the ShchK support system can move in two modes: - - by frontal movement of the entire support due to pressure of the bypassed rock (\"kick\"), Fig. 1; - - by separate movement of the ceilings and footings due to the pressure of the bypassed rock and the forces of the shifting jacks, Fig. 2. In the first case, coal is cut out initially on the floor of the bed and in the middle part. The shield ceiling is supported on the coal mass at the roof of the bed, and the shifting jacks are in the extended position. Then, coal is cut from the roof of the bed by the drilling and blasting method, simultaneously along the entire length of the working face, and the shield system is moved along the dip of the bed under the influence of the pressure of the bypassed rock. As shown by observations on models and in mines, with a dip angle of the bed greater than 60 \u00b0, the tilting moment exceeds the force of resistance of the support to tilting; and with a rigid connection between the ceiling and the footings, their tilting part begins to break away from the floor of the bed" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002590_0022-2569(71)90034-6-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002590_0022-2569(71)90034-6-Figure2-1.png", + "caption": "Figure 2. Cylinder traced by point A4.", + "texts": [ + " A point on link 3 can lie anywhere on a surface when both revolutes 12 and 23 are perfectly free. The point Aa, at a distance b from the axis of 23 (as shown in Fig. 1) must be somewhere on the surface of a circular torus whose pr imary and secondary radii are respect ively q and b. Nex t suppose a link 4 to be connected to the same base 1 through acyl indr ica l pair 14. A point A4 at a distance r from the axis of 14 must lie somewhere on the surface of a circular cylinder, of radius r, axis 14 (see Fig. 2). Axis 14 is shown intersecting 0z orthogonally in P and lying parallel with 0)': 0P = p. In general this cylinder intersects the torus in a spatial closed curve that has two separate branches. If Aa and At are made to coincide at the centre A of a spherical joint 34, a closed-loop linkage is formed, and A is then constrained to follow one branch of this curve of intersection. By suitably choosing the proport ions of the torus and the size and position of the cylinder, one branch of the curve of intersection can be made to have third order At the points E and G the spatial curve of intersection has stationary zero curvature; at points F and H it has stationary curvature of radius r" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002755_tvlsi.2012.2227848-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002755_tvlsi.2012.2227848-Figure8-1.png", + "caption": "Fig. 8. Multifin FinFET (a) bulk and (b) SOI structures. Dielectric regions are not shown.", + "texts": [ + " It also permits iterative optimization for a large number of layouts in a practical timeframe. We harness the setup in Fig. 7(b) to analyze multifin multigate FETs and 6T multigate SRAMs in subsequent parts of this section. Owing to the width quantization property, multigate FETs with large electrical widths need to have multiple fins. We synthesized multifin FinFETs using the bulk and SOI FinFETs generated earlier at the 22-nm/14-nm/10-nm nodes. They consisted of four fins each, with shared raised source/drain epiregions that are via-contacted and connected using metal-1, as shown in Fig. 8(a) and (b). We varied the fin pitch, FP, which is the distance between the centers of consecutive fins, and computed the parasitic (FEOL+BEOL) capacitances for each layout using the setup described in Fig. 7(b). From Fig. 9(a), we can see that the trends in CDRAIN,TOT are in stark contrast to the single-fin results in Section III. While moving from SOI to bulk FETs, there is a 11.5%, 10.8%, and 8.8% increase in CDRAIN,TOT for the 22-, 14-, and 10-nm nodes, respectively, which can be attributed to the shared drain-to-bulk fin capacitances in bulk FETs" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000022_robot.1999.770392-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000022_robot.1999.770392-Figure1-1.png", + "caption": "Fig. 1: An overview of the developed Multiple Active Antenna", + "texts": [ + " One big advantage is that the contact localization can be achieved by one simple rotating motion, irrespective of the slope of the object. In this paper, we begin by explaining the basic working principle of the sensor. We demonstrate that two beams make interaction only for a particular case where the slope is exactly perpendicular to the motion plane of two beams, while it is hard to see in the actual utilization. Then, we explain how to obtain the contact angle of each beam. Finally, we show a couple of experimental results to confirm the validation of the idea. 2 Sensing principle Fig.1 shows an overview of the developed Multiple Active Antenna which is composed of two parallel piano wires with the length of 150 x 10-3[m] and diameter of 0.9 x 10-~[m], an one-axis moment sensor, DC servo motor, and an optical encoder to measure the rotation angle of the motor. The two beams are equipped on a platform which is rotated by the motor. They are implemented in such a way that they may rotate in the same searching plane. The torque acting on the beams is measured through the moment sensor installed at the root of the beams", + " Although the peak of the residual provide the exact boundary for an ideal case, it does not for an actual case. This is because the peak easily shifts due to the noise included in the moment sensor output. Based on this consideration, we take the following way for computing the boundary point. We set a threshold for the residual as shown in Fig.10 and execute a line fitting for the data whose residual is less than the threshold. Finally we compute the intersection between each line. 5 Experiments for contact localization The developed system shown in Fig.1 is utilized for examining the working principle discussed in previous sections. A stainless steel column with the diameter of 15 x 10-3[m] is utilized as an object. The motor shaft is rotated with the rotational velocity of 20[deg/sec]. After the first beam makes contact with the object, the motor shaft is rotated until the moment sensor output achieves to a predetermined value. Fig.11 shows experimental results when the object is perpendicularly placed against a horizontal plane, where the circles represent the average of experimental results and the solid line denotes the line on which the experimental points should appear" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003029_2010-01-0639-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003029_2010-01-0639-Figure8-1.png", + "caption": "Fig. 8. Accelerance of the six-axis load cell (axis cell direction y) (.a). FE model of the sensor (.b).", + "texts": [ + " An experimental modal analysis of the sensors has been SAE Int. J. Mater. Manuf. | Volume 3 | Issue 1 295 performed to identify the accelerance of the system, defined as in equation 1. The structure has been excited via a PCB Piezotronics 086C5 dynamometric hammer and the accelerations have been sensed via a tri-axis Br\u00fcel & Kj\u00e6r DeltaTron\u00ae 4506 accelerometer. The obtained transfer functions have been compared with accelerance response of an elastic element such as for the suspension frame (see Sect. SUSPENSION FRAME COMPLIANCE). The graph in Fig. 8.a shows an experimental accelerance amplitude (load cell axis direction) compared with inertance (eq. 1) belonging to stiffness systems. The experimentally identified load cell stiffnesses1 are shown by the data in Tab. 2 and compared with the stiffnesses calculated by means of the FE model (Fig. 8.b). Tab. 2. Load cell stiffness. Tab. 2, Load cell stiffness. For frequencies higher than 300 Hz, the effect of the load cell natural frequency is relevant (the experimental curve increases, while the spring element does not have any resonance peak). Two different types of indoor test for suspension systems linked to the connecting structure above RuotaVia drum can been performed \u2022 Cleat test: where the wheel passes over an obstacle at different drum tangential speeds (simulated vehicle speed) up to 100 km/h to characterize the vibration performance \u2022 Flat surface test: speed ramp test from 0 to 250 km/h: the purpose of this test is to evaluate the tire/wheel unbalance effect" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000922_amm.821.229-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000922_amm.821.229-Figure1-1.png", + "caption": "Fig. 1: General arrangement of the journal bearing model used: a) Floating ring bearing model scheme, b) General bearing dimensions", + "texts": [ + " Generally, there are two basic types of journal bearing used in the design of modern turbochargers \u2013 fully floating ring bearing and non-rotating (semi-floating) ring bearing. Since the fully floating ring is more difficult to model and has more issues with oil film stability, the rational conclusion is to model this bearing first. If it works properly, the model can be altered quite simply to represent the semi-floating ring bearing. The fully floating ring bearing consist of three main parts, as shown in Fig. 1 a). The shaft lies in the housing bore (sleeve) and between these two parts the floating ring is inserted. The shaft is separated from the housing bore by two oil films \u2013 inner and outer oil film. Therefore, there are two parameters describing the shaft eccentricity \u2013 inner eccentricity (ei) and outer eccentricity (eo). The overall eccentricity (e) can be defined, which represents displacement of the shaft relative to housing bore. All important bearing parameters, shown in Fig. 1 b), have to enter the solution. All rights reserved. No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of Trans Tech Publications Ltd, www.scientific.net. (#527184154, Link\u00f6pings Universitetsbibliotek, Link\u00f6ping, Sweden-05/01/20,14:03:03) The hydrodynamic pressure and consequential forces and moments are numerically calculated from the modified Reynolds equation, which is based on the modified Navier-Stokes equation and continuity equation transformed for cylindrical shapes of the bearing oil gap" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000084_6.1992-1082-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000084_6.1992-1082-Figure5-1.png", + "caption": "Figure 5. Thermal Expansion Joint Concept", + "texts": [ + " Ifthe wing were designed as a single unit it would thermally expand over 4 inches (if titanium structure) along the root chord during the cruise portion of flight. The accumulation of these displacements would result in a large residual thermal stress distribution. By breaking the inner wing box into smaller units of about 240 inches (maximum) and fixing each unit at its midpoint, each cell is able to expand freely in the fore and aft directions. Spanwise, each cell is connected by thermal expansion joints (Figure 5) . which allow the cells to grow longitudinally, yet still canyrunningloadsandmaintainthe torsional rigidity ofthe entire box structure. Wing bending loads are carried through the fuselage in two different ways: The majority of the load is carried through a continuous beam spar (Figure 6 ) located at the center of the cell, and the remainder of the bending loads are transmitted as force couples through the huss spar (Figure 7) attachment lugs. The loads then travel through a wing box type structure made up of the lower fuselage and an upper wing skin that continues through the fuselage" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000452_imws-bio.2013.6756225-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000452_imws-bio.2013.6756225-Figure1-1.png", + "caption": "Fig. 1. Geometry of the proposed antenna.", + "texts": [ + " Earlier a compact antenna with a full ground plane and electromagnetically coupled feed for body area network devices operating in industrial, scientific, and medical (ISM) band at 2.45 GHz has been reported [3]. It exhibits a radiation pattern that is along the body surface. An electromagnetically coupled feed was used to partially fill a null that is otherwise present towards the direction opposite to the feed. Proposed antennas provide maximum coverage along the body surface. Further, antenna performance under bending conditions is investigated and it was noted that the operating frequency band can be controlled by tuning the stub. II. ANTENNA GEOMETRY Figure 1 shows the geometry of the proposed antenna with dimensions. It is to be fabricated on 14 x 80 mm 2 FR-4 substrate with a dielectric constant of 4.4 and a height of 1.6 mm. This Antenna consists of two radiating elements and a full ground plane. Full ground plane is considered to reduce radiation toward the body that can potentially harm human tissues. Radiating elements consist of two rectangular strips each having a length of 30 mm and a width of 6 mm. They are electromagnetically coupled to a microstrip transmission line feed with dimensions of 13 mm x 2" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003887_iemdc.2015.7409094-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003887_iemdc.2015.7409094-Figure1-1.png", + "caption": "Fig. 1. Optimal spoke type design, 2D topology.", + "texts": [ + " In this paper, a high performance ferrite based spoke design obtained from a multi-physical optimization is introduced. Furthermore, the optimal material for the rotor support from structural, magnetic, thermal and cost point of view is assessed and proposed. As part of an electric vehicle development project, with the major requirements presented in Table I, a spoke type motor with ferrite magnets and distributed windings has been designed, following a multi-objective multi-physical optimization procedure. The two dimensional (2D) topology is shown in Fig. 1. To cope with the structural requirements at high speed, the rotor is composed of two parts which are joined via a fir tree root, Fig. 1. The performance in terms of peak torque at base and top speed, as well as the effects from the chosen one slot pitch skew, are shown in Fig. 2; the 1 per unit torque corresponds to the peak transient torque, and is equal to 270 Nm. P AND VIABLE MATERIAL OPTIONS To achieve the optimal fir-tree design in Fig. 1, a series of structural-magnetic optimizations have been made, using 2D FE tools; the objective was to control the peak stress below the tensile and fatigue limits of the rotor pole and rotor support materials. Schematic illustrations of a tensile stress-strain curve and fatigue stress-life (S-N) curve are shown in Figs. 3(a) and 3(b), respectively. Some of the early generations of fir tree design are shown in Fig. 4, where some optimization outcomes such as increasing the bubble cut-out radius, or adjustment of the circumferential teeth height are illustrated" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000999_bf00987123-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000999_bf00987123-Figure1-1.png", + "caption": "Fig. 1. The elementary multibody system.", + "texts": [ + " A computer code has been implemented and its effectiveness has been verified in some test cases. In developing this work a particular notation has been used. It requires the definition of some matrix operators which are naturally related to rigid body dynamics. Such elements are defined in the Appendix. In this section the kinematic relations between two constrained bodies will be introduced for a particular class of constraints: rotational and embedded constraints. To this end, consider the elementary multibody system in Figure 1. Let us call: O the origin of an Inertial Reference Frame (IRF), P1 and P2 two reference points on the bodies chosen as origins of two Embedded Reference Frames (ERF), Q1 and Q2 two points on the bodies on which acts the constraint. If we represent the angular position of the two bodies by means of the finite rotation vectors p~ and P2 and the relative angular position by means of the finite rotation p which rotates the first ERF onto the second, a kinematic equation can be set up: n(p2) = n ( p ) n ( p I ), (1) R ( p I ), R ( p 2) and R (p) being the rotation matrices associated to Pl, P2 and p respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003378_bf02765177-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003378_bf02765177-Figure1-1.png", + "caption": "Fig. 1. Diagram of frontal simultaneous shift of support along the roof and floor of bed (\"kick\").", + "texts": [ + " The kinetic scheme of the Sheh_K support differs from that of the known ShchRP support system in that it allows tilting of the shield ceiling on the face under the influence of pressure of the bypassed rock, while the footing remains immobile relative to the floor; as a consequence, the support remains in a stable state regardless of the dip angle of the bed. Tilting of the ceiling, independent of the footings, takes place because the jacks for shifting the support are not connected rigidly with the ceiling, but rather are connected through sliders that can move freely in guides that are rigidly attached to beams of the shield ceiling. The same as the ShchR.P, the ShchK support system can move in two modes: - - by frontal movement of the entire support due to pressure of the bypassed rock (\"kick\"), Fig. 1; - - by separate movement of the ceilings and footings due to the pressure of the bypassed rock and the forces of the shifting jacks, Fig. 2. In the first case, coal is cut out initially on the floor of the bed and in the middle part. The shield ceiling is supported on the coal mass at the roof of the bed, and the shifting jacks are in the extended position. Then, coal is cut from the roof of the bed by the drilling and blasting method, simultaneously along the entire length of the working face, and the shield system is moved along the dip of the bed under the influence of the pressure of the bypassed rock" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000016_9780470276280.ch31-Figure31.6-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000016_9780470276280.ch31-Figure31.6-1.png", + "caption": "Figure 31.6 Spinnerets with various flow angles (Cao et al., 2004).", + "texts": [ + " It is revealed that nodules in the outer skin changing from random arrangement to obviously tidier align along the direction of dope extrusion when the shear rate is increased. In addition, both nodule sizes in the fiber spinning and transversal directions decrease with increasing shear rate, possibly because of chain disentanglement and thermodynamically favored. Furthermore, the roughness of the outer surface of hollow-fiber UF membranes decreases with an increase in shear rate. The combined influence of elongation and shear rate induced by the geometry of spinnerets on membrane performance for gas separation has been studied as illustrated in Figure 31.6 (Cao et al., 2004). The flow profiles of dope solution and the elongation and shear rates at the outermost point of the outlet of spinnerets can be simulated by the computational fluid dynamics (CFD) model. The preliminary conclusion indicates that the elongation rate has more contribution portion in permselectivity than in permeance, while the shear rate has more contribution portion in permeance than in permselectivity. For UF membranes, experimental results suggest hollow fibers spun from a conical spinneret have smaller mean pore sizes with larger geometric standard deviations, thus exhibiting lower water flux and greater solute separation than hollow fibers spun from a traditional straight spinneret (Wang et al" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002956_09544100jaero196-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002956_09544100jaero196-Figure1-1.png", + "caption": "Fig. 1 Spacecraft model with single-axis rotation", + "texts": [ + " To realize this approach on the flexible spacecraft, a modelling analysis is first performed. Next, the analysis of the proportional\u2013derivative (PD) and shaped command control strategies is conducted to recommend the proposed method. Then, numerical simulations performed on a five-mode model of the flexible spacecraft demonstrate the efficacy of the method. Finally, the paper is completed with some concluding comments. The slewing motion of a rigid hub with two appendages attached to the hub is graphically presented in Fig. 1. The rotational motion without any translation of the centre of mass of the whole structure only is considered in the current paper. Define the OXY and oxy as the inertial frame and the frame fixed on the hub, respectively. The attitude angle \u03b8 denotes the relative motion between these two frames. Denote w(x, t)as the flexible deformation at point x with respect to the oxy frame. It is assumed that the control torque is applied to the rigid hub only. The governing differential equations of motion can be derived from the extended Hamilton\u2019s principle which states that [10] \u222b t1 t0 (\u03b4L + \u03b4W ) dt = 0 (1) where L = T \u2212 U represents system Lagrangian as a difference between kinetic and potential energy of the system, and \u03b4W = \u03b4\u03b8u is the virtual work by nonconservative control torque applied at the centre body" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002699_icorr.2015.7281298-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002699_icorr.2015.7281298-Figure3-1.png", + "caption": "Fig. 3. The appearance of the PAU in developing", + "texts": [ + " In [24], tipping lever is explained as: \u201cThe tipping lever extends from the bottom of the frame and is designed to make it easier to move the wheelchair over obstacles, such as curbs. The person pushing the wheelchair will put weight on the tipping lever, which causes the wheelchair to tip backwards\u201d. It exists in almost all of manual wheelchairs as well as it is at the rear of a manual wheelchair, the most convenient position to access by attendant. For this two reason, we determine our PAU should be attached to the tipping lever. As shown in Fig.3, two clamps are used to fasten the PAU to the two tipping levers on the two sides of the frame of the wheelchair, respectively. In this way, the PAU can be very easily attached to and removed from the manual wheelchair by any untrained persons including the attendants only with their bare hands but without the help of any tools . In addition, since the tipping levers are at the rear of the wheelchair, the PAU is at the rear of the wheelchair, too. This is the most convenient place to mount or remove the PAU to or from the wheelchair", + " To make the mechanism of our novel PAU as simple and cheap as possible, we do not adopt an automatic active suspension system to keep a constant contact of the drive wheel to the ground. In stead, a mechanical lever is used to manually adjust and lock the height of the drive wheel so that the drive mode and free mode can be easily switched while the entire structure of our PAU is kept very simple. Fig. 5 shows the PAU is in lifted state in which the drive wheel of the PAU does not contact the ground and the wheelchair is operated as a normal manual wheelchair. In our PAU, we use shield lead acid battery, that is cheap and heavy. Moreover, as shown in Fig.3, the battery pack is at the front of the PAU. This is to increase the exerted normal force of the drive wheel (in-wheel motor) to the ground so that the grip performance and traction force of the PAU are enhanced and the slippage between the drive wheel and the ground could be greatly prevented. The ultimate purpose of a PAU is to help an attendant to push the wheelchair. Therefore, its most basic function is to control the speed that is usually pre-selected via the handle control panel or something like that to prevent the speed variance mainly caused by load or environment condition" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003293_icelmach.2016.7732546-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003293_icelmach.2016.7732546-Figure1-1.png", + "caption": "Fig. 1. 2D FEA Magnetic flux density and flux lines at 86% rated load", + "texts": [ + " Akiror, Arezki Merkhouf, Pragasen Pillay, Claude Hudon, Charles Millet O model this machine according to its magnetic and winding symmetry. Transient time stepped simulations with a time step of 36 s is selected. A fine mesh totaling to 275945 elements is used. External circuits are used to impose the load, include end winding impedances and model the damper bar impedances. Due to the turbine limitations, the only available measured data is at no load and 86% of the rated load. Therefore, the simulations are limited to these two operating conditions. Fig. 1 shows the flux density distribution and flux lines at 86% of the rated load. A large percentage of the stator yoke has a flux of less than 0.4 T. The FEA model of the machine is validated using measured airgap flux at both no load and 86% of rated load. The air gap flux in the machine was measured using two 1 inch diameter, circular, 10 turn coil sensors C1 and C3 located in different positions in the air gap, glued on the stator tooth. Flux density in the air gap was obtained by integrating the induced EMF in the coils" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002960_iembs.2009.5332882-Figure12-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002960_iembs.2009.5332882-Figure12-1.png", + "caption": "Fig. 12. MATLAB simulation of foot-plate movement for our gait trajectory guiding device.", + "texts": [ + " Now following the %duration of different phases of a complete walking cycle suggested in [8], time division for movements along two vertical axes is shown in Tables II and III: From the above information we can easily derive the vertical velocity along Z1 and Z2 direction over time as shown in Fig. 11 (Z1 and Z2 were described in Fig. 8). On the basis of the time divisions for horizontal and vertical movements along X, Z1 and Z2 axes shown in Table I, II and III, we have simulated our proposed person specific gait algorithm in MATLAB. Fig. 12 depicts motion of the foot-boards during backward and forward movements. In this figure we have also shown the co-ordinate points for ankle, knee, hip and base rib joints derived from sample data of a normal subject, as according to our first hypothesis we are expecting that the foot-boards movement will also help the patients to follow an ideal trajectory and angle for these joints. The main theme of task specific approach is motor learning through repetitive training of a real and ideal walking practice suitable for individual patient\u2019s body structure" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003205_ecce.2014.6954109-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003205_ecce.2014.6954109-Figure8-1.png", + "caption": "Fig. 8: Coupled EM/thermal optimal machine design", + "texts": [ + " Six cores out of the 8-core desktop are dedicated in this optimization, with one core running the main MATLAB code and the other five cores conducting FE analyses in parallel. This configuration can accelerate the computation speed approx. 5 times higher than with a singlecore computer. The optimization was run and converged after the 50th generation, with a total number of 1,500 designs evaluated. The best design has a total active mass (stator and rotor) of 13.5 kg. It can continuously produce the required torque at its maximum current density of 9.84 A/mm2. The optimized machine using the coupled EM/thermal model (Fig. 8) has a mass reduction of 7.17 kg compared to the machine optimized previously using the EM-only model (Fig. 7), in which the current density is a fixed value of 4.6 A/mm2. The key parameters of these two machines are listed in Table III, demonstrating a significant improvement in torque density for the machine using the coupled EM/thermal model. A large share of this improvement can be attributed to the action of the coupled model to significantly raise the current density within the maximum temperature limits of the winding insulation, set at 155\u2070C for this exercise (assuming Class F wire insulation)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002367_iros.2009.5354127-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002367_iros.2009.5354127-Figure3-1.png", + "caption": "Fig. 3 When the Xm axis motor is rotated", + "texts": [ + " Motor bearings We now focus on and explain the motor bearing supporting the Xm axis motor shown in 1. This motor bearing is formed by the great circle of the rotating ball and the concentrically curved hollow cylinder component. Also, for the surface on the side of the Xm axis motor, and the surface on the side of the rotating ball, a long hole is formed in a longitudinal direction with the insertion width of the hollow motor output axis. Also, a component known as a slider is incorporated into the hollow part of the motor bearing (see Fig.3). There are a number of bearings located in the slider. The dimensions of these bearings are set so that there cannot be any gap between the inner surface of the hollow part of the motor bearing, the inner surface on the motor side, and the inner surface of the rotating ball. The slider ensures that the internal components do not jiggle in any direction. In this way the slider, by means of the rolling motion between each inner surface of the internal components yielded by the bearings, makes it is possible for the inner part of the motor bearing to slide with low friction in a longitudinal direction while keeping the rotating ball at a constant distance. The anterior end of the Xm axis motor is fixed via a support plate coupled with the slider. In this case, the output axis extending from the anterior end of the Xm axis motor passes through the slider and is joined to the rotating ball. According to the structure stated above, there is no jiggling in the Xm axis motor between the slider and the hollow part of the motor bearing, so the rotation is regulated around the axial direction of this part itself (see Fig. 3). Moreover, due to the slider there can be no reciprocation along the motor bearing, so among the other 2 orthogonal axes the Xm axis motor can turn around the Y0 axis (see Fig.4). With this type of motor bearing, the ends are fixed to the rotation axes, and there is axial support in such a way that rotation is possible at the support-end located on the Z axis of the clamp. There are two bearings installed internally in the support-end, and by aligning the two bearings in the axial direction of the rotation axis, even when an external force is applied to tilt the rotation axis the two bearings act together to support the rotation axis and prevent it from tilting" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002575_125906-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002575_125906-Figure4-1.png", + "caption": "Figure 4. Electromagnetic drive principle of two GMM. (a) Hollow cylindrical GMM\u2019s drive principle. (b) Solid cylindrical GMM\u2019s drive principle.", + "texts": [ + " The giant magnetostrictive structure composed of a solid cylindrical GMM is as shown in figure\u00a03(a), Terfenol-D rod is magnetized by the working current in the excitation coil, and it will occur as a magnetostrictive effect. The pre-tightening structure is as shown in figure\u00a03(b), the output force of the GMM structure is transferred to nuts by hinge-levers. Electromagnetic field of giant magnetostrictive structure can be analyzed by finite element method [12\u201315]. GMM is magnetized to generate magnetostriction, its operating principle is as shown in figure\u00a04. The hollow cylindrical GMM is coaxial with the screw, its radial thickness is 7 mm and axial length is 60 mm, the diameter of the screw is 40 mm, the air gap between the screw and the yoke is 7 mm. There\u2019s an aluminum sleeve between the screw and the GMM, its radial thickness is 3 mm. The hollow cylindrical GMMs magnetic field lines distribution, magnetic flux density and magnetization intensity are as shown in figures\u00a05(a), 6(a) and 7(a). The solid cylindrical GMM (\u03c6 14mm \u00d7 60 mm), yoke iron and air gap constitute a finite element model" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000092_8.247750-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000092_8.247750-Figure3-1.png", + "caption": "Figure 3 shows the direction of the main beam of the array. Mutual couplings between the array and the MSA were measured for the beam scan within a range of -90\" I 8 I 90\". The measurement was carried out at one degree step, and time fluctuation of the measured signal at every 8 was averaged utilizing the function of the network analyzer.", + "texts": [ + "j)f - (s^*f)$ (25) radiating elements of the other array A , because interference occurs between them. = - p 2 7 (26) SAi is calculated as follows: SAi = C S n i Z n , (33) { ( r 2 - r , ) .j>\u2019 n e A _ (27) I n = j \u2019 e j 4 n n 3 I In12 = 17 (34) s^.f 1 - = 1, n e A 2(d, + d2I2 where n is the number of radiating elements in the array A, the sum Cn E A runs over all the radiating elements in A , Sni is the mutual coupling between radiating elements ( r 2 - r , > . j s * y Gz (28) dl + d2 . MIYASHITA et al.: ANALYSIS OF ANTENNA COUPLING BETWEEN ARRAYS 1245 14.11 A LECTIONOF y - IARRAY I Fig. 3. Antennas on a polyhedron used in the experiment. MSA ARRAY MSA PHASE SHIFTERS 128 DIVIDER NETWORK ANALYZER Fig. 4. Configuration of the experiment. n in-the array A and the radiating ejements i in the array B, I, is an exciting amplitude, and 4, is an exciting phase of n. In the above expressions we assume that the exciting power of the array A is normalized by a unit power. IV. EXPERIMENTAL VERIFICATION Figures 3 and 4 show the configuration of the experiment. The scatterer is a rectangular parallelepiped which has a size of 36h x 36A X 554 where A is the operating wavelength" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002460_5138-ms-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002460_5138-ms-Figure2-1.png", + "caption": "Fig. 2 Basic Dimensions of Nodes", + "texts": [ + " Overlapped Cast Steel Node The term \"node\" is defined as a structural element connecting incoming members in a space frame structure. A node consists of one main member known as a \"chord\", and several sub-members known as \"stubs\" which are connected to brace members. The term \"overlapped node\" refers to a node in which its stubs are overlapping each other. Typical types of overlapped nodes and the terminology of the dimensions are shown in Figs.1 and 2. The overlapping ratio \" defined by f2 s/ f2 c is very important measure to describe structural characteristics and static strength of the over Iapped node. In Fig.2, the dimension ''min.350\" is determined from welding workability. Diaphragms between stubs of overlapped cast steel node are removed from casting technology point of view. 2. Description.Qi. FE Analysis The structural characteristics of overlapped cast steel nodes are determined on the basis of a lot of numerical experiments conducted with finite element analysis. The computer program for structural analysis is the latest version of MSC!NASTRAN. 2.1 Analysis ~del A typical mesh layout of a finite element model is shown in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002597_j.ijthermalsci.2007.06.020-Figure16-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002597_j.ijthermalsci.2007.06.020-Figure16-1.png", + "caption": "Fig. 16. Configuration and parameters of the electric submersible pump flat power cable.", + "texts": [ + " In addition, it is worth mentioning that single-phase operation generates a pulsating magnetic field according to Eq. (9) and hence higher localized corrosion on the motor casing, see Fig. 1. OPERA-2D is also used to simulate an oil well cross-section with an electric submersible pump cable touching its tubing. This cross-section consists of several layers with different materials which are shown in Fig. 15 and the dimensions are summarized in Table 3. The used electric submersible pump flat power cable is a 3-phase flat cable with the shown configurations in Fig. 16. The different motor operating conditions are shown in Table 4 with Table 4 Simulated line currents in A (rms) Conductor No. Balanced case Single phasing 1 100 0\u25e6 0 2 100 120\u25e6 173.2 0 3 100 240\u25e6 \u2212173.2 0 Fig. 17. Configuration of the oil well with power flat cable. their operating current flowing in the cable for each condition. These results describe the distribution of the magnetic-flux density (B) and the current-density (J ) at the well casing and its tubing. Different oil well models have been simulated and studied" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001187_intmag.2003.1230623-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001187_intmag.2003.1230623-Figure2-1.png", + "caption": "Fig. 2. Radial traction iluctuiltiim duc to rotor \\Intic ccccntricity: (;I) Rotor ecccntricitv (b) Ridi.tI tmctiun fluctuiltiun", + "texts": [ + " Using the linile element analysis, this paper analyLes unbalance force according to rotor eccentricity and deals with the comparisons of unbalance force of two types, radial and Halbach magnetized rolors of high spccd slo~kss P M machinc. Figure I shows a 4-pole, 3-phase slot-less permanent magnet motor, with (a) conventional radial magnetized permanen1 magnets. and (h) a multi-pole Halhilch inagnetired miigneti. 'The lopologicr were designed as I kW permanent m~~giiet molor/gencrator wilh Ihc raled speed 0140.000 rpnl. Figure 2 shows the geometric configuration of permanent magnet machine with rotor eccentricity and unbalance force fluctuation to be analyzed. When pernunent magnet rotor rotates around the axis through 0, at the angular velocity m,, Fig. 2(b) shows the radial traction fluctuation. The fluctuation of radial trdction causes vibrslion and noise. Figure 3 shows radial traction per airgap flux density according 10 rotation angle e, and fluctuation of radial traction at the eccentricity ratio ~=0..5. respectively. It can be seen that the Halbach maglietized topology has a smaller iluctuation of radial force than conventional radial magnetized topology. The permanent magnet mutor with Halbach magnetization has a lvwer fluctualivn uf radial traction ahaut 82" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003283_870182-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003283_870182-Figure8-1.png", + "caption": "Figure 8", + "texts": [ + "3 TEST PROGRAM A simple test arrangement shown in Figure 9 was used to establish deflection and burst strength of the rad lators. Each rad lator was SUbmerged under water and pressurized with air. Pressure was 1ncreased from zero psig untIl it reached a point where leaks occurred at the tank to header jotnts. Pressure was increased in Increments of 5 psig and deflection measurements were taken at the locations shown. (Refer to Figure 9.) After leaks occurred. air More accurate ana I ysls was per formed by mode 11 n9 the thre~-dlmensiona I Quarter sectlons of the two dss1gns as shown in FIgure 8. Boundary conditIons were taken 4 pressure was released. The radIator was then completely filled wtth water, and air pressure was reapplled as prevIously descrIbed. Deflection measurements were recorded until the plastIc tanks eIther blew off the header, or leaks developed to a degree where a1 r pressure coul d no longer be Increased. DESCRIPTION OF RADiATORS TESTED A total of four radIators were burst tested as prevlously descrlbed. Two radiators were buIlt wlth the \"HSK\" tank attachment method" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003283_870182-Figure10-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003283_870182-Figure10-1.png", + "caption": "Figure 10", + "texts": [], + "surrounding_texts": [ + "4 pressure was released. The radIator was then\ncompletely filled wtth water, and air pressure was reapplled as prevIously descrIbed. Deflection measurements were recorded until the plastIc tanks eIther blew off the header, or leaks developed to a degree where a1 r pressure coul d no longer be\nIncreased.\nDESCRIPTION OF RADiATORS TESTED\nA total of four radIators were burst tested as prevlously descrlbed. Two radiators were buIlt wlth the \"HSK\" tank attachment method. Radtator number 1 used .040\" thick header material, whIle radIator number 2 used .03211 th 1ck header mater 1a I. The other two radl ators were bu 1\\ t wi th the tabbed header pi ast Ie tank attachment method. Rad 1ator number 3 used .040\" thIck header materIal, whIle radiator 4 used .032'1 thick header material. (Refer to Table 1.)\n3. RESULTS AND DISCUSSION\nSTRI\u00a3TlJlJ\\L ANALYS is\nTo compare the ability ot the jotnts to retain a water tight seal, the relatlve vert1cal displacement between\nthe tank and header, where the gasket Is located, was cal cu I ated. Tab I e 2 11 sts these va I ues. In essence, this dlsplacement relates to the loss ot gasket compression to 1nternal pressurlzatton of the tank.\nThe 3-D rode I results for the linear elastic displacemenT shows that for the same load, the amount of gasket decamp ress Ion for HSK 1s about one-ha I f that for TAB. The joint stiffness against the loss of gasket seal can be defined as the inverse of gasket decompress Ion. The ana Iys 15 shows that the HSK joint is stiffer than the tab joint by 2/3 or lTOre SCCI'om, - Eac~ 2 Ir>c~~' long IOf 4 1~:llC~ TOlall\nGASKET COMPRESSION\n/~ ;0',\n010 f---+--t---+--i>-\u00ab:.--+---j---+---j---+--I /f-\"\n050f--+--+---,~j-=-+--+---t--+--+---t---I1/ -w\n0'\" f-_-+_~j_I'---'-;'j_; __-j-_-+__-j-_-+__+-_-+_-If- ,\", ~ (MO f----j,!:::::-\"I-.,---t--\n\"0 j-...,/4---+_-+-\n/" + ] + }, + { + "image_filename": "designv6_24_0003869_851385-Figure3-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003869_851385-Figure3-1.png", + "caption": "Figure 3. Combined Environments Test Equipment", + "texts": [], + "surrounding_texts": [ + "851385 5\ncombustion of SOFI decomposition products. Aft-dome combustion is sustained by turbulent mixing which provides sufficient oxygen at the SOFI surface to promote burning. Wind-tunnel tests evaluated char recession under co~bined aerodynamic heating and shear, and determined ablation performance at protuberance heating levels. Minitanks are 3 ft diameter X 5 ft high aluminum tanks used to assess TPS performance under repeated cryogenic fill/drain cycles and pressurization. Minitank tests evaluated substrate-primer-TPS strain compatibility, primer and TPS adhesion, ab1ator and SOFI cracking and susceptibility to cryopumping. Design features such as c10seout/ repair configurations and instrumentation wire bending were also evaluated with these tests. To assure that the applied TPS materials did not delaminate and falloff with catastrophic consequences the flexstrain also known as the cryof1ex test was developed. The tester is a single axis device capable of applying any combination of tensile and flexural loads simultaneously in an environmental test chamber which allows the simu1atation of thermal loading (Figure 2). The key element in the TPS qualification was the design- verification tests conducted on 4 ft X 4 ft combined-environments panels. The biaxial test panels were representative of specific critical areas of the external tank structure and TPS design. In these tests, the TPS experiences the timed phased combined effects of cryogenic SUbstrate, biaxial strain, acoustics, ascent heating and pressures. Figures 3 and 4 show a typical combined enviroments test panel and test apparatus. The qualification methodology was to combine the ablation wind-tunnel tests with specific design-verification tests conducted under combined environmental conditions to achieve confidence in the total TPS system.", + "6 851385\n~~TERIAL/PROCESS VERIFICATION\n\"The verification test program for the ET TPS included the wind tunnel tests discussed above together with several cryogenic, radiant heating, and combined environment tests. These tests were designed to verify the TPS integrity under the various predicted flight induced environments. There was no one test (other than flight) which simulated all of the pertinent flight parameters. Confidence in the TPS system was achieved by the successful results of these tests taken together. \"Minitank tests were used to evaluate TPS cryo-strain compatibility, primer and TPS adhesion, and TPS cracking and susceptibility to cryopumping. The minitanks were 3-ft diameter aluminum tanks with TPS applied and were tested under repeated cryogenic fill, drain and pressurization cycles. These tests did not simulate any ascent pressure, heating or acoustics loads. \"A larger 10-ft cryogenic tank was also tested similar to the minitanks to assess any large scale application issues. The 10-ft tanks, like the mini tanks, were tested with LH under repeated cryogenic cycles. The 10-ft tank also included a radiant heat test to assess TPS recession and propellant quality on a large scale tank. Note that the 10-ft tank was the largest scale application of flight type TPS prior to STS-l. \"Radiant heating tests were conducted to verify the TPS recession characteristics under the aft dome environment where heating was due primarily to the exhaust plume radiation and recirculation, rather than aeroheating as simulated in the windtunnel tests. These tests were conducted in two facilities; one simulated radiant heat and acoustics, and the other\nradiant heat and ascent pressure decqy. \"The key element in the TPS qualification was the \"combined environment\" tests conducted on four TPS panels configured to represent the substrate and TPS in specific critical areas on the ET. These panels were subjected simultaneously to biaxial substrate loads, cryogenic backface temperature, ascent heat load (radiant), and either acoustics or ascent pressure. The panels were tested in a thennal-vacuum chamber, and/or in a thermal-acoustic facility depending on the specific test objective. Both facilities employed a large load cell structure which could be programmed to induce biaxial load profiles in either tension or compression. These were used to simulate various degrees of predicted flight substrate loads to demonstrate the TPS structural margin. The panels were cooled with liquid helium to simulate hydrogen tank substrate temperatures and the flight heat loads were simulated by an infrared lamp bank. These unique tests allowed the TPS to be subjected to nearly all the flight conditions (except aeroheating) and on a scale large enough to verify production methods (Bachtel et al., 1983).\"\nFLIGHT RESULTS\nActual flight data from the first six ETs provided flight verification of the predicted thermal environments. Specially designed instrument islands measured aerodynamic heating and boundry-lqyer conditions at 30 locations on the ET sidewalls and 8 locations on the LHZ aft dome. In addition, the temperature of the tank structure was measured at 59 locations. Flight instrumentation was augmented by cameras mounted in the orbiter which photographed the", + "851385\nET during ascent and separation from the orbiter. Photographs confirmed the locations of higher heating regions. Initially the flight instrument data was used to compare the environments predicted with the actual environments. This data together with the photographs confirmed the adequacy of the TPS materials and design. Currently this same data is being used to further reduce cost by further reducing the coverage of the ablator based on the actual temperatures, not the higher predicted heating.\nSUMI4ARY\nSince the start of the project, the External Tank has progressed from development to production. The thermal protection system has changed with changing requirements and improvements in materials and processes (Table 2), however, it is still true to the original concept. Today's changes represent an upgrading and maturing of an already efficient design. Figure 5 shows the current TPS configuration. As the production rate increases, most changes are directed toward\nproducibility improvements and elimination of sole-source dependencies. Increased use of robotics and net-molding techniques will greatly reduce touch labor and material usage. The qualification of NCFI 22-65 for use on areas other than the aft dome would provide an alternative to CPR-488 and could potentially reduce the amount of ablator used. Other efforts are underway to develop cheaper/lighter ablators and/or cheaper foams with improved thermal capability to replace SLA-56l ablator completely. Based on flight and production experience, the already efficient design will continue to be optimized to enhance producibility, consistency and reliability. Continuing concerns include the streamlining of launch operations by minimizing the number of TPS closeouts and simplification of active thermal- control systems (purges and heaters). With the forthcoming launches from Vandenberg Air Force Base (VAFB), a new set of environmental parameters must be considered. Vandenberg's more severe weather conditions and launch profile present new challenges.\n7" + ] + }, + { + "image_filename": "designv6_24_0002584_1999-01-0782-Figure13-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002584_1999-01-0782-Figure13-1.png", + "caption": "Figure 13. Detail of folded spoke construction.", + "texts": [ + " PROTOTYPE FABRICATION \u2013 The second generation prototype was similar to the first in that it was designed and fabricated out of five separate spokes yet it proved to be significantly more difficult to prototype. This was primarily due to the large number of spokes having to be scrapped from excessive cracking and tearing in the bend regions. To fabricate the spokes, an initial \u201chorseshoe\u201d shape was lasercut out of 1040HR steel. The shape was then folded into a channel section and as a final step, two tabs were then folded outward at the end of each spoke. An example of the completed spoke is shown in Fig. 13 below. To assemble the prototype, a fixture was constructed which located all of the spokes for initial tack welding. Once tack welded, the assembly was then removed from the fixture and finish welding was completed. As a final step, the bolt pattern and center pilot diameter were machined. TESTING \u2013 All testing was performed by Goal Automotive Technical Services in El Monte, California. A total of two samples were evaluated on the cornering fatigue test to the SFI 5.22A and SAE J328 test specifications" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002388_2015-01-0090-Figure10-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002388_2015-01-0090-Figure10-1.png", + "caption": "Figure 10. Strain distribution (Von-Mises) in case of front left wheel at the top", + "texts": [ + " In this case, a front left spring experienced its maximum bump travel observed by a zero gap between bump stopper, attached on the axle, and stopping plate, welded to the side rail lower flange. In Figure 9, vertical distances between the wheel centre and level road of the front left wheel at start position (yi) and stop position (yf) were measured. truck (Location 3, lateral direction), in case of Front-Left wheel on ramp stop positions of the acting wheel. The distances from wheel centre to level road at start position (yi) and stop position (yf) were measured. In the simulation, vertical displacement of yf\u2212 yi (Figure 9) was assigned to the left end of the front axle. Figure 10 is a contour plot of equivalent strains on the truck frame predicted by the simulation. For convenience, lateral and longitudinal strains were defined to comprehend our discussion later on. The lateral strain was normal strain measured in a direction along the length of cross members near their ends as showed in Figure 11. The longitudinal strain is normal strain located on the parallel flanges of side rails (U-shaped channel), Figure 12. With difficulty in mounting the strain gauges to a truck due to limit access, it was necessary to attach a strain gauge inside the channel for the top flange and outside the channel for the bottom flange" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000378_28.739003-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000378_28.739003-Figure2-1.png", + "caption": "Fig. 2. Coil shapes for AFPM\u2019s with coreless winding. (a) Trapezoidal coil. (b) Rhomboidal coil.", + "texts": [ + " In multistage machines, only the two external rotor discs must be made of material with good magnetic properties (typically, mild steel), since they are used to provide a return path for the main flux. The intermediate rotors, on the other hand, are used merely for mechanical support of the magnets, so that lightweight nonmagnetic materials (e.g., aluminum) can be used for their construction, thus enhancing the machine compactness and lightness [13]. As in the earlier disc-shaped machines with ironless winding [1]\u2013[6], the winding coils may have the trapezoidal shape shown in Fig. 2(a). This coil shape allows the maximum coil flux linkage, as each coil embraces the entire pole area, but, on the other hand, the coil end windings necessitate a three-tier arrangement and may have significant length if compared with the length of the conductor active sides. Therefore, disregarding the case of machines having a very large number of poles and, thereby, very short end windings, the use of trapezoidal coils negatively affects the value of torque produced per unit of , and this is undesirable for machine applications with high power ratings. As discussed in [12], higher values of torque per unit of can be achieved if the winding coils have the rhomboidal shape shown in Fig. 2(b). In fact, due to the inclined arrangement of the coil active sides, in rhomboidal coils, the end windings are greatly shortened, but with only a small reduction of the coil flux linkage if compared with the conventional trapezoidal-shape coils. In addition to that, the use of rhomboidal coils allows a two-tier arrangement of the end windings with consequent reduction of the winding axial thickness and simplification of manufacturing. The geometry of rhomboidal coils is characterized by the inclination angle and the ratio between the inner radius and outer radius " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002584_1999-01-0782-Figure12-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002584_1999-01-0782-Figure12-1.png", + "caption": "Figure 12. Second generation FEA analysis.", + "texts": [ + " Traditional methods were not considered feasible due to the increasing complexity of the spoke geometry and attachment. To simulate the prototype wheel under a dynamic test, a three-foot shaft was attached to the hub and the entire assembly was meshed as one contiguous part using 20 node polyhedron elements. To simulate the laboratory dynamic cornering test, a transverse load was then placed at the end of the shaft. The wheel was then constrained by fixing the inboard beadseat of the rim and the analysis was run. Resultant Von Mises stresses are shown Fig. 12 below. Stress levels in the range of 12 to 26Mpa were evident on the back and top sides of the spoke which fell in line with the transverse shaft load. Based on the FEA modeling, critical stress regions appeared to be located in the spoke-hub welded joints. Although maximum stress levels were fairly low in general, they presented a challenge for the material selection choice because of their location in the spoke. This was because improved mechanical properties resulting from forming would not take place in critical regions of the so a medium carbon steel was selected for fabrication" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003657_ccece.2013.6567714-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003657_ccece.2013.6567714-Figure8-1.png", + "caption": "Fig. 8. Bicycle trajectory, in blue and start point (-10,-5), tracking counterclockwise path, outlined in red, consisting of straight lines and circles.", + "texts": [ + " Convergence is attained in less than one revolution of the path. Finally, a path consisting of multiple line and circle segments was tested. To construct the path, waypoints are provided (or, as in this case, randomly chosen) and lines are drawn between each pair of waypoints. Each corner is then rounded by a circle, of given radius, so that the bicycle is not subjected to any very large instantaneous changes in direction. The result of the bicycle travelling such a path, starting and ending at the origin, is shown in Figure 8. Figure 9 shows how much the bicycle deviates from the path along its journey. It can be seen that after the bicycle is on the path it deviates by less than 1.15m. This is a reasonable value for the bicycle and path parameters; it can be reduced by decreasing the bicycle speed v or by increasing the gains h1 and h2, but in the latter case the bicycle must start closer to the path. Lastly, the input torques are shown in Figure 10. A systematic multi-loop control approach has been developed for the Whipple Model of a bicycle using linear control techniques to simultaneously achieve stabilization and path tracking" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003395_1.1778720-Figure15-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003395_1.1778720-Figure15-1.png", + "caption": "Fig. 15 The best-fit displacement field when limited members are allowed to actuate for the target displacement field wd \u00c4A0z2ez", + "texts": [ + " (24) Here, B*\u2020 is the Moore-Penrose generalized inverse of B*. The vertical displacements of the solid face nodes are also easily calculated: w\u0303 i5Bi j*e\u0303 j* . (25) Reconsider the target displacement field that corresponds to constant curvature: wd5A0z2ez . (26) Now, however, assume that only members of the Kagome plane in the row corresponding to z'Lz/2 can actuate. These members are located in the middle of the periodic cell, as shown in Fig. 14~a!. The Moore-Penrose best-fit displacement field is shown in Fig. 15. Note that the structure displays only local curvature along the line z5Lz/2. Reconsider also the target sinusoidal displacement field: wd5A0ez sinS 2pz Lz D . (27) Here, however, assume that only the selected members of the Kagome plane are actuated corresponding to those aligned in rows having z'Lz/4 and z'3Lz/4, as shown in Fig. 14~b!. These members lie within the zones of maximum curvature magnitude of the target displacement field. When the Moore-Penrose analysis is run under these conditions, the resulting displacement field is displayed in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000855_0954406220936732-Figure5-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000855_0954406220936732-Figure5-1.png", + "caption": "Figure 5. Distribution of limbs coordinate system.", + "texts": [ + ",5) are the coordinates of the origin of the reference coordinate system {Op} in the local coordinate system {Oj} in terms of ith flexible series limb. ri1 \u00bc 4rs \u00fe lsin \u00fe L1 lcos 0 T ri2 \u00bc 2rs \u00fe lsin \u00fe L1 lcos 0 T ri3 \u00bc 2rs \u00fe L1 0 0 T ri4 \u00bc L1 0 0 T ri5 \u00bc 0 0 0 T \u00f013\u00de Stiffness modeling of flexible parallel force sensor When calculating the stiffness matrix of the flexible parallel force sensor, the thickness of the moving platform is neglected. The sensor is symmetrical, and the establishment of each series limb coordinate system and moving platform reference frame are shown in Figure 5. {O} is the reference frame of moving platform, {Oi} (i\u00bc 1, 2, . . . , 12) are the local coordinate systems of the reference points (i.e. the geometric centers of the end sections) of each series limb and heavyload-bearing limb. {Oi} (i\u00bc 1, 2, . . . , 8) are distributed along the circumference of the radius R. xi-axis is parallel to X-axis. represents the angle between the vector OO1 and Y-axis. The direction of yi-axis is the projection of series limb i in the plane of YOZ, which is from the moving platform to the fixed platform. The local coordinate systems {Oi} (i\u00bc 9, 10, 11, 12) of the heavy-load-bearing limbs are distributed along the circumference of radius R\u2019 at equal intervals of p/2, the acute angle between Y-axis and OOi (i\u00bc 9, 10, 11, 12) is p/4. The coordinate axes of each local coordinate system are parallel to the reference coordinate system of the moving platform, and the directions are shown in Figure 5. According to the static balance condition and equation (11), the stiffness matrix of the sensor with hybrid limbs can be obtained K \u00bc X12 i\u00bc1 J Ti KipJ 1 i \u00bc J T1 J T2 J T12 K1p K2p . . . K12p 2 66664 3 77775 J 11 J 12 .. . J 112 2 666664 3 777775 \u00f014\u00de where Kip are the stiffness matrices of the ith flexible limb, Ji (i\u00bc 1, 2, . . ., 12) are the pose transformation matrices between the ith local coordinate system {Oi}, K is the global stiffness matrix of the force sensor with hybrid limbs, and the reference coordinate system {O}, Ji are as follows J1 \u00bc O O1 R S r1\u00f0 \u00de O O1 R O3 3 O O1 R \" # , J2 \u00bc O O2 R S r2\u00f0 \u00de O O2 R O3 3 O O2 R \" # , " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0003746_rem49740.2020.9313869-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0003746_rem49740.2020.9313869-Figure2-1.png", + "caption": "Fig. 2: Disassembled hybrid actuator.", + "texts": [], + "surrounding_texts": [ + "The actuator is designed to carry a defined load at reference air gap only by permanent magnet forces. If the air gap changes due to additional load, a current is applied to set the air gap back to its reference value (see sec. IV). The applied permanent magnets of the hybrid actuator are NdFeB magnets with a minimum remanence flux density of B R = 1.17T. The height of the magnet is h pM = 1.0mm Authorized licensed use limited to: University of Brighton. Downloaded on June 20,2021 at 18:58:55 UTC from IEEE Xplore. Restrictions apply. (17) (16) (18) (20) G (8) - K . TN,el 8+ 1. 11 . 1 \u00b0 el - P el T T + 1 T + 1N el 8 l8 s 8 The controller gain K P,el is calculated such that the absolute value of the reference transfer function is kept close to one for the largest possible frequency range, resulting in K _ TN L P,el - 2 K s T E 2 T s ' where K s is the static gain of the numerator of the open loop transfer function and TE the smaller time constant of the transfer function Ga,el. Consequently, the resulting current transfer function of the closed-loop control is defined as G () = Go,el 1 eel 8 . , 1 + Go,el 2 T s 8 (1 + T s 8) + 1" + ] + }, + { + "image_filename": "designv6_24_0002797_dscc2011-6167-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002797_dscc2011-6167-Figure2-1.png", + "caption": "Figure 2. NEW CATHETER DESIGN", + "texts": [ + " The new design presented in this paper has the ability to handle this problem as well. The new flexible design consists of two concentric tubes: the inner rigid tube and the outer flexible one. There is a cable driven joint attached to the tip of the inner tube. This joint is composed of four disc-shaped sub-elements that are strung together by means of three cables that are 120 degrees apart. This arrangement of cables makes it possible to steer distal flexible portion of the catheter in any desired direction. Figure 2 shows mechanical components of the new design. When the cables are pulled by motor, discs are compressed to each other and form a rather rigid joint due to the friction existing between their surfaces. This joint can eventually be rotated up to 30 degrees in both directions. It is worth noting here that the introduction of joint flexibility into dynamic equations of the system adds an extra level of accuracy to the whole robot analysis. At the time of 2 Copyright \u00a9 2011 by ASME Downloaded From: http://proceedings" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001650_pcicon.2015.7435110-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001650_pcicon.2015.7435110-Figure1-1.png", + "caption": "Fig. 1 A typical motor rotor", + "texts": [ + " The rotor winding is embedded inside of the lamination and is constructed of copper or aluminum bars which are inserted or casted in the rotor core. The bars are then joined to short circuit rings. The short circuit rings can be cast together with the bars or brazed to the bars. Some rotor windings are constructed with stranded coils, similar to the coils that are wound into the stator core. The core is made up of different materials with different yield strengths, thermal expansion coefficients, conductivities and magnetic properties. These are joined together by pressing, casting, brazing or by mechanical fasteners (Fig.1). Rotors can be even more complex with arms (or spiders) that are welded to the shaft. When considering the limits of manufacturing tolerances, the electric motor rotors, even in their simplest construction, are likely to lack perfect mechanical and magnetic symmetry. Even in the longitudinal direction, a perfect shrink fit distribution on the shaft is hard to achieve consistently. Axial lamination pressures may vary along with the conductivity of the portion of the core that contacts the shaft, leading to a somewhat nonhomogeneous heat distribution" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0000685_j.jeurceramsoc.2016.01.002-Figure2-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0000685_j.jeurceramsoc.2016.01.002-Figure2-1.png", + "caption": "Fig. 2. Collocation points of the solid shell element for the alleviation of the transverse shear and trapezoidal locking pathologies.", + "texts": [ + " Specifically, for the treatment of the transverse shear ocking [36], the interpolation of the constant transverse shear esultants p13 and p23 are evaluated at the points A, B, C and D in ig. 2, whose coordinates are: A = (0, \u2212 1, 0), B = (1, 0, 0), C = (0, 1, 0) PRESS eramic Society xxx (2016) xxx\u2013xxx 5 and D = (\u22121, 0, 0). Thus, the ANS interpolation of such components can be expressed as:{ pANS 13 pANS 23 } = { (1 \u2212 2)p13( A) + (1 + 2)p13( C ) (1 + 1)p23( B) + (1 \u2212 1)p23( D) } . (23) As far as trapezoidal locking is concerned [37], the following collocation points are employed for the interpolation of the strain components p33, see Fig. 2: E = (\u22121, \u2212 1, 0), F = (1, \u2212 1, 0), G = (1, 1, 0) and H = (\u22121, 1, 0). The ANS interpolation of this component renders: pANS 33 = 4\u2211 m=1 Nm( 1, 2)p33; Nm( 1, 2) = 1 4 (1 + 1 m 1)(1 + 2 m 2), with 1 m, 2 m = \u00b11 (24) Finally, it is worth mentioning that the use of the EAS method requires further considerations, since the computation of the modified right Cauchy\u2013Green strain tensor that accounts for the metrics modification has to be performed. This is especially suitable for material models which consider the corresponding postulation of the free energy function in terms of the right Cauchy\u2013Green strain tensor or the deformation gradient tensor" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001941_1045389x12451194-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001941_1045389x12451194-Figure1-1.png", + "caption": "Figure 1. Reference coordinate system and displacement designations for the link.", + "texts": [ + " The link itself is assumed as a tubular structure clamped at the base, and with a tip mass featuring an offset between the center of mass and the longitudinal axis of the link. Such an arrangement induces an inertial coupling, between bending and torsion displacements, which will be used in future research to emulate the type of mechanical loads that a serial chain manipulator link undergoes while maneuvering. The link reference coordinate system and the designation of the degrees of freedom that is used throughout this article are presented in Figure 1. The main reference is the plane that is defined by the longitudinal axis of the tube plus the center of mass of the tip plate. An elastic displacement in this plane is designated as in plane (InP). An elastic displacement in the plane perpendicular to InP that also contains the longitudinal axis of the link is called out of plane (OoP). Also, a rotation along the longitudinal axis of the link is called a torsional (Tor) displacement. In designing the link actuation, the starting point is a description of the actuation at the laminate level" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002405_bf03225312-Figure1-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002405_bf03225312-Figure1-1.png", + "caption": "Figure 1: Fully variable valve control Valvetronic with integrated actuator", + "texts": [ + " The combination of centrally positioned direct injection with Valvetronic constitutes a great challenge for the package design in the cylinder head. Efforts must be concentrated on arranging the components in such a way as to comply with all of the requirements relating to kinematics, mechanical strength, cooling and combustion system. A newly developed actuator, with a much more compact design, gives the redesigned Valvetronic system much higher dynamic performance, as well as enabling complete integration of the actuator into the actual cylinder head. The result is shown in Figure 1. Attention must also be drawn to the fact that there is no longer any need for a separate angle sensor as it has been possible to combine this component with the actuator in a single housing [2]. Bank separation is unavoidable in order to achieve a high naturally aspirated and turbocharged full load combined with high specific performance values and direct response from a six-cylinder engine. After considering all of the pertinent aspects, such as low-end torque, responsiveness, maximum output, emissions and costs, a decision was finally made in favour of a TwinPower mono-turbo concept" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002376_j.acme.2017.11.002-Figure12-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002376_j.acme.2017.11.002-Figure12-1.png", + "caption": "Fig. 12 \u2013 Distribution of bainite (Fb) and perlite (Fp) in the coil after cooling processes and in measurement points (P1, P2) during cooling process for steel S355 for early (a) and late (b) cooling mode.", + "texts": [ + " The lack of impact of laminar cooling strategy change on the level of residual stresses can be explained by the similar evolution of phase transformations during cooling process. Ultimately for this steel, the change of laminar cooling strategy has no essential influence on the level of residual stresses. In the case of steel S355, the analysis of numerical simulation showed that phase transformations ends during cooling on ROT or in the coil depending on the applied laminar cooling strategy (Fig. 12). The calculation results of residual stresses are considerably higher (Fig. 14), when the transformation ends during cooling in the coil. For this type of steel, higher carbon and manganese content in combination with different cooling modes varies significantly the evolution of phase transformation. In this work, the further development of a model of coil cooling which is a part of a model of residual stresses in strip has been performed and applied for analysis of residual stresses and temperature" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0002824_ias.2007.4347776-Figure8-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0002824_ias.2007.4347776-Figure8-1.png", + "caption": "Fig. 8. Shape design parameters", + "texts": [ + " 7 shows the cogging torque characteristics of a shape design model by an analytical prediction comparison with the result of initial model. The peak-peak value of cogging torque of initial model is 5.84 (kgcm), but the shape design model is generated to peak-peak value of 3.63 (kgcm). From the results of FE analysis, the reduction of cogging torque by rotor shape design which is eliminated by cogging torque of fundamental and third harmonics is more than 38%. It is found that the position of magnetization dead zone, , and the dead zone angle, , are very effective for the reduction of cogging torque. Fig. 8 shows the shape design parameters which are alteration of dead zone angle and depth of 1. The characteristics of cogging torque characteristics according to changing each of dead zone angles, 1 and 2, is shown in Fig. 9. From the results of the analysis, the cogging torque is slightly influenced by 2 whereas that is severely affected by the dead zone angle, 1, of magnetization. Fig. 10 shows the analysis result according to dead zone depth at the position for fundamental harmonics elimination of cogging torque" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv6_24_0001568_s13246-016-0502-6-Figure4-1.png", + "original_path": "designv6-24/openalex_figure/designv6_24_0001568_s13246-016-0502-6-Figure4-1.png", + "caption": "Fig. 4 View of ankle and foot designed in CATIA software", + "texts": [ + " Also, the dimensions of the male proile were considered as 2 \u00d7 4 \u00d7 20 cm3 which was the same for both bones and facilitates the adjustment of length [9]. To simulate the extension and lexion movements of the leg joint, two bearing were used at the end of these joints (Fig.\u00a03). A compact plastic was embedded in exoskeleton in order to remove the load from the human sole. The end side of the plastic was curved that let the person to continue his movement in stance phase. The sole was attached to the ankle by hinge that can simulate abduction and adduction movements of the ankle (Fig.\u00a04). Human gait cycle begins when the toes leave the ground and ends when heels touch the ground (swing phase). The motion span of the hip joints varies from \u221215\u00b0extension to 30\u00b0 lexion. Also, this span is between 0\u00b0 extension and 60\u00b0 lexion in the case of the knee joint (Fig.\u00a05). In order to make pressure on the joints to move forward or dragging them backwards, jacks need torque. Its value can be obtained through equations (1) and (2) [9]: where TmaxPull is the maximum required torque to drag the joint backwards and T max Push is the maximum required torque to drag the joint forward; Ps is the pressure provided (1)T max Push = P S \u22c5 ( D 2 bore 4 ) R( ) (2)T max Pull = P S \u22c5 ( D 2 bore \u2212 D 2 rod 4 ) R( ) 1 3 by the system; Dbore is the diameter of the cylinder; Drod is the diameter of piston and R(\u03b8) is the desired joint angle" + ], + "surrounding_texts": [] + } +] \ No newline at end of file