diff --git "a/designv10-14.json" "b/designv10-14.json" new file mode 100644--- /dev/null +++ "b/designv10-14.json" @@ -0,0 +1,7971 @@ +[ + { + "image_filename": "designv10_14_0002413_tie.2021.3078388-Figure7-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002413_tie.2021.3078388-Figure7-1.png", + "caption": "Fig. 7. 1/3 electromagnetic FE model.", + "texts": [ + " The FE modeling process is briefly described as follows: 1) Geometry import The structure of axial-flux motors inherently requires the use of a 3-dimensional electromagnetic FE model for accurate analysis. The axial-flux motor studied in this paper has three cycles in space, a 1/3 electromagnetic FE model is built to save the computation time. Meanwhile, the electromagnetic model only needs to contain the effective magnetic circuit, thus active parts including the rotor, stator, PMs, and windings are considered, as shown in Fig.7. The geometry model is input to the FE software firstly. 2) Material definition The materials of the rotor, PMs, windings, and stator are then defined, respectively. The material of the PMS is NdFeB. The relative permeability is 1.2 T and the magnetization pattern is along the axial direction. The material of the windings is set to be copper. The material of the stator and rotor is set to be the silicon steel 35JN270 with a lamination coefficient of 0.97. 3) Winding connection The 30p27s PMSM is a fractional motor with the star connection" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001762_j.triboint.2020.106536-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001762_j.triboint.2020.106536-Figure3-1.png", + "caption": "Fig. 3. Parameters definition associated with the root filling power loss Prfs [17].", + "texts": [ + " In reality, the churning loss contributes a little less than nothing to the total churning losses. Thus, a spur gear with the gear parameters in the heel of the gear is adopted to estimate the churning loss Pcrf . Ignoring gravitation, the churning loss Pcrf of the flow assumed to be a creeping flow is found as (Seetharaman et al. [17]): Prfs \u00bc n\u03bcoil\u03c9\u03b8c\u00f0ri rr\u00de 2 \u00f0ri \u00fe rr\u00de \ufffd D3 rirr D4 \ufffd (17) where n denotes the average number of tooth space below oil and \u03b8c refers to the angle across the tooth space. ri represents the external radius of the creeping flow (rr \ufffd ri \ufffd ro), see in Fig. 3. D3 and D4 are calculated based on the boundary conditions \u03c8 \u00bc 0 atr \u00bc rrj\u03b8j \ufffd \u03b8c orr \u00bc rij\u03b8j \ufffd \u03b8c (18) and d\u03c8 dr \u00bc \ufffd \u03c9rr at r \u00bc rr j\u03b8j \ufffd \u03b8c \u03c9ri \u03c9ri at r \u00bc ri j\u03b8j \ufffd \u03b8c (19) to the biharmonic equation for radial flow: \u03c8 \u00bcD1 \u00fe D2r2 \u00fe D3 log r \u00fe D4r2 log r (20) For a splash lubricated spiral bevel gear partially immersed in oil and partially surrounding by air, the windage power losses similar to the churning losses are composed of three individual parts Pw \u00bcPwp \u00fe Pwf \u00fe Pwts (21) where Pwp denotes the power loss caused by the air acting on pitch cone\u2019s flank of the spiral bevel gear; Pwf refers to the power loss in the toe/heel; Pwts represents the air drag power loss induced within the tooth space/cavity" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001497_icra.2019.8793692-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001497_icra.2019.8793692-Figure1-1.png", + "caption": "Fig. 1. The considered cable-driven instrument passing through the left channel of the STRAS system. Top: top view, bottom: face view. Figure shows left channel frame, the degrees-of-freedom of the instrument, configuration variables and important construction parameters.", + "texts": [ + "5 mm) and a one degree of freedom (DOF) bending tip (length : 18.3 mm, made of 11 mobile vertebrae). Each instrument has 3 motorized DOFs: the bending of the tip in a plane (cable-driven transmission, motor position denoted qbend), the rotation of the instrument in the channel of the endoscope (motor position denoted qrot ) and the translation in the channel (qtrans). The configuration space of the tip of the instrument can be parameterized by the bending angle \u03b2 , the translation t and rotation \u03c6 with respect to the channel (see figure 1). Note that the rotation modifies the orientation of the bending plane. For more information on these instruments, see [14]. 978-1-5386-6027-0/19/$31.00 \u00a92019 IEEE 7913 The objective considered here is to control the cartesian position of the tip of the instruments in the frame of the channel Fch, which describes the task space of the instrument. The reference position for the instrument in this frame will be noted P\u2217 = (x\u2217,y\u2217,z\u2217)T . By assuming that the curvature of the bending tip is constant (uniform) [15], it is possible to convert the configuration variables (bending angle, rotation, translation) into Cartesian positions", + " By using a cylindrical parameterization of the task space, with: \u03b8 = atan2(z,x) and \u03c1 = \u221a x2 + z2, (1) the relations between the configuration variables and the task space variables are partly decoupled. Firstly, one directly obtains \u03c6 = \u03b8 \u2217. Then one can derive the value of the desired bending angle \u03b2 \u2217 by numerically inverting the relation \u03c1 \u2217 = L \u03b2 \u2217 (1\u2212 cos(\u03b2 \u2217))+dsin(\u03b2 \u2217) (2) where L is the length of the bending tip and d the length of the rigid part at the tip of the instrument (see [16]). From \u03b2 \u2217 and the reference depth y\u2217, one can finally get t by noting that (see figure 1): y = t +L+d\u2212\u2206y (3) with \u2206y = L+d\u2212 L \u03b2 sin(\u03b2 )\u2212dcos(\u03b2 ). (4) Assuming a direct control of the configuration space variables from the motors, one can obtain the desired joint positions. For the bending joint, the direct relation is qbend = \u03b2 D 2 , where D is the diameter of the instrument. The other joints being rigid, the transformation is simply qrot = \u03c6 and qtrans = t. This simple kinematic model was tested onto the STRAS instruments by comparing the theoretical 3D positions given by the forward kinematic model with measurements provided by an external camera system (see section III-C)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000769_978-3-662-43893-0-Figure4.40-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000769_978-3-662-43893-0-Figure4.40-1.png", + "caption": "Fig. 4.40 Load distribution and tooth flank stress", + "texts": [ + " The local flank stress resulting from the load determined for the ith flank region Fni (normal force) may be calculated as: \u03c3Hi \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi Ei Fni 2\u03c0\u0394b 1 \u03bd2\u00f0 \u00de\u03c1i s \u00f04:327\u00de Consideration of all discrete points i of a contact line provides the pressure distribution for a tooth in one meshing position. Repeating the procedure through the zone of action provides the total flank stress for a tooth during its engagement (Fig. 4.40). For more advanced calculations, it is also possible to draw in all the components of the spatial stress state and calculate a comparative stress state depth wise. Shear stress \u03c4 caused by friction due to the normal load may also be superimposed to the stress components. 4.4 Stress Analysis 181 The required tooth root stress may be determined from the previously calculated load distribution. Three-dimensional methods of numerical simulation (FEM, BEM, Finite Strips) can be used to calculate the local tooth root stress (stress due to the notch effect) in the face width direction using the known load distribution along the contact line" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002225_s10846-021-01463-6-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002225_s10846-021-01463-6-Figure1-1.png", + "caption": "Fig. 1 Schematic of quadrotor model", + "texts": [ + "eywords Quadrotor \u00b7 Sliding mode control \u00b7 Backstepping control \u00b7 Nonlinear disturbance \u00b7 Nonlinear disturbance observer \u00b7 Flight control Among many kinds of unmanned aerial vehicles (UAV), a multi-rotor UAV, quadrotor, has got a significant interest of researchers for developing various automatic control schemes during past couple of decades. The mathematical model of quadrotor is highly-nonlinear underactuated due to four inputs and six outputs. A general schematic of the quadrotor is illustrated in Fig. 1. In literature, both the linear and nonlinear controllers have been developed and presented for the UAV quadrotor. However, in recent years the trend has been shifted to develop nonlinear controllers to avoid the procedures and assumptions of linearization. A popular nonlinear control scheme, sliding mode control (SMC) being robust to uncertainties and disturbances with known bounds has been used widely by Nigar Ahmed Nigar.Ahmed.pk@gmail.com 1 College of Automation, NUAA, Nanjing, China 2 College of Electrical and Mechanical Engineering, NUST, Islamabad, Pakistan 3 School of Electrical Engineering, SEU, Nanjing, China the researchers" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000769_978-3-662-43893-0-Figure4.39-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000769_978-3-662-43893-0-Figure4.39-1.png", + "caption": "Fig. 4.39 Hertzian stress between two cylinders", + "texts": [ + "38): \u2013 shaft angle \u03a3 shaft angle deviation \u0394\u03a3 \u2013 hypoid offset a hypoid offset deviation \u0394V \u2013 pinion apex distance tz1 pinion axial displacement \u0394H \u2013 wheel apex distance tz2 wheel axial displacement \u0394J 180 4 Load Capacity and Efficiency The non-linear relationship between load and deflection can be taken into account by section-wise linearization or by an iterative procedure. The maximum normal stress acting in the contact area between meshing tooth flanks, also known as Hertzian stress, is employed as an equivalent variable in characterizing tooth flank stress, as in [DIN3990, DIN3991, ISO6336, ISO10300] (Fig. 4.39). In a manner similar to the calculation of local tooth deflection, the original Hertzian model of two contacting cylinders is modified in a model of two contacting \u2018inclined\u2019 cones pointing away from one another, and subdivided into sections (cylindrical segments). The local flank stress resulting from the load determined for the ith flank region Fni (normal force) may be calculated as: \u03c3Hi \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi Ei Fni 2\u03c0\u0394b 1 \u03bd2\u00f0 \u00de\u03c1i s \u00f04:327\u00de Consideration of all discrete points i of a contact line provides the pressure distribution for a tooth in one meshing position" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002095_j.jmapro.2021.01.020-Figure7-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002095_j.jmapro.2021.01.020-Figure7-1.png", + "caption": "Fig. 7. Schematic representation of the flow conditions in the secondary nozzle.", + "texts": [ + " In the case of the modified set-up, the secondary nozzle provides a knife-edge shaped high-speed secondary gas jet, that shears the molten material and provides an additional drag force in the direction perpendicular to the melt ejection, as shown in the Fig. 6(b). The secondary gas jet removes the molten material from the bottom of the cut and thus, causes a significant reduction in the dross, as shown in the Fig. 6(d). A.K. Singh et al. Journal of Manufacturing Processes 64 (2021) 95\u2013112 Estimation of the thrust force generated by the secondary nozzle Fig. 7 shows the schematic representation of the flow conditions in the secondary nozzle. The inlet of the secondary nozzle was connected to the highly pressurized argon cylinder, and the outlet was directed towards the dross. The calculation of the thrust force in the present study was based upon the application of Bernoulli\u2019s equation (Eq. 2 [36]) to the steady-state flow of the compressible gas through the nozzle. V1 2 2 \u00d7 g + z1 + Rc \u00d7 T g \u00d7 ln ( p1 p2 ) = V2 2 2 \u00d7 g + z2 (2) Here, z1 = Elevation head at the inlet (m), z2 = Elevation head at the outlet (m), V1 = Flow velocity (m/s) of the gas at the inlet, p1 = Flow pressure (N/m2) of the gas at the inlet, V2 = Flow velocity (m/s) of the gas at the outlet, p2 = Flow pressure (N/m2) of the gas at the outlet, Rc = Characteristic gas constant (J/kg" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001874_s00170-019-04738-3-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001874_s00170-019-04738-3-Figure3-1.png", + "caption": "Fig. 3 The arc-edge cutter of gear", + "texts": [ + " M 1p \u00bc a11 a12 a13 a14 a21 a22 a23 a24 a31 a32 a33 a34 a41 a42 a43 a44 2 664 3 775 where a11 \u00bc cos\u03b31cos\u03c6p; a12 \u00bc \u2212cos\u03b31sin\u03c6p; a13 \u00bc sin\u03b31; a14 \u00bc Sr1cos\u03b31cos \u03c6p \u00fe q1 \u2212Xb1sin\u03b31\u2212XG1; a21 \u00bc cos\u03c61sin\u03c6p\u2212cos\u03c6psin\u03b31sin\u03c61; a22 \u00bc cos\u03c61cos\u03c6p \u00fe sin\u03c6psin\u03b31sin\u03c61; a23 \u00bc cos\u03b31sin\u03c61; a24 \u00bc Em1cos\u03c61\u2212X b1cos\u03b31sin\u03c61 \u00fe Sr1sinq1 cos\u03c61 sin\u03c6p \u00fe cos\u03c6p \u00fe sin\u03b31sin\u03c61 sin\u03c6p\u2212cos\u03c6p h ; a31 \u00bc \u2212sin\u03c61sin\u03c6p\u2212cos\u03c6psin\u03b31cos\u03c61; a32 \u00bc \u2212cos\u03c61sin\u03b31sin\u03c6p\u2212cos\u03c6psin\u03c61; a33 \u00bc cos\u03b31cos\u03c61; a34 \u00bc \u2212Em1sin\u03c61\u2212X b1cos\u03b31cos\u03c61\u2212Sr1 sin\u03c61sin q1\u00fe\u03c6p \u00fe sin\u03b31cos\u03c61cos q1\u00fe\u03c6p h i ; a41 \u00bc a42 \u00bc a43 \u00bc 0; a44 \u00bc 1: The arc-edge cutter of gear is shown in Fig. 3. The main profile for machining the gear working surface is as follows: rh \u00bc sg 0 h sg 1 T \u00f06\u00de where sg and h(sg) are the parameters of the main profile. Base on the coordinate transformation, the position vector and unit normal vector of cutting cone are as follows: rg \u00bc Rg\u2212sgsin\u03b12 \u00fe h sg cos\u03b12 cos\u03b8g Rg\u2212sgsin\u03b12 \u00fe h sg cos\u03b12 sin\u03b8g \u2212sgcos\u03b11\u2212h sg sin\u03b12 1 2 664 3 775 \u00f07\u00de ng \u00bc \u2212 cos\u03b12 \u00fe h0 sg sin\u03b12 cos\u03b8g \u2212 cos\u03b12 \u00fe h0 sg sin\u03b12 sin\u03b8g \u2212 sin\u03b12\u2212h0 sg cos\u03b12 2 4 3 5 \u00f08\u00de where \u03b8g is the parameters of the main profile" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000493_2015039-Figure8-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000493_2015039-Figure8-1.png", + "caption": "Fig. 8. Simplified modelling of a track.", + "texts": [ + " The latter has the lowest vaporisation temperature (1090 \u25e6C) amongst all other compounds present; it is thus possible that part of the magnesium vaporised during the laser fusion process. This phenomenon was also observed at the LERMPS laboratory [9]. Powders are thus supplied with an increased quantity of magnesium to ensure the conformity (according to NF EN 1706 [8]) of the alloy after selective laser melting. According to Tissot [10], with a simple modelling of a track with a hemispherical melt (Fig. 8) which is quite close to reality (see Fig. 22), it is possible to determine the theoretical maximum temperature difference in the track from Equation (4): \u0394T = q 2\u03bb\u03c0 r (4) with: q: power passing through the surface of the melt bath. q = 40 W, allowing the power transmitted to the track to be 20% (estimation) of the laser power (200 W); this is due to the high reflectivity of the aluminium alloy [11]. \u03bb: thermal conductivity of the AlSi10Mg alloy made by SLM. \u03bb = 103 W.m\u22121.\u25e6C\u22121 [12]. r: mean radius of the track" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000897_j.cad.2015.12.001-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000897_j.cad.2015.12.001-Figure1-1.png", + "caption": "Fig. 1. The face-hobbing cutting system.", + "texts": [ + " In the present paper, Gleason TRI-AC R\u20dd cutting system, nongenerating kinematic chains and workpiece geometry are described in Section 2. In Section 3, the projection of un-deformed chip geometry on the rake face of the cutting blades, Chp, is derived. The cutting forces are predicted in Section 4. Finally, in Section 5, two case studies are presented which for both, the inprocess model of the workpiece and the un-deformed chip geometry are derived and the cutting forces are predicted. In face-hobbing, a cutter head and groups of cutting blades make a cutting system (Fig. 1). In the concept of half profile blades, one inside and one outside blade create a blade group. Each blade is mounted inside a cutter head slot with a special orientation. In the present paper, themathematical representation of the cutting edge and cutting surfaces of outside blades are mentioned. However, in case of inside blades, all the formulations can be used easily since the inside blade is a symmetry of the outside blade. The cutting edge, rake and relief surfaces are the most important features of the cutting blades" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001712_j.ymssp.2020.106823-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001712_j.ymssp.2020.106823-Figure2-1.png", + "caption": "Fig. 2. Gear backlash nonlinearity for normal forces Fjwg :", + "texts": [ + " The lubricant mesh stiffness can be obtained by employing a trace method to approximate the hysteresis loop of the lubricant reaction [11], and the rectangular wave function could be used to approximate the solid mesh stiffness [12]. Generally, the mesh damping is proportional to the gear teeth meshing stiffness, as follow C j l;s \u00bc 2nl;s ffiffiffiffiffiffiffiffiffiffiffiffi K j l;sIeq r \u00f08\u00de where Ieq \u00bc IpIg= IpR 2 bg \u00fe IgR 2 bp is the gear pair\u2019s equivalent mass. fl;s is the critical viscous damping ratio of lubricant and solid, respectively. The normal forces incorporated both solid and lubricant contact is depicted in Fig. 2, in which the dashed line is the trace of the hysteresis loop. The hysteresis loop, demonstrated by a closed cycle, consists of two solid lines: one is the line for _x t\u00f0 \u00de > 0 and another is the line for _x t\u00f0 \u00de < 0. Sliding on the gear tooth surface causes a frictional force along the off-line of action direction. Because the hydrodynamic flank friction has little influence on the gear pairs [28], only the friction force for the solid contact are considered. Thus, the nonlinear friction force Fjfg acting on the wheel for the jth tooth pair is Fjfg t\u00f0 \u00de \u00bc 0 x t\u00f0 \u00dej j < L lFjwg t\u00f0 \u00design ujs t\u00f0 \u00de x t\u00f0 \u00dej j > L ( \u00f09\u00de here, lj s denote the solid friction coefficient for the jth tooth pair" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001703_j.jmapro.2020.02.035-Figure8-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001703_j.jmapro.2020.02.035-Figure8-1.png", + "caption": "Fig. 8. The residual ease-off of pinion concave side after machine tool setting modification.", + "texts": [ + " The prescribed easeoff of tooth flank before machine tool setting modification is shown in Fig. 7. It can be seen from the Fig. 7 that the predesigned ease-off at each discretized point has good smoothness and the discretized point distributes along the tooth face by 9 points and along the tooth height by 5 points [19]. Where, the root mean square error of the ease-off on the pinion concave side is 21.57um, the minimum ease-off is -25.46um at the middle area of tooth flank and the maximum ease-off is 41.76um at the toe and root area. Fig. 8 depicts the residual ease-off of pinion concave side after machine tool setting modification. The root mean square error of the residual ease-off is 1.35um while the maximum is 3.16um at the toe of tooth flank and the minimum value is -1.53um at the middle area of tooth flank. The large error is mainly concentrated in the boundary of the tooth flank. To implement the optimal machine setting modification, we need to calculate the sensitivity coefficients and the sensitivity matrix J of the tooth flank error to the machine settings" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000665_tmag.2017.2764260-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000665_tmag.2017.2764260-Figure1-1.png", + "caption": "Fig. 1. Structure and operation principle of the proposed MS-PMECC. (a) Structure. (b) lsd = 0 mm. (c) lsd = 8 mm. (d) lsd = 20 mm.", + "texts": [ + " Finally, the analytical results of the electromagnetic characteristics, including air-gap magnetic field and torque-slip speed curve, are verified by 3-D FEM. Manuscript received June 27, 2017; revised September 5, 2017; accepted October 12, 2017. Date of publication November 9, 2017; date of current version February 21, 2018. Corresponding author: H. Lin (e-mail: hyling@seu.edu.cn). 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.2017.2764260 The structure of the proposed MS-PMECC is shown in Fig. 1. The stationary MSR is formed by ferromagnetic material, which can be shifted along the axial direction steel track by a stepping motor-driven mechanical manipulator [6], [7]. The rotor consists of a coaxial PMR and a CR. The PMR is characterized by two axially parallel consequentpole rotors and an axial magnetized PM ring located between the back-iron cores. The copper sheet (CS) is fixed on the iron core of the CR. For the proposed MS-PMECC, the PM ring serve as an airgap flux adjuster. The fluxes produced by the PM ring passing through the iron poles can be controlled by adjusting the axial displacement of the MSR. As shown in Fig. 1(b)\u2013(d), when the shifting distance (lsd) increases from 0 to 20 mm, the PM ring fluxes short circuited by the MSR increase gradually. In other words, the effective air-gap fluxes decrease, which results in the flux-weakening effect. Consequently, the slip speed of the coupling under a given load-torque operation can be adjusted with the aid from the MSR shifting control. Considering the special structure of the proposed MS-PMECC, the 3-D FEM is usually employed to compute 0018-9464 \u00a9 2017 IEEE. Personal use is permitted, but republication/redistribution requires IEEE permission" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001621_s40192-019-00148-1-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001621_s40192-019-00148-1-Figure1-1.png", + "caption": "Fig. 1 Schematic of the bridge structure and the cuts necessary for X-ray samples with leg indices highlighted. The red lines denote EDM cuts", + "texts": [ + " During the build, a Nd:YAG laser was operated at 195\u00a0W with an approximately Gaussian-shaped power distribution function at the sample position and an estimated spot size (full width at half maximum) of 50\u00a0\u00b5m. The scanning speed was 800\u00a0mm/s. The hatch distance was 100\u00a0\u00b5m. The build was conducted in N2 with an approximate oxygen level of 0.5%. The height of each layer is \u2248 20\u00a0\u00b5m. The AM builds of IN625 and 15-5 follow the same bridge structure geometry that has 12 legs with different sizes. This bridge structure is illustrated in Fig.\u00a01. More details about the build can be found elsewhere. To prepare the X-ray samples, three different sets of wire electron discharge machining (EDM) cuts were made. The first cut, highlighted as \u201cCut location 1\u201d in Fig.\u00a01, removes the top of the bridge. The second set of cuts consisted of 8 vertical cuts from the centers of the 5-mm-thick legs, which were made to produce 4 specimens of 0.5\u00a0mm thickness. The last cut was a single EDM cut that removed the legs from the base plate. All these cuts were made in the as-built condition, i.e., without a stress relief heat treatment. The X-ray specimens were mechanically thinned and polished to remove surface abnormalities, following a standard metallographic procedure. High-resolution synchrotron XRD experiments were conducted at the dedicated powder XRD beamline 11-BM-B at the Advanced Photon Source (APS), Argonne National Laboratory [15]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001778_tsmc.2020.3009405-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001778_tsmc.2020.3009405-Figure1-1.png", + "caption": "Fig. 1. One-link manipulator.", + "texts": [ + " Remark 4: Compared with the control scheme in [27], the one we presented in this article improves the accuracy of the setting time. The proposed scheme in [27] is just a special form of the proposed scheme in this article. In addition, our method simplifies the calculation procedure in dealing with the form of nonstrict feedback. Moreover, the system (6) in this article is more complex than the one of [27]. IV. SIMULATION EXAMPLES In order to prove the feasibility of the method proposed in this article, a one-link manipulator with kinematic dynamics and disturbance shown in Fig. 1 is studied [29]. The dynamic equation of the one-link manipulator can be described as \u23a7\u23aa\u23aa\u23a8 \u23aa\u23aa\u23a9 x\u03071 = x2 + f1(x) x\u03072 = x3 + f2(x) x\u03073 = 10u + f3(x) y = x1 (30) where f1(x) = 0, f2(x) = \u2212x2 \u2212 10 sin x1 + \u03d1d , and f3(x) = \u22122x2 \u2212 10x3 with x1(0) = 0, x2(0) = 0, x3(0) = 0, \u03b8\u03021(0) = 0.5, \u03b8\u03022(0) = 0.5, and \u03b8\u03023(0) = 0.5. Select the trajectory r = (\u03c0/2)([1/2] sin([3/2]t)+sin([1/2]t)). The adaptive quasi-fast finite-time neural tracking control scheme can be presented for (30) by Theorem 1. In the simulation, the parameters are chosen as \u03c11 = 10, \u03c12 = 10, \u03c13 = 100, \u03c41 = 21, \u03c42 = 21, \u03c43 = 20, \u03bc = 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002083_bf02123920-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002083_bf02123920-Figure1-1.png", + "caption": "Fig, 1. Hydraulic circuit of a pump and a tube with resistance.", + "texts": [ + "' volume I electric current v velocity L electric inductance II: power 31 hydraulic inertance X hydraulic gyro-reactance m mass; mechanical inertance X e electric gyro-reactance P hydraulic potential .r coordinate p pressure a',,, nlechanicat gyro-reactance \u00a7 2. Hydraulic systems. By hydraulic systems we shall for the moment unders tand systems in which an incompressible fluid (e.g. water) circulates between rigid walls. be the dischaJ~e (currant, volume velocity) through the tube, i.e. the volume of liquid t ransmi t ted per uni t time. Let H denote the total head at some point, that is the sum of the geometric height z of the point above some zero level (see fig. 1), the pressure head p/~g (~o = density) and the velocity head v2/2g. Head is the q u a n t i t y commonly used in hydraul ics , and therefore i t is most prac t ica l in appl icat ions to define hydraul ic impedances as, re la t ions between head and discharge. Present ly , however, we in tend to give a more general physical discussion of hydraul ic systems, and then it is preferable to use the q u a n t i t y which we shall call the hydraulic potential. Fric t ion in the conduit of fig. I causes a loss of potent ia l where R = 8o31/~a 4, and R = 3[n,l/4ab a for a round tube and a crevice respect ively (v = k inemat ic viscosity)", + " We suppose that the upper liquid is covered by a rigid boundary (practically the same effect is obtained by making the difference in density small compared with the densities themselves, because the surface disturbances at tending the internal waves are then very small). The system may be schematized by sections of parallel branches formed by bi-terminal storage elements of the type of fig. 2b, and two lines of inertances in series, one line formed by the inertances of the upper liquid (Ms) an the other line by those of the lower liquid (21I,) (fig. 1 lb). The electric analogue is a set of two conductors, each with series inductance, and parallel capacity between them (fig. 1 l c). \u00a7 5. MechaJzical systems. By mechanical systems we shall for the moment understand systems of solid bodies moving relative to one another. - - i i - - i I , , i - - ~' . I., I I,. I,, ; I. ,, .. ,. ~, ]Zig. 12. a. Mechanical resistance; b. Mechanical compliance: c-e. Mechanical inertances. other (fig. 12a). Let v I be the velocity.of the first and v 2 of the second body. Then the force F exerted by each body upon the other depends on the relative velocity v ~ - v 2. In case of lubricated surfaces we may put F = r(v I - - v 2 ) , where r = e~,S/6, (e = density, 7, = kifiematic viscosity of the lubricat ing liquid, S = area of the lubricated surfaces, and 6 = distance between them) may be called the mechanical resistance of the system" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001644_rnc.4758-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001644_rnc.4758-Figure3-1.png", + "caption": "FIGURE 3 Flexible wing unmanned aerial vehicle (UAV) [Colour figure can be viewed at wileyonlinelibrary.com]", + "texts": [ + " Then, the deflection of the trailing edge of the parafoil canopy, so-called flap deflection, is the horizontal control quantity of this system. The flap deflection is realized by changing the length of the control ropes. The asymmetric flap deflection and symmetrical flap deflection are shown in Figure 2. The asymmetric flap deflection is applied to control the flight direction and the symmetrical flap deflection is applied to change its airspeed in a small range.1 From Figure 2A, it can be observed that the flap deflection will change the shape of the parafoil canopy irregularly. Furthermore, with the power plant, as shown in Figure 3, the system will have faster flight velocity. In this condition, it will enlarge the oscillation of the control quantity of both the horizontal and vertical controllers. Furthermore, in this system, it also exists coupling influence of these two controllers. By changing the aerodynamics of the parafoil canopy FIGURE 1 State of the parafoil canopy. A, Parafoil canopy in the quiescent state; B, Parafoil canopy in flight state [Colour figure can be viewed at wileyonlinelibrary.com] (A) (B) FIGURE 2 Flap deflection" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000756_j.ifacol.2018.11.549-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000756_j.ifacol.2018.11.549-Figure1-1.png", + "caption": "Fig. 1. Quadrotor model with coordinate axes", + "texts": [ + " The remainder of this paper is organized as follows: Next section includes preliminaries for deriving an attitude model of the quadrotor UAV. Then, detailed theoretical analyses of the proposed SMC technique are presented and compared against the conventional SMC. Next, the application of the proposed methods for UAVs using numerical simulations and analyses are given. Finally, the paper is ended with some concluding remarks. 2. QUADROTOR ATTITUDE MODEL A quadrotor UAV includes four rotors with bolted propellers in cross-shaped frame as illustrated in Fig. 1. The attitude change of the quadrotor results from variations on forces and moments produced by adjusting rotors\u2019 speeds. The quadrotor attitude dynamics can be written in the following form [H. Nemati and Jamei (September 2017)]: \u03c6\u0308 \u03b8\u0308 \u03c8\u0308 = Jyy \u2212 Jzz Jxx \u03b8\u0307\u03c8\u0307 Jzz \u2212 Jxx Jyy \u03c6\u0307\u03c8\u0307 Jxx \u2212 Jyy Jzz \u03c6\u0307\u03b8\u0307 \u2212 Jr\u2126r \u03b8\u0307 Jxx \u2212 \u03c6\u0307 Jyy 0 + u2 Jxxu3 Jyy u4 Jzz (1) where [ \u03c6 \u03b8 \u03c8 ] T are Euler angles and known as roll (rotation around x\u2212axis), pitch (rotation around y\u2212axis) and yaw (rotation around z\u2212axis)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000104_s12598-015-0461-1-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000104_s12598-015-0461-1-Figure3-1.png", + "caption": "Fig. 3 Tensile specimen (mm)", + "texts": [ + "4 Mechanical testing The static tensile tests were performed on a Zwick Universal Testing Machine with a displacement rate of 1 mm min-1 at room temperature. The tensile specimens were cut out from the fabricated components in laser scanning x and the wall height z orientations, i.e., with mechanical loading direction parallel to the direction of the laser scanning x and parallel to the direction of the wall height z (designated with x and z, respectively). The dimensions of the specimens were 0.4 mm in thickness and 1.0 mm in width, with a gauge length of 4 mm. The detailed size of the tensile specimens is shown in Fig. 3. 3 Results and discussion 3.1 Microstructure analysis A thin-wall component consisting of 50 layers with height of about 7 mm and wall thickness of about 0.6 mm was selected to analyze the microstructure. Figure 4 shows the microstructure within the longitudinal cross section (x\u2013z plane) of the component. Long columnar grains of irregular sizes and shapes form in the sample, and the layer bands resulting from the remelting of previously deposited layers can also be seen within the cross section (Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002061_tec.2020.3048442-Figure6-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002061_tec.2020.3048442-Figure6-1.png", + "caption": "Fig. 6. Coil EMF phasor and winding connection illustration with Nr=7 and Nr=8 as examples (at the rotor position of maximum phase A flux linkage). (a) Coil EMF phasor of Nr=7. (b) Armature winding connection of Nr=7. (c) Coil EMF phasor of Nr=8. (d) Armature winding connection of Nr=8.", + "texts": [ + " 1 (b), which potentially results in unbalanced magnetic force (UMF). The asymmetric structure and UMF can be eliminated by doubling the stator and rotor poles. The mechanical and electrical angles between two armature coils can be defined in (3) according to the previous analysis, in which the stator tooth number of n=4k (k=1, 2, 3) is wound with field coil. To illustrate the armature winding configuration, the coil EMF phasors and winding connections of 12/7 and 12/8-stator slots/rotor poles HE machines are shown in Fig. 6. Similarly, the armature winding connections of other rotor pole number with Nr=4,5,7,8,10,11\u2026 can be obtained. It is worth emphasizing that the EMF phasors in Fig. 6 indicate the phasor angle of corresponding tooth wound armature coil, whereas the amplitudes of different coils can be different as discussed in Fig. 3. Obviously, the armature winding distribution factor is 1 when the rotor pole number Nr=4k, which can result in higher phase flux linkage as well as back-EMF. 1 1 360 1 , 4 360 1 , 4 n m s n e r s n n k N N n n k N (3) In order to evaluate the torque density of the proposed DSHE machine, global optimization is conducted in this section", + " Furthermore, global optimization is conducted on the proposed DSHE machine with different stator/rotor pole number combinations for maximum average torque under flux enhancing. After global optimization, the torque densities of the proposed DSHE machines are compared in Fig. 9. The field winding regulation ratio is defined by (8), where T+, T0, and Trepresent average torque with flux-enhancing field excitation, no field excitation, and flux-weakening field excitation. The proposed DSHE machines with 8- and 16-rotor poles exhibit higher torque density. This can be explained by the EMF phasor in Fig. 6, and the winding distribution factor is 1 when the rotor pole number satisfies Nr=4k. As reported in [29] and [30], the flux switching HE machine possesses 25% field regulation ratio with maximum field excitation. Therefore, the flux regulation ratio is satisfied for the 12/8 and 12/16-stator slots/rotor poles DSHE machines, and are chosen for further investigation. 0 DC T T T (8) V. INFLUENCE OF IRON BRIDGE ON HYBRIDIZATION Since the magnetic field of HE machine is excited by both PM and field winding, it is worth investigating the hybridization of two magnetic sources" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000098_cca.2014.6981467-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000098_cca.2014.6981467-Figure1-1.png", + "caption": "Fig. 1. Wingcopter, coordinates systems and some notations", + "texts": [ + " In addition, alike classical quadrotor UAVs, it is also possible to monitor the control torque vector by controlling the 4 propellers\u2019 speed, in addition to aerodynamic torques from control surfaces. The vehicle is equipped with 4 control surfaces, namely an aileron on each wing with a deflection \u03b4a1 and \u03b4a2, respectively, an elevator with a deflection \u03b4e, and two rudders with deflection angles \u03b4r1 and \u03b4r2, respectively. The airspeed vector Va makes an angle \u03b2 with the plane (xb, zb), which is also known as the sideslip angle. The angle of attack is \u03b1 as shown in Fig. 1. 978-1-4799-7409-2/14/$31.00 \u00a92014 IEEE 1011 Let I = {0;xI , yI , zI} denotes an inertial frame with zI pointing downward, which is consistent with the common use of NED (North-East-Down) frames in aeronautics. Let B = {G;xb, yb, zb} denotes a body frame, with G the vehicle\u2019s center of mass. The orientation of the aircraft bodyfixed frame B with respect to the inertial frame I can be represented by a rotation matrix RBI , which can also be related to a quaternion representation qBI = [q0, qv] > using the Rodrigues\u2019 rotation formula: RBI = I3 +2q0qv\u00d7+2(qv\u00d7)2, with I3 the 3 \u00d7 3 identity matrix, and (\u00b7)\u00d7 denoting the skew-symmetric matrix associated with the cross product, i", + " The propeller rate dynamics are modeled as a first order system with time constant \u03c4n, which corresponds to the dynamics of a brushless motor as used on the Wingcopter : $\u0307i = \u2212 1 \u03c4n $i + 1 \u03c4n $c . (14) IV. AERODYNAMIC FORCES: Fa B The air flow acting on the airframe is responsible for the aerodynamic forces. The air flow is described by the airspeed vector Va. Its norm is Va and its direction relative to the airframe is defined by two angles, namely the angle of attack \u03b1 and the sideslip angle \u03b2. As shown in Fig. 1, the angle of attack \u03b1 is the angle between the projection of the airspeed vector Va onto the (xb, zb) plane and the xb axis. The sideslip angle \u03b2 is the angle between the projection of the airspeed vector Va onto the (xb, zb) plane and the airspeed vector itself. The wind axes coordinate system is such that the xw axis points along the airspeed vector Va. The rotation matrix RWB is necessary to transform vectors and point coordinates from the aircraft body-fixed frame (B) to the wind frame (W) and vice versa according to the following formulae: AW = RWB AB or AB = (RWB )TAW = RBWAW with RWB = cos\u03b1 cos\u03b2 sin\u03b2 sin\u03b1 cos\u03b2 \u2212 sin\u03b2 cos\u03b1 cos\u03b2 \u2212 sin\u03b1 sin\u03b2 \u2212 sin\u03b1 0 cos\u03b1 " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000520_s40684-016-0008-4-Figure6-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000520_s40684-016-0008-4-Figure6-1.png", + "caption": "Fig. 6 Misaligned roller bearing", + "texts": [ + "15 The elastic deformation of the rolling elements in a radial roller bearing can be described by a thin lamina model. To calculate the elastic deformation of a roller, a roller is divided into at least 30 thin, lamina of equal size, as shown in Fig. 5. The load-deformation equation of thin lamina k of roller j is given by Eq. (9). (9) The elastic deformation of roller j with respect to the radial deformation of the raceway \u03b4r is given by Eq. (10), and the misalignment angle between the inner and outer raceways is given by Eq. (11), as shown in Fig. 6. The elastic deformation of the kth thin lamina of the j th roller can be obtained from Eq. (12) using Eqs. (10) and (11). If the right-hand side of Eq. (12) is negative (-), the deformation is zero. (10) (11) (12) SF \u03c3FG \u03c3F --------= \u03c3F \u03c3F0KAK\u03b3KVKF\u03b2KF\u03b1= \u03c3F0 Ft bmm --------- YFYSY\u03b2YBYDT= \u03c3FG \u03c3FlimYSTYNTY\u03b4relTYRrelTYX= SH \u03c3HG \u03c3H ---------= \u03c3H ZB D, \u03c3H0 KAK\u03b3KVKH\u03b2KH\u03b1= \u03c3H0 ZHZEZ\u03b5Z\u03b2 Ft d 1 b ------- u 1+ u ----------= \u03c3HG \u03c3HlimZNTZLZRZVZWZX= qj k, cs\u03b4j k, 10 9 ----- = \u03b4j \u03b4rcos\u03d5j s 2 --\u2013= \u03c8j arctan tan\u03c8 cos\u03d5j( )= \u03b4j k, \u03b4j xktan\u03c8j\u2013( )= The following force-moment equilibrium relation exists between the external force acting on the two raceways and the rolling element load with respect to the moment" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002011_j.ijthermalsci.2020.106610-Figure4-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002011_j.ijthermalsci.2020.106610-Figure4-1.png", + "caption": "Fig. 4. Spring\u2013mass\u2013damper model. Fig. 5. Decomposition on the excitation force.", + "texts": [ + " The following sections, for the purpose of assessing the energy loss of angular contact ball bearings more accurately, aim at exploring the causes of bearing temperature rise in vibration. During vibration, energy is periodically transformed back and forth between kinetic and potential until all the energy is lost through damping. As a result, the energy loss from kinetic and potential energy leads to a marked rise in bearing temperature. The lumped parameter model is typically depicted as shown in Fig. 4, where k is the linear spring constant, and r is the viscous damping coefficient. To facilitate analyzing the input energy in vibration, let\u2019s set a harmonic function F(t) to simulate an excitation force for now. F(t)=F0 sin(\u03c9t) (1) where F0 is the amplitude of excitation force F(t), and \u03c9 represents the excitation frequency. According to paper [21], its steady solution is expressed as x(t)=B sin(\u03c9t \u2212 \u03d5) (2) When the exciting force F(t) is displaced a micro displacement dx, the energy consumption is dW =F(t)dx = F(t)x\u0307dt (3) For a complete cycle t = ( 0,2\u03c0 \u03c9 ) , so the energy input into the system can be calculated by [23]", + " 5, we have { F1 = F0 cos \u03d5 sin(\u03c9t \u2212 \u03d5) F2 = F0 sin \u03d5 sin(\u03c9t \u2212 \u03d5) (5) here, F1 denotes the component in phase with the displacement of F0, and F2 is the component in phase with the velocity of F0. So, the energy input by the exciting force F(t) in one cycle is the sum of the work done by the two components F1 and F2[21]. here, hence, Eq. (4) can be rewritten as WF =WF1 + WF2 = \u03c0F0B sin \u03d5 (8) Clearly, the work done by a harmonic exciting force in a period relies on the amplitude, the exciting force and the angle between the exciting force and the amplitude. For the spring\u2013mass\u2013damper system as seen in Fig. 4, the damping force can always be expressed as [21]. Substituting Eq. (2) into the above formaula gives In one cycle of vibartion, the energy expenditure induced by the damping can be formulated as Wr = \u222b T 0 Fdx\u0307dt = \u222b 2\u03c0 \u03c9 0 rB2\u03c92cos 2(\u03c9t \u2212 \u03d5)d(t) = \u03c0r\u03c9B2 (11) At this point, it is clear the power loss of a forced vibration system only depends on the amplitude because the damping relates to the physical system as well as the frequency lies on the excitation source. So we need to know only the amplitude of vibration response to calculate the vibration-induced heat" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000893_j.ymssp.2015.10.028-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000893_j.ymssp.2015.10.028-Figure1-1.png", + "caption": "Fig. 1. Piezoelectric strain sensor and its dimensions [5].", + "texts": [ + " This method is based upon measuring the alteration of tangential surface strain on the fixed ring gear due to cycling of the bearing damage. The measured strain alteration is caused by periodically recurring alterations of the load distribution of the planets. In the work on hand piezoelectric strain sensors are applied on the ring gear of a WEC-gearbox that is deployed on a WEC-system test rig and the measured signals are analyzed regarding their mechanical basics, measurement chains and inferable information). For the strain measurements a piezoelectric strain sensor produced by PCB Piezotronics, Depew, New York, USA (Fig. 1 and Table 1), and for reference strain gauges (SG) of the company Hottinger Baldwin Messtechnik GmbH, Darmstadt, Germany are used. Strain sensors and strain gauges are applied using an adhesive based on Cyanoacrylate, because additional screw joints on the ring gear should be avoided. For some measuring points, the coated surface is grinded before applying the sensor. Additionally the signals of an angle of rotation sensor and a strain-gauge based torque sensor at the slowly spinning gearbox input shaft (low-speed shaft, LSS) are recorded" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001352_j.mechmachtheory.2016.02.015-Figure7-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001352_j.mechmachtheory.2016.02.015-Figure7-1.png", + "caption": "Fig. 7. Coordinate system applied for generating pinion.", + "texts": [ + " The position vector r2 and the unit normal n2 of the coast side of the pinion tooth surface at the reference point M2 \u2032 , can be represented as r2 \u00bc r1c R p; \u03b82\u00bd \u00f024\u00de n2 \u00bc n1c R p; \u03b82\u00bd \u00f025\u00de Here, R[p, \u03d12] is a transformation matrix that denotes the rotation angle \u03d12 about the vector p. According to step 4, the curvature parameters of the pinion tooth surface based on the tangency of the generating and pinion tooth surfaces at the reference points are determined in this section. The configuration in Fig. 7 is the same as that in Fig. 5. Coordinate systems Sm{Xm, Ym, Zm} are rigidly connected to the cutting machine. The top (A\u2013A), bottom and middle of Fig. 7 are the machine's front view, top view and side view (projection of the head cutter), respectively. The cradle rotates about the G-axis. The p-axis is the unit vector of the pinion spindle. O1 is the cross point of the pinion. The manufacturing coordinate systems of the gears and pinions and the installation coordinate system of the hypoid gear sets are represented in the coordinate system Sm. Some of the basic machine-tool settings for generating pinion are: the machine root angle \u03b3m1, the machine center to back Xp, the cradle angle q1, the radial distance Sr1, the blank offset Em1, the sliding base Xb1, the swivel angle J1, the tilt angle I1, the horizontal and vertical setting (Hp and Vp) of pinion head-cutter" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000769_978-3-662-43893-0-Figure4.16-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000769_978-3-662-43893-0-Figure4.16-1.png", + "caption": "Fig. 4.16 Geometry data to calculate tooth root stress (left) and gradient of local tooth root stress (right)", + "texts": [ + "2 Load Capacity Calculation 117 118 4 Load Capacity and Efficiency 4.2 Load Capacity Calculation 119 In order to calculate the tooth root load capacity of a bevel gear, the maximum bending stress which occurs in the tooth root is determined. This stress value is compared to a permissible stress derived from standardized cylindrical test gears (standard strength values in [ISO6336]). Bending stress \u03c3F0 is calculated at the contact point of the 30 tangent in the fillet of the loaded flank since tooth root breakage usually starts at this point (Fig. 4.16). The tooth root stress \u03c3F is calculated from the local tooth root stress \u03c3F0 multiplied by load factors and load distribution factors which consider overloading and real load distribution in actual operation of the gears. The local tooth root stress \u03c3F0 results from a nominal stress and a number of corrective factors Y for specific influences on tooth root stress, e.g., the notch effect or the complex stress condition in the tooth root regarding compressive stress from the radial load Fn\u2219 sin \u03b1 and shear stress from the tangential load Fn\u2219 cos \u03b1 (Table 4" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002195_s12206-021-0606-0-Figure4-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002195_s12206-021-0606-0-Figure4-1.png", + "caption": "Fig. 4. Inner-race shoulder/cage interaction.", + "texts": [ + " (5) In the abovementioned equations, the exponent value, 10/9 is for line contact for model simplification [16]. In addition, the normal contact forces ijN and ojN are effective only when the amounts of contact deformation ij\u03b4 and oj\u03b4 are positive. The cage types of roller bearing can be divided into innerrace landing and outer-race landing depending on the shoulder position where cage motion is guided. For the inner-race landing type, where the cage is close to the inner-race shoulder, geometric interactions between cage and inner-race shoulder can be expressed in Fig. 4. The center of inner-race shoulder is identical to that of the inner-race, and translational motion of the center of cage is expressed in cx and cy . The amount of contact deformation ( ci\u03b4 ) between the inner-race shoulder and cage can be calculated as follows. ( ) ( )2 2 ,= \u2212 + \u2212 \u2212 +ci i c i c ci sh ix x y y r r\u03b4 . (6) The azimuth angle of the position of contact point between the inner-race shoulder and cage is calculated as follows. ( ) ( ) 1 2 2 cos\u2212 \u2212= \u2212 + \u2212 i c ci i c i c x x x x y y \u03b8 . (7) \u2032\u2032\u2032 In the case of the outer-race landing type, geometric interaction is defined as shown in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001901_j.mechmachtheory.2020.103823-Figure5-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001901_j.mechmachtheory.2020.103823-Figure5-1.png", + "caption": "Fig. 5. Slice model of gear pair with MEs.", + "texts": [ + " Therefore, a new slice model of the gear pair with MEs in multi-DOFs is established, and the requirement that the planes of the line of action of the meshing force and profile equations be coplanar when solving for the contact behaviour based on bending, axial compression, shear, foundation and Hertzian contact deformations would be satisfied. This model can calculate the initial meshing position of a gear pair quickly and accurately and does not require many loop iterations or solutions to complicated optimization problems. As shown in Fig. 5 (which is for illustrative purposes only because the position of the gear and pinion on the meshing CS varies with the working angle), because the surface profile equation of the gear in the ideal CS coincides with that of the actual CS, the surface profile equation of the gear represented by E e,g i is parallel to the plane X e O e Y e of the meshing CS. However, for the pinion, the surface profile equation represented by E ea ,p i is not parallel to the plane X e O e Y e of the meshing CS due to the MEs" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000457_j.ijmachtools.2014.05.009-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000457_j.ijmachtools.2014.05.009-Figure2-1.png", + "caption": "Fig. 2. Model of the high speed press system. 1 \u2013 upper part of linkage, 2 \u2013 lower part of linkage, 3 \u2013 middle part of slider, 4 \u2013 left part of rail, 5 \u2013 motor, 6 \u2013 upper part of press frame, 7 \u2013 upper part of left bearing, 8 \u2013 upper part of right bearing, 9 \u2013 right part of rail.", + "texts": [ + " In Section 5, the thermal characteristics of the high speed press are analyzed and discussed, and the validity of the thermal model is verified. Section 6 summarizes the proposed thermal model and the conclusions as well as future work. The physical figure of the high speed press, as shown in Fig. 1, consists of the frame, base, motor, crank shaft, linkage, slider, flywheel, bearings, etc. The solid model of the high speed press system and locations of the measured points on the press structure (numbers 1, 2, 3, \u2026,12) are shown in Fig. 2. The motion of the motor is passed through the belt to the crank shaft. The linkage connects with the crank shaft as well as the slider, which can translate rotary movement of crank shaft into linear reciprocating motion of slider. The upper mold is installed on the slider while lower mold on the working plate. When the sheet metal is put between the upper and lower dies, blanking or other operation techniques can be carried out by the press system. The slider of the press can move or stop optionally as a result of the clutch and brake" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000723_0959651818791027-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000723_0959651818791027-Figure1-1.png", + "caption": "Figure 1. Types of inverted pendulum systems. (a) The x inverted pendulum system. (b) The x-y inverted pendulum system. (c) The x-z inverted pendulum system. (d) Rotary inverted pendulum system.", + "texts": [ + " In general, the vector form of equation (1) can be rewritten in the following matrix form M q\u00f0 \u00de\u20acq+C q, _q\u00f0 \u00de _q+G q\u00f0 \u00de=F q\u00f0 \u00deu \u00f02\u00de where M(q) 2 Rn3 n denotes the inertia matrix, C(q, _q) 2 Rn3 n is the Coriolis and centrifugal matrix, and G(q) 2 Rn is the vector of gravitational forces. Example: inverted pendulum systems. Inverted pendulum systems can be divided into four types: x inverted pendulum systems, x-y inverted pendulum systems, x-z inverted pendulum systems, and rotary inverted pendulum systems: 1. The x inverted pendulum system The x inverted pendulum system shown in Figure 1(a)36 is driven by a horizontal control force. The control objective is to implement the trajectory-tracking control of the car under the condition of maintaining the balance and upright position of the pendulum. The Lagrange equations of the x inverted pendulum system can be expressed as M+m ml cos u cos u l \u20acq+ 0 ml sin u 0 0 _q + 0 g sin u = Fx 0 \u00f03\u00de where q= \u00bd x u T. 2. The x-y inverted pendulum system In contrast to the x inverted pendulum, the x-y inverted pendulum system shown in Figure 1(b)36 is powered by two horizontal control forces applied to the car in the x- and y-directions simultaneously and in a complex way. When the car moves in the x- and y-directions, the pendulum can become stabilized at an unstable equilibrium point. The dynamic model is M+m 0 0 M+m ml cos u 0 ml sin u sinu ml cos u cosu cos u sin u sinu 0 l cosu l 0 0 l cos u 2 6664 3 7775\u20acq+ 0 0 g sin u cosu g sinu 2 664 3 775 + 0 0 ml sin u _u 0 0 0 ml sin u cosu _u+ml cos u sinu _u ml sin u cosu _u+ml cos u sinu _u 0 0 0 l cos u cosu _u 0 0 0 0 2 6664 3 7775 _q= Fx Fy 0 0 2 6664 3 7775 \u00f04\u00de where q= \u00bd x y u u T. 3. The x-z inverted pendulum system The structure of an x-z inverted pendulum system is shown in Figure 1(c);36 here, the car can track the desired positions in the x- and z-directions using two horizontal control forces. However, the pendulum balances only in the x-direction. The Lagrange equation is expressed as M+m 0 ml cos u 0 M+m ml sin u cos u sin u l 2 64 3 75\u20acq + 0 0 ml sin u _u 0 0 ml cos u _u 0 0 0 2 64 3 75 _q+ 0 M+m\u00f0 \u00deg g sin u 2 64 3 75= Fx Fz 0 2 64 3 75 \u00f05\u00de where q= \u00bd x z u T. 4. Rotary inverted pendulum systems The rotary inverted pendulum system, as shown in Figure 1(d),37 contains two main parts: the rotating arm and the pendulum. Relative to other inverted pendulum systems, the rotary inverted pendulum system needs less space, and building a relatively accurate dynamic model is straightforward because of the power transmission mechanism. The rotating arm can be driven by a motor to implement a revolving movement, and this arm directly controls the pendulum rotation. The control objective is to track the expected position of the rotating arm and balance the pendulum at an equilibrium point" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000709_s11071-018-4461-1-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000709_s11071-018-4461-1-Figure1-1.png", + "caption": "Fig. 1 Aerodynamic configuration of the considered AHVs", + "texts": [ + ", ||c|| \u221a cTc; ||d||\u0393 \u221a dT\u0393 d, where d is a vector, \u0393 is a positive matrix with the compatible dimension; sgn(\u00b7) denotes the sign function; tanh(\u00b7) denotes the hyperbolic tangent function. In this section, the COSM of AHVs, which is utilized for design purpose, will be briefly presented. Different from the existing COMs of AHVs, this COSM utilizes a set of sub-models to describe the hypersonic flight dynamics in a more precise way. We consider the longitudinal flight for a class of AHVs, which have the aerodynamic configuration as shown in Fig. 1 and are controlled by a couple of elevators. Since velocity and AOA are the determining factors of aerodynamic force and moment coefficients, we first construct a switching signal \u03c3 , whose rule is decided by the real-time values of velocity and AOA as the \u201cSwitching Rule\u201d block in Fig. 2. It should be emphasized that the thrust of scramjet is very sensitive to AOA, and a drastic decrease in thrust may occur as the increase in the magnitude of AOA.Due to the tight integration between airframe and propulsion of AHVs (as shown in Fig. 1), the geometry of inlet is usually designed for scramjet working under specific ranges of Mach number and AOA. If the magnitude of AOA becomes excessively large during hypersonic flight, the shock wave deviates from the lip of scramjet, and thus, the inlet cannot capture enough air flow tomaintain the efficient supersonic combustion in this situation [8]. Additionally, as pointed by [5,30], the excessively large magnitude of AOA will induce the so-called \u201cinlet unstart\u201d phenomenon, which also has a direct influence on the thrust of scramjet" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000142_s00170-012-4678-y-Figure4-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000142_s00170-012-4678-y-Figure4-1.png", + "caption": "Fig. 4 Circular cutting edge of the inner blade", + "texts": [ + " The vector along the TCE can be derived as Ci s = \u23a1 \u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a3 cos \u03b1i f cos \u03c6i c(sin \u03b1i o sin \u03b3 i f \u2212 cos \u03b1i o sin \u03c6i c cos \u03b3 i f ) cos \u03b3 i f (cos \u03b1i f sin \u03b1i o\u2212 cos \u03b1i o sin \u03b1i f sin \u03c6i c) cos \u03b1i o sin \u03b1i f sin \u03b3 i f \u2212 sin \u03b1i o cos \u03b1i f sin \u03c6i c sin \u03b3 i f \u2212 cos \u03b1i o cos \u03b1i f ( cos \u03c6i c )2 cos \u03b3 i f \u23a4 \u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a6 (3) and its unit vector can be represented as Ui cs . The point width pw describes the width of the blade tip along xi b, and the depth of blade db represents the length of the side cutting edge along zi b (see Fig. 4). From Ui ct and Ui cs , the angle \u03c6i a between them can be calculated as \u03c6i a = arccos(Ui ct \u00b7 Ui cs). The TCE is the line with two ends Ji and Ki , which can be represented in Si b as TCEi b(u1) = (1 \u2212 u1)J i b + u1K i b where 0 \u2264 u1 \u2264 1 (4) and J i b = r cot \u03c6i a 2 Ui ct , K i b = pw 2 Ui ct . The SCE is in line with two ends Li and Mi , which can be represented in Si b as SCEi b(u2) = (1 \u2212 u2)L i b + u2M i b where 0 \u2264 u2 \u2264 1 (5) and Li b = r cot \u03c6i a 2 Ui cs , Mi b = db Z Ui ct Ui cs . So far, the TCE and SCE are determined; the next step is to define the CCE (see Fig. 4). The CCE is an arc on the rake plane 2. It is tangent to TCE and SCE with radius r , and can be obtained by transforming an arc with radius r from xi bz i b plane to 2 plane. Finally, it can be represented in Si b as CCEi b(\u03c6 i)= \u23a1 \u23a2\u23a2\u23a2\u23a3 r cos \u03c6i b cos \u03c6i + XOi br r cos \u03b1i o sin \u03c6i b cos \u03c6i\u2212r sin \u03b1i o sin \u03c6i+YOi br r sin \u03b1i o sin \u03c6i b cos \u03c6i+r cos \u03b1i o sin \u03c6i+ZOi br , \u23a4 \u23a5\u23a5\u23a5\u23a6 (6) where 0 \u2264 \u03c6i \u2264 \u03c0 2 \u2212 \u03c6i a , and here [XOi br YOi br ZOi br ]T are the coordinates of the point Oi r in Si b. It can be calculated from the equation Oi br = r cot \u03c6i a 2 Ui ct + r(Ui ct \u00d7 UE), (7) and \u03c6i b = arcsin( XE\u221a (XE)2+(YE)2+(ZE)2 )" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001563_j.triboint.2019.01.002-Figure11-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001563_j.triboint.2019.01.002-Figure11-1.png", + "caption": "Fig. 11. Test apparatus of rolling bearings: (a) schematic of the test apparatus, and (b) photograph of the test apparatus.", + "texts": [ + " These dynamic characteristics also indicate that when the ball bearing is subjected to combined load, the frequency components of the radial excitation frequency can be reflected in the axial displacement and friction torque spectrum. 3.7. Experimental verification Furthermore, in order to validate the proposed method, an experiment is presented in this section. The measurements of friction torque are completed based on the instrument for the dynamic accuracy of a single rolling bearing described previously [34]. The schematic and photograph of the test apparatus is shown in Fig. 11. The axial load is applied by an air-bearing, and the formation of a gas film between the air flotation plate and the bearing housing reduces the mechanical friction effectively. Hence the bearing housing is free. The force balance method is used to measure the bearing friction torque. The force needed to restrain the bearing housing (outer ring) from rotating is measured by the force sensor, and a tangential connection between the bearing housing and the force sensor is with a flexible rope. Thus, multiplication of the force by the force arm of action point on the bearing housing will give the friction torque of the ball bearing" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001515_j.addma.2019.100892-Figure25-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001515_j.addma.2019.100892-Figure25-1.png", + "caption": "Fig. 25: Optimized C-frame with supports (blue).", + "texts": [ + " The preferable support type in the framework of SLM is the so-called block support which exhibits sufficient strength and heat conduction properties (EOS (2017)). The supports are generated with Materialise Magics, adapted manually according to the requirements and equipped with teethlike substructures at the contact areas with the manufactured part which facilitates their removal in the postprocessing stage. In order to save material, in some areas the supports are angled. Fig. 24 shows an example for such angled supports at the example of the matrix area. Fig. 25 depicts the supported SLM solution centered on the building platform of an EOS M400 SLM system. In order to demonstrate the general manufacturability of the SLM solution including the selected build strategy and support concept, a scaled prototype made of AlSi10Mg has been manufactured on an EOS M290 SLM system (fig. 26). The manufacturing was performed without any interruptions, damages or other problems that might occur during an SLM process which proves the manufacturability of the generated optimized SLM solution" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001897_j.fusengdes.2020.111522-Figure5-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001897_j.fusengdes.2020.111522-Figure5-1.png", + "caption": "Fig. 5. a) Full restriction of the two rotation joints (Case 1). b)Restriction of only joint 2 (Case 2). c)Restriction of the translation joint (Case 3).", + "texts": [ + " This dynamic accuracy algorithm would search in solution space more precisely, and won\u2019t trap into a local minimum. The kinematics of a robot is the foundation for robot control. According to the DH method, the homogeneous matrix Tji represents the relation of jth joint to ith joint, which can be also denoted as a matrix Ai. The forward kinematics of MPD can be easily calculated from Eq. 10. = = +T A s t j i. . 1j i j (10) = \u2219 \u2219\u2219\u2219\u2219\u2219\u2219\u2219 \u2219 \u2219A A A A AT0 11 1 2 9 10 11 (11) There is a position restriction of cask for 1st, 2nd and 3rd joint in Table 3. And these cases map different geometric situations shown in Fig. 5a, 5b and 5c respectively. In Case 1 (Fig. 5a), a prismatic joint restriction of the cask leads to a result that joint 2 and joint 3 cannot have any rotation. When the 1st joint goes forward and the 3rd joint goes into the vacuum vessel completely, the 3rd joint can rotate without any restriction in Case 2 (Fig. 5b). And in Case 3 (Fig. 5c), there is no geometric restriction on the 2nd and 3rd joint, and both joints can move in the workspace freely. Thus, the MPD is a redundant robot that can \u2018freely\u2019 position and orient an object in the Cartesian workspace [23]. Some DOF restrictions, such as self-collision free and surrounding collision-free, should be considered so that solutions can be reasonable. To simplify the model, a novel method, which is dividing the whole arm into two parts, analyzes the MPD\u2019s inverse kinematics (Fig. 7)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000423_0954405414553979-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000423_0954405414553979-Figure1-1.png", + "caption": "Figure 1. Interference of the cutter with the adjacent blade in flank milling.", + "texts": [ + " First, with the increasing demand for performance designs, components with complex geometries and surfaces are required to be manufactured in tight tolerance; therefore, the machining error should satisfy the precision requirement. Second, for widely utilized ballend cutters, the ball end of the cutter should be tangential to the hub surface to avoid interference.11 Third, the cutter should be interference-free with the adjacent blade. As the distance between two adjacent blade surfaces is relatively small in some centrifugal impellers, interference between the cutter and adjacent blade may occur if an unsuitable cutter is chosen, as shown in Figure 1. In order to generate interference tool paths, some strategies have been proposed. Tsay et al.12 developed an algorithm to generate tool paths with global interference checking for five-axis point milling of turbo-machinery components. The interference between the surface of the work-piece and the cutter can be detected based on the projected distance. Lee and Suh13 presented an interference-free tool path planning method to machine twisted ruled surfaces. Increasing the stiffness of the cutter can effectively improve the rigidity of the whole cutting system and consequently reduces the deflection and vibration in the milling process", + " As illustrated in Figure 9, dpi,S0 (w) r0 =0 means that the ball end of the cutter is tangential to the hub surface. For a set of tool reference points fpi 2 R 3, i=1, . . . , n2g, we have the following constraint functions C2 dpi,S0 (w) r0 4e i=1, . . . , n2 \u00f022\u00de where e is the prescribed geometric tolerance. Constraint of collision avoidance of the cutter with the adjacent blade One of the major problems outlined in five-axis flank milling of impellers is the interference between the cutter and the adjacent blade, which is usually caused by an unsuitable cutter size, as shown in Figure 1. For a set of data points fpi 2 R 3, i=1, . . . , n3g sampled from the adjacent blade surface, as shown in Figure 10, we have the following constraint functions C3 dpi,SAE (w)50 i=1, . . . , n3 \u00f023\u00de Model and algorithm for optimization of cutter geometry According to the analysis presented in sections \u2018\u2018Stiffness of a conical cutter\u2019\u2019 and \u2018\u2018Geometric constraints,\u2019\u2019 we propose the following model for optimal design of tool path and shape considering stiffness and multi-constraint. The objective is to improve tool stiffness" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002131_j.wear.2021.203859-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002131_j.wear.2021.203859-Figure3-1.png", + "caption": "Fig. 3. Schematic of the test specimen configuration in the hot strip drawing tribometer.", + "texts": [ + " The shot blasted surfaces had an areal average surface roughness value nearly four times higher compared to the milled and ground surfaces. The reference hot work tool steel was provided with a ground surface having an areal average surface roughness of 0.19 \u00b1 0.45 \u03bcm since this is the typical state for a refurbished tool for hot stamping. The tribological tests were performed using a hot strip drawing tribometer, which enables tests under conditions prevalent in press hardening. A detailed description of the test equipment is given in Ref. [25]. In the tribometer (Fig. 3), the strip specimen is mounted vertical on its long edge, to prevent accumulation of wear debris, and clamped by two hydraulic jaws and held straight by a pneumatic pre-tension cylinder. The two tool steel pins are mounted in a moving tool assembly driven by a servomotor and ball screw. The tool steel pins are loaded against the strip by means of a pneumatic bellow. The strip specimen is heated via resistive heating (Joule effect) by passing a current through it. The tool steel pins can be heated independent from the strip by means of resistive heating coils placed inside the holders" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000925_s11044-016-9500-4-Figure4-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000925_s11044-016-9500-4-Figure4-1.png", + "caption": "Fig. 4 Force simulation model of the 2RPU+SPR mechanism", + "texts": [ + " Then, the parameters of the cross-section of links R1P1, R2P2 and SP3 can be obtained by Ai1 = \u03c0d2 i1/4, Ii1 = \u03c0d4 i1/64, Ipi1 = \u03c0d4 i1/32, and the parameters of links P1U1, P2U2 and P3R3 can be obtained by Ai2 = \u03c0d2 i2/4, Ii2 = \u03c0d4 i2/64 and Ipi2 = \u03c0d4 i2/32. In the initial configuration, the moving platform is parallel to the base, l12 = l22 = 200 mm, l32 = 205.5 mm, and \u03b81 = \u03b82 = 10.16\u25e6. The external wrench imposed on the moving platform expressed in the reference coordinate frame O-XYZ is /SF = (10 N 10 N \u2212 10 N 10 N\u00b7m 10 N\u00b7m 10 N\u00b7m)T. By resorting to the Adams simulation software, the force simulation model of the 2RPU+SPR mechanism is built up, as shown in Fig. 4, in which all links within each limb are built as flexible bodies, and the moving platform, base and all kinematic joints are considered as rigid bodies. Then, both the theoretical values of all constraint forces/couples (including the driving forces) of the 2RPU+SPR mechanism solved by Eq. (32) and the corresponding simulation values measured by the simulation model are listed in Table 2. It can be seen from Table 2 that the maximum error between the theoretical values and the simulation values is less than 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000893_j.ymssp.2015.10.028-Figure7-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000893_j.ymssp.2015.10.028-Figure7-1.png", + "caption": "Fig. 7. Isometric representation of the overall meshed model.", + "texts": [ + " The axial degree of freedom is restrained at the gearbox housing, because the axial force between the sun and the ring gear is transmitted via the housing. In the real construction the ring gear and the gearbox housing are connected by bolted connections via frictional contacts. The influence of the bolted joints' drill hole on the strain condition at the outer surface of the ring gear is considered. However, the modeling of frictional contact is not part of this research project. An overview of the FE-Model is given in the Fig. 7. The simulation is split into two parts in order to study the influence of different ring gear geometries. In the first model the drill holes are neglected (referring to Fig. 8). For the contact boundary condition between the ring gear and the gearbox housing a tie-constraint is chosen. In the real construction, the ring gear is attached to the housing via bolted joints. By using the tie-constraint, friction between housing and ring gear as well as stresses introduced by the bolted joints are neglected" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001108_tia.2018.2829769-Figure15-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001108_tia.2018.2829769-Figure15-1.png", + "caption": "Fig. 15. The 3D-FEM model of the 120 W SyRM motor.", + "texts": [ + " In the calculations, the material characteristics taking into account the punching effect were used to model the stator teeth and the rotor core, whereas the new measured characteristics of the sample yoke material (taking into account clamping or welding) were used for modeling the stator yoke. As mentioned before two fractional kilowatt motors have been studied using the 3D-FEM models; they are rated 120 W and 550 W, respectively. The machines use the same stator and rotor laminations, but they have different core lengths (35 mm and 90 mm, respectively) and different turn number in the three-phase stator winding. Their rotor structure is the conventional geometry used for line start motors, characterized by salient pole pieces \u2013 see Fig.15. The analyzed motor has unskewed anisotropic rotor, and a shorter core axial length compared to the end winding length. Therefore, a 3-D model is mandatory for such motor construction. For example, in [19] it is recognized that 2-D FEM models lead to significant errors in the estimations of the winding linked fluxes. Computer simulations were carried out using the OPERA 3D commercial package. The field-circuit modeling, where the integral parameters in the circuit model are calculated from the field model, has been adopted" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000769_978-3-662-43893-0-Figure1.7-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000769_978-3-662-43893-0-Figure1.7-1.png", + "caption": "Fig. 1.7 Principle of a flap drive mechanism", + "texts": [ + " This is achieved by means of a central shaft driving individual angular gear sets and 6 1 Fields of Application for Bevel Gears a crank mechanism moving the flap backwards horizontally and at the same time increasing their angle of attack to the airflow. For aerodynamic reasons, the wings of a modern aircraft are backswept and the airfoil varies along the wing. Because of flight safety requirements, all flap movements must originate from a central shaft. If the trailing edge of the wing is not straight, the central shaft actuating the flaps must be interrupted in several places. Figure 1.7 shows the rotary mechanisms, cranks and bevel gear drives used to actuate the flaps. At each change in direction along the central shaft, a bevel gear set is used to transmit the rotary motion which moves the flap. Unlike the high speed bevel gears in the turbine, flap drive bevel gears perform only slow speed servo movements and therefore require a completely different design. The classic power train concept for large vessels where the rudder is installed behind a propeller mounted on a shaft driven by the ship\u2019s engine is becoming less and less common" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002015_j.asr.2020.09.040-Figure11-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002015_j.asr.2020.09.040-Figure11-1.png", + "caption": "Fig. 11. Basic layouts of spiral folding pattern for membrane parameters consider N \u00bc 6;R0 \u00bc 17:5 mm; t \u00bc 0:1 mm; S0 \u00bc 59:33 mm.", + "texts": [ + "e denotes the ceil function, the term k denotes the half of the edge of the concentric polygon with apothem R 2 R0;Rf . For a regular polygon with N edges k \u00bc R tan p N \u00f021\u00de The modulo operator in Eq. (20) renders either 0 or 1, thus for apothem R 2 R0;Rf , segments of Fold A0 and Fold B are outlined as mountain (F \u00bc 0) or valley fold (F \u00bc 1). The solution of the system of differential equations in Eq. (14) to Eq. (16) enables us to render spiral folding pat- terns whose generalization to other regular geometries and lengths is straightforward, as Fig. 10 and Fig. 11 show. We extend the above-mentioned principles to develop the multi-spiral folding patterns. The basic idea of our approach is portrayed by Fig. 12. Here, in the left side of Fig. 12, we portray the spiral folding on a single membrane, in which the orientation is deemed to be counterclockwise; and in the right side of Fig. 12, we show the basic concept of our proposed approach combining two types of spiral folding patterns: (1) the counterclockwise spiral pattern as shown by the membranes with blue and green color in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001897_j.fusengdes.2020.111522-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001897_j.fusengdes.2020.111522-Figure3-1.png", + "caption": "Fig. 3. DH coordinate system established for MPD.", + "texts": [ + " In Section 4, the procedure of the MPD\u2019s IK is introduced in detail, such as forward kinematics, inverse kinematics and collision. The results and conclusions are shown in Section 5. The fundamental analysis of a robot is building up coordinate systems by the Denavit\u2013Hartenberg (DH) convention and find the DH parameters [19]. There are 2 branches\uff1athe standard DH method and the modified DH method. The difference between them is the sequence defining \u03b1, a, d, and \u03b8, which are the DH parameters of a robot. Using the standard DH method, the coordination of MPD can be established in Fig. 3, and its DH parameters are listed in Table 1. Accordingly, a relationship in adjacent coordinates (coordinate i bonded to joint i and coordination i-1 bonded to joint i-1) can be given by a homogenous transformation matrix Ai = \u23a1 \u23a3 \u23a2 \u23a2 \u23a2 \u2212 \u2212 \u23a4 \u23a6 \u23a5 \u23a5 \u23a5 \u03b8 \u03b8 \u03b1 \u03b8 \u03b1 a \u03b8 \u03b8 \u03b8 \u03b1 \u03b8 \u03b1 a \u03b8 \u03b1 \u03b1 d A cos sin cos sin sin cos sin cos cos cos sin sin 0 sin cos 0 0 0 1 i i i i i i i i i i i i i i i i i i (1) where \u03b8i, \u2212ai 1, di, and \u2212\u03b1i 1 are link parameters of joint i. The CFETR\u2019s geometric model can be rebuilt by the point cloud", + " The specific procedure of the ACO of IK will be introduced in the next section. This paper adopts a geometric & algebra method to analyze this problem. According to the MPD structure, the whole robot is divided intotwo parts: the first part from joint 0 to joint 7, and the rest joints as the second part, such that the kinematics solutions for the joints in each part can be calculated easily. Meanwhile, there is a good geometric feature that the origin of coordination 6 and coordination 7 is coincident as shown in Fig. 3. The forward kinematic relationship in the two parts is listed in Eq. 12. \u23a7 \u23a8 \u23aa \u23aa \u23a9 \u23aa \u23aa = \u2219 = \u2219 = \u2219 \u2219 \u2219 = \u2219 \u2219 \u2219 \u2219 \u2219 \u2219 = \u2219 \u2219 \u2219 \u2219 \u2219 T T T T T T T T T T T T T T T T T T T T T T T T T T END MPD CFETR 11 0 11 0 7 0 11 7 11 7 8 7 9 8 10 9 11 10 7 0 1 0 2 1 3 2 4 3 5 4 6 5 7 6 6 0 1 0 2 1 3 2 4 3 5 4 6 5 (12) An iteration procedure is in Fig. 6. Through setting different target matrices for two parts, the DAACO, which is an intelligent ant colony optimization algorithm, is applied to analyze each part respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000423_0954405414553979-Figure10-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000423_0954405414553979-Figure10-1.png", + "caption": "Figure 10. Constraint of collision avoidance of the cutter with the adjacent blade: (a) Three-dimensional sketch map and (b) Twodimensional sketch map.", + "texts": [ + " , n2g, we have the following constraint functions C2 dpi,S0 (w) r0 4e i=1, . . . , n2 \u00f022\u00de where e is the prescribed geometric tolerance. Constraint of collision avoidance of the cutter with the adjacent blade One of the major problems outlined in five-axis flank milling of impellers is the interference between the cutter and the adjacent blade, which is usually caused by an unsuitable cutter size, as shown in Figure 1. For a set of data points fpi 2 R 3, i=1, . . . , n3g sampled from the adjacent blade surface, as shown in Figure 10, we have the following constraint functions C3 dpi,SAE (w)50 i=1, . . . , n3 \u00f023\u00de Model and algorithm for optimization of cutter geometry According to the analysis presented in sections \u2018\u2018Stiffness of a conical cutter\u2019\u2019 and \u2018\u2018Geometric constraints,\u2019\u2019 we propose the following model for optimal design of tool path and shape considering stiffness and multi-constraint. The objective is to improve tool stiffness. The geometric constraints are to meet the machining accuracy and interference-free requirements P1 max w2Rm+2 k s:t: dpi,SEE (w) 4d, i=1, " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000080_j.engfailanal.2013.03.008-Figure14-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000080_j.engfailanal.2013.03.008-Figure14-1.png", + "caption": "Fig. 14. Shape of cracked gear (5th order, a = 0.8, root crack).", + "texts": [], + "surrounding_texts": [ + "Assuming a crack length of a = 0.8 mm at the pitch circle, when the other parameters are fixed, and only the magnitude of the load is changed, the relationship between load and SIF can be obtained by analysis as shown in Figs. 23 and 24. From these figures, it can be seen that the relationship between load and SIF is linear, that is to say: KI = A1w, KII = A2w, where A1 and A2 are coefficients that are relevant to the main parameters of the gear, position of load, crack length, and so on. From the figures, it can also be seen that the rate of changed for KI, KII is fastest when the load acts at point E." + ] + }, + { + "image_filename": "designv10_14_0000771_978-3-319-16823-4-Figure7.6-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000771_978-3-319-16823-4-Figure7.6-1.png", + "caption": "Fig. 7.6 Evolution of \u2018plasmodium\u2019 positions and interaction fronts in the particle model for the G2 gate with inputs 011. Top Row: Particle positions. Bottom Row: Chemoattractant gradient. Arrows indicate propagation of gradient from food sources. Dashed arcs represent boundary regions separating competing gradients. Dashed circles represent diffusion from food sources suppressed by engulfment See text for explanation.", + "texts": [ + " The dynamical gradient interface represents a fragile boundary between two separate swarms, two separate food gradients or a combination of both swarm and food gradients. The third criterion \u2014 fusion of plasmodia can be represented in the particle model when movement of separate particle paths is limited and perturbation of the dynamic boundary occurs. This can result in fusion of network paths which corresponds to fusion of plasmodia. The complex evolution of gradient fields can be seen in an example run of G2 with the inputs 011 in Fig. 7.6. The top row shows the particle positions and the bottom row shows the chemoattractant gradient field enhanced by a local method of dynamic contrast enhancement. The first column shows the propagation of chemoattractant gradient from the two food sources and the interfacial region (dashed arcs). Note that the gradient from the right suppresses the gradient from the bottom source. The second column shows 134 7 Approximating Classical Computing Devices with the Multi-agent Model the effect of suppression of the rightmost food source when engulfed by the particle population which has migrated towards it" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002093_s40194-020-01043-6-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002093_s40194-020-01043-6-Figure1-1.png", + "caption": "Fig. 1 Schematic representation of the WAAM deposition process of Inconel 625 (a), samples fabricated by CMT-WAAM with the short-circuiting mode of droplet transfer (b), deposited with short-circuiting with pulse mode of droplet transfer (c), and conventionally cast pipe (d)", + "texts": [ + "999% purity) using a Fronius-TranspulsTM synergic 4000 CMT welding power source in short-circuiting and short-circuiting with pulse mode of droplet transfers with optimised parameters given in Table 2. Samples manufactured using the short-circuiting mode of droplet transfer consisted of 13 layers with a total height 50 mm and width of 12 mmwhile samples manufactured with short-circuiting with pulse mode of droplet transfer consisted of 13 layers with a total height of 60 mm and width of 20 mm (Fig. 1). Samples produced by WAAM were compared with samples extracted from a commercial centrifugally cast Inconel 625 pipe (Fig. 1d). During the multi-pass deposition, it was ensured that the previously deposited layer was completely cooled to room temperature. Before the onset of next deposition, the previously deposited layer was cleaned with a stainless steel wire brush. An InfiniVisionTM digital storage oscilloscope was used to record the current and voltage waveforms during the deposition with a recording frequency of 70 kHz. A Photon FASTCAM high-speed camera was used to observe melt pool stability and droplet transfer behaviour during the multi-pass deposition" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002124_s11665-021-05591-w-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002124_s11665-021-05591-w-Figure2-1.png", + "caption": "Fig. 2 SLM scanning strategy", + "texts": [ + " The particle size distribution is displayed in Fig. 1(b). DV(10), DV(50), and DV(90) of the powder are 16.2 lm, 37.7 lm, and 79.1 lm, respectively. And the result of EDS characterization is shown in Fig. 1(c), it can be confirmed that the content of silicon is about 3% (wt). A series of cubic specimens with dimensions of 8 mm 9 8 mm 9 8 mm were prepared using a selective laser melting machine MCP Realizer SLM 250 system (MCP-HEK Tooling GmbH, Germany). It was equipped with an Nd-YAG fiber laser. As described in Fig. 2, a zigzag scanning pattern with the neighboring laser track rotation of 90 was utilized during the SLM process. Five different laser powers (P, W) with 50 W, 70 W, 100 W, 140 W, 180 W were used to prepare a series of SLM cubic samples. The laser spot (lm), laser scanning speed (V, m/s), hatch distance (lm), and layer thickness (lm) were kept in a constant value with 35 lm, 0.33 m/s, 60 lm, and 30 lm, respectively. To avoid oxidation, the chamber was adopted under an argon atmosphere to maintain the content of residual oxygen content below 1" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001304_j.mechmachtheory.2019.103697-Figure8-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001304_j.mechmachtheory.2019.103697-Figure8-1.png", + "caption": "Fig. 8. End position coordinate systems of force-sensitive elements.", + "texts": [ + " C fj denote the end flexibility matrices of part1, part2, part3. T ig f j denotes the corresponding coordinate transformation matrix, shown as follows, T ig f j = [ E 3 \u00d73 S ( r ig f j ) 0 E 3 \u00d73 ] (16) Here, r ig f 1 = ( 0 0 \u22122 r \u2212 l 1 ) T , r ig f 2 = ( 0 0 \u22122 r \u2212 2 l 4 ) T , r ig f 3 = ( 0 0 0 ) T C ig ( i =1 , 2 ... 6 ) can be obtained through substituting Eq. (16) into Eq. (15) . 3.4. Stiffness matrix modeling of the six-axis force/torque sensor The end position coordinate systems of force-sensitive elements are shown as Fig. 8 . o p x p y p z p is the reference coordinate and located above the surface center of the upper platform. o ig x ig y ig z ig ( i = 1 , 2 , . . . , 6 ) are the end local coordinate systems of force-sensitive elements. The corresponding axes of o p x p y p z p and o 1 g x 1 g y 1 g z 1 g are parallel. The direction of axes z ig are perpendicular to the end surface of force-sensitive elements. The direction of axes x ig ( i = 1 , 2 , 3 ) are outward from the point o p . The direction of axes x ig ( i = 4 , 5 , 6 ) are vertically upward" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000067_s00170-015-7171-6-Figure5-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000067_s00170-015-7171-6-Figure5-1.png", + "caption": "Fig. 5 Intersection between the cutting surface and the axial plane", + "texts": [ + " The cutting surface generated with the tooth cutting edge of the inside blade is called inside cutting surface CSi(l,\u03bb) and is formulated with parameters l and \u03bb as CSi l;\u03bb\u00f0 \u00de \u00bc X i CS Y i CS Z i CS 2 4 3 5 \u00bc hb\u2212l\u22c5ei1 ei2\u22c5l\u22c5cos\u03bb\u2212e i 3\u22c5l\u22c5sin\u03bb\u00fe ei4\u22c5sin\u03bb ei2\u22c5l\u22c5sin\u03bb\u00fe ei3\u22c5l\u22c5cos\u03bb\u2212e i 4\u22c5cos\u03bb 2 4 3 5 \u00f07\u00de where ei1 ei2 ei3 ei4 2 664 3 775 \u00bc \u2212cos \u03b1i r \u22c5cos \u03b1i h \u22c5cos \u03b2i tcp sin \u03b1i r \u22c5cos \u03b1i h \u22c5sin \u03b2i tcp \u2212cos \u03b1i r \u22c5sin \u03b1i h \u22c5cos \u03b2i tcp cos \u03b1i r \u22c5cos \u03b1i h \u22c5sin \u03b2i tcp Rac\u2212 1 2 \u22c5Pw 2 66666664 3 77777775 \u00f08\u00de where parameter l is within the range of [0,cl i/[cos(\u03b1r i) \u22c5 cos(\u03b1h i ) \u22c5cos(\u03b2tcpi )]], and the parameter \u03bb is the rotation angle about the Xch axis in the range of [0,2\u03c0]. To find the mean point projection Mi on the inside flank of the cutting surface profile, the Xch\u2013Zch plane intersects the cutting surfaces at the cutting surface profile H-I-J-K (see Fig. 5). The point Mi on the inside flank H-I can be found with their parameters as l \u00bc hM ei1 \u00f09\u00de \u03bb \u00bc tan\u22121 ei2 ei3\u2212ei4 \u22c5l \u00f010\u00de Thus, the coordinates xchMi ; ychMi ; zchMi h iT of point Mi in the cutter head coordinate system can be calculated by substituting Eqs. 9 and 10 into Eq. 7. As the core of the CNC programming model for gear face milling, the formulation of the cutter system location and orientation is established here. The geometric principle of this model is that the cutter system is located and oriented with respect to the gear so that the profile of a gear tooth slot at its mean point is coincided with a cutting surface profile of the cutter system as shown in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000735_s00521-018-3795-4-Figure10-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000735_s00521-018-3795-4-Figure10-1.png", + "caption": "Fig. 10 Coordinates system and degree of freedom of UAV", + "texts": [ + " However, there exists a certain error in T\u2013S FCMAC. The proposed IRWTC achieves a favorable tracking response, and the errors are much reduced. Table 2 shows the comparison result of root mean square error (RMSE) between the T\u2013S FCMAC and the proposed IRWTC. From these comparisons, our findings indicate that the proposed IRWTC achieves stability and desired control performance, and the number of hypercubes is adjusted to obtain favorable control performance. Coordinate system and degree of freedom of UAV are shown in Fig. 10. Aircraft contains six degrees of freedom, and based on its motion it is divided into the longitudinal motion and the lateral motion. There are three control surfaces in the aircraft, one is used to control longitudinal motion and the other two are used to control lateral motion. All of the symbol meanings are given in Table 3, physical meanings of longitudinal partial derivatives are shown in Table 4, physical meanings of lateral partial derivatives are shown in Table 5, and the data of UAV are given in Table 6" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000769_978-3-662-43893-0-Figure6.32-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000769_978-3-662-43893-0-Figure6.32-1.png", + "caption": "Fig. 6.32 Position of the lapping nozzles", + "texts": [ + "31 Material removal by means of lapping 286 6 Manufacturing Process The jet of lapping medium is not usually fed directly into the meshing point but, depending on the direction in which the gear set is rotating, ahead of it. By directing the lapping jet at the toe of the wheel, the spinning effect distributes the lapping compound conveniently over the face width. The lapping nozzles are adjusted radially such that two-thirds of their crosssection overlaps with the face width. The distance from the nozzles to the gear teeth should be about 10\u201315 mm, in the direction of the wheel shaft (Fig. 6.32). As established in Sect. 3.4.4, the tool radius crucially affects the load-free displacement behavior of a gear set. For lapping, this means that with a large tool radius the contact zone is easy to change, whereas with a relatively small tool radius this is possible to only a limited extent. The hypoid offset influences the local sliding velocities which vary over the tooth flank (see Sect. 2.4.3). Since pure rolling takes place on the pitch cone (Fig. 6.33) of a spiral bevel gear set without offset, scarcely any lapping removal is possible about this area such that if lapping time is too long, a clear elevation will appear about the pitch cone" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000718_j.jsv.2018.07.037-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000718_j.jsv.2018.07.037-Figure2-1.png", + "caption": "Fig. 2. Torsional dynamic model of multistage gear transmission system.", + "texts": [ + " Using Poincar e section analyses the evolution process, instability condition, and the influence of fixed-axis gear tooth crack to the sub-harmonic resonance and stability of the system. The system studied in this paper is a test rig of the multistage gear transmission systemwhich contains a two-stage fixedaxis gear and a one-stage planetary, as showed in Fig. 1. where spur gears 1, 2 compose the 1st stage fixed-axis gear for the input, spur gears 3, 4 compose the 2nd stage fixed-axis gear, the planet carrier is for the output. The torsional dynamic model is established by using the lumped mass method as showed in Fig. 2. The model does not consider the transverse vibration displacement of gears. Gear parameters are simulated with a spring and a damper. where, qs, qc, qpn, q1, q2, q3, q4 represent the angular displacement of sun gear, planet carrier, planetary gear n (n\u00bc 1, 2, 3, 4), spur gears 1, 2, 3, 4, respectively. Throughout this paper, the subscripts s, c, pn, r, 1, 2, 3, 4 denote sun gear, planet carrier, planetary, ring gear and spur gears 1, 2, 3, 4. Quantities rs, rc, rpn, r1, r2, r3, r4 are the base circle radius of gears" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002192_s12541-021-00556-4-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002192_s12541-021-00556-4-Figure2-1.png", + "caption": "Fig. 2 KR500L340-2 kinematic model", + "texts": [ + " (25) \u23a7\u23aa\u23a8\u23aa\u23a9 i1 \u2208 [ i1min, i1max] i2 \u2208 [ i2min, i2max] \u22ee i120 \u2208 [ i120min, i120max] According to Sect.\u00a02, the kinematic model and kinematic error model of the robot KUKA KR500L340-2 are established. To simulate the actual work environment, the drilling robot is imposed constraints on each joint as shown in Table\u00a01, due to the limited measuring range of the measuring device or the obstructions around the drilling robot, some points among the optimal measurement configurations cannot be obtained. The kinematic model and nominal kinematic parameters of the drilling robot are presented in Fig.\u00a02 and Table\u00a01 respectively. It\u2019s assumed that the transformation parameters between the base frame and the measuring frame, the flange frame and the end-effector frame of the drilling robot are shown in Table\u00a02. Set the scale of swarm m = 30 , the maximal iteration time T = 200 , the particle velocity boundary Vmax = 10 , the upper and lower boundary of the inertia weight min = 0.4 and max = 0.9 . The proposed improved PSO algorithm in this paper and the traditional PSO algorithm are used to find 20 1 3 optimal measurement configurations of the drilling robot in a limited workspace respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000769_978-3-662-43893-0-Figure4.38-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000769_978-3-662-43893-0-Figure4.38-1.png", + "caption": "Fig. 4.38 Displacements of the gear axes", + "texts": [ + " The factors include displacements and tilting of the pitch cones due to: \u2013 deflection of the shafts, bearings, gear bodies and housing \u2013 bearing clearance \u2013 positional deviations of the bearing bores caused in manufacturing \u2013 errors in gear assembly Load-dependent deflections are caused by applied loads including torque, tooth forces of the bevel gear stage, and additional loads on the shafts such as forces of a cylindrical gear stage, shear forces and bending moments at the ends of the shafts. The relative positions of the pinion and wheel are, for example, described by the four possible forms in axis displacement and their deviations in a hypoid bevel gear set (Fig. 4.38): \u2013 shaft angle \u03a3 shaft angle deviation \u0394\u03a3 \u2013 hypoid offset a hypoid offset deviation \u0394V \u2013 pinion apex distance tz1 pinion axial displacement \u0394H \u2013 wheel apex distance tz2 wheel axial displacement \u0394J 180 4 Load Capacity and Efficiency The non-linear relationship between load and deflection can be taken into account by section-wise linearization or by an iterative procedure. The maximum normal stress acting in the contact area between meshing tooth flanks, also known as Hertzian stress, is employed as an equivalent variable in characterizing tooth flank stress, as in [DIN3990, DIN3991, ISO6336, ISO10300] (Fig", + " Combining the greatest possible effective total contact ratio in low load operation with adequate displacement capability in full load operation is an important goal for the development of bevel gear micro geometry. 208 5 Noise Behavior When calculating the amount of crowning required, it is necessary to account for the \u201cenvironment\u201d of the gear set (see Sect. 4.4.3.3), where displacements are analyzed in four directions, i.e. variations along the pinion axis, along the wheel axis, along the pinion offset and of the shaft angle (Fig. 4.38) [HAGE71]. Crowning must be chosen in such a way that tooth flank damage due to unfavorable load concentration does not occur in a representative load spectrum for the gear set (Fig. 5.16). The way in which crowning is designed has a dominant influence on the running noise of a gear set. A number of important relationships are accordingly shown in Figs. 5.17, 5.18 and 5.19. The gear set being modified is a randomly selected bevel gear with 35 mm face width and 10 mm whole depth, of the kind generally used in passenger car rear axle units" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001761_012028-Figure4-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001761_012028-Figure4-1.png", + "caption": "Figure 4. DC HTS dynamo concept to be integrated with the field coil. The right image shows the device with the HTS and sapphire support removed to illustrate the rotating magnet.", + "texts": [ + " The field coil is circular in shape with a cross-section of 40 mm x 8 mm and has a mean radius of 212 mm. It has 60 turns of 3 mm wide REBCO conductor \u2013 12 turns/layer and 5 layers. The overall dimensions of the coil cryostat are 70 mm x 33 mm. Inductance of the field coil is 1.19 mH. Field ICMC 2019 IOP Conf. Series: Materials Science and Engineering 756 (2020) 012028 IOP Publishing doi:10.1088/1757-899X/756/1/012028 currents at no-load and rated full-load are 188 A and 364 A, respectively. The dynamo to be integrated with this coil is shown in Figure 4. and will be capable of managing field current over this range. Description Current Leads Dynamo Thermal conduction through cryostat, W 30 30 Thermal conduction through exciter, W 36 4 Total thermal load, W 66 34 Power input to refrigerator, kW 1.94 1.05 Weight of refrigerator, kg 4 2 ICMC 2019 IOP Conf. Series: Materials Science and Engineering 756 (2020) 012028 IOP Publishing doi:10.1088/1757-899X/756/1/012028 The AC homopolar concept currently represents least risky option for aircraft applications in near term" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002190_s43452-021-00242-2-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002190_s43452-021-00242-2-Figure2-1.png", + "caption": "Fig. 2 Bending stiffness boundary condition", + "texts": [ + " The static stiffness involves bending stiffness and torsional stiffness. For the bending stiffness\u2019s boundary condition, the front left and right shock absorber point is constrained for the second and third degree of freedom, respectively, the shock absorber point of rear left and right are constrained for the first, second, third degree of freedom, and the first, third degree of freedom, respectively, meanwhile, vertical forces of 1000\u00a0N is imposed at the center of longitudinal beam, as shown in Fig.\u00a0 2a. Meanwhile, displacement measure points are arranged at 100\u2013150\u00a0mm intervals on top of the beam, as shown in Fig.\u00a02b. For the torsional stiffness\u2019s boundary condition, the front left and right shock absorber point are constrained for the first and second degree of freedom, respectively, the shock absorber point of rear left and right are constrained for the first, second, third degree of freedom and the first, third degree of freedom, respectively, as shown in Fig.\u00a03(a). Meantime, 1689\u00a0N vertical forces is imposed at the front shock absorber, it is equivalent to applying a torque of 2000Nm on the front end. Meanwhile, measure 1 3 the vertical downward displacement of the loading point, as shown in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000663_s40195-017-0668-2-Figure5-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000663_s40195-017-0668-2-Figure5-1.png", + "caption": "Fig. 5 Schematic diagram of the mechanical anisotropy", + "texts": [ + " In the second grade, r0.2, rb, the elongation and reduction of the area in x direction are 485, 1050 MPa, 0.67 and 1.9%, respectively. However, r0.2, rb, the elongation and reduction of area in z direction are 469, 761 MPa, 10 and 8.1%, respectively. The similar phenomenon was observed in the third grade. The tensile strength (rb) increased gradually from the first grade to the third grade in x direction. However, rb was not increased in z direction. The phenomenon can be explained in the following parts as shown in Fig. 5. Mechanical anisotropy in different grades is characterized. The values of r0.2 and rb measured in x direction were higher than those in z direction. An inverse relationship for elongation and reduction of the area was observed. The reason for the mechanical anisotropy can be explained by two factors as shown in Fig. 5. For the alloys that were built in z direction, the weak metallurgical bonding between layers accounted for the relatively lower values in tensile properties and higher ductility [35]. As shown in Fig. 5, the metallurgical bonding in x direction is track\u2013track bonding like fried dough twists. The metallurgical bonding in z direction is layer\u2013layer bonding like plane bonding. Another factor is that the pores in the DLMD samples which is known as lack-of-fusion pores and is extremely harmful to mechanical properties [31]. The loading direction is normal to the orientation of pores. Therefore, the pores tend to be torn apart from the samples in z direction. The loading is nearly parallel to the flat pores imposing compression stress upon the pores locally, which helps seal and close the pores for the samples in x direction [41]. Fracture morphologies of the tensile samples are shown in Fig. 6. As for the 1-x and 1-z samples, it is observed that the primary fracture mechanism is ductile mode, and the mean size of the dimples is similar to that of cellular grains. The undesirable closed pores on a scale of 10\u201320 lm are observed on the fracture surfaces (Fig. 6a\u2013d) which were explained in the schematic diagram of Fig. 5. Table 2 Mechanical properties of the samples Materials Yield strength, r0.2 (LPa) Tensile strength, rb (LPa) Elongation (%) Reduction of area, w (%) First grade-x 529 \u00b1 30 951 \u00b1 79 6.8 \u00b1 1.5 8.4 \u00b1 3.9 First grade-z 501 \u00b1 19 765 \u00b1 20 12.0 \u00b1 0.17 10 \u00b1 1.7 Second grade-x 485 \u00b1 40 1050 \u00b1 91 0.67 \u00b1 0.39 1.9 \u00b1 1.4 Second grade-z 469 \u00b1 38 761 \u00b1 49 10 \u00b1 0.57 8.1 \u00b1 1.0 Third grade-x 675 \u00b1 114 1056 \u00b1 207 0.87 \u00b1 0.55 1.1 \u00b1 0.99 Third grade-z 507 \u00b1 23 721 \u00b1 30 12 \u00b1 0.77 8.9 \u00b1 0.42 LENS SS16L [33] (x direction) 576 776 33 \u2013 LENS SS16L [33] (z direction) 479 703 46 \u2013 LMDS part [35] (z direction) 558 639 21 \u2013 LMDS part [35] (x direction) 352 536 46 \u2013 DLD/Single-Built 316L(as-built) [39] 405\u2013415 620\u2013660 32\u201340 \u2013 DLD/Single-Built 316L(heat-treated) [39] 325\u2013355 600\u2013620 42\u201343 \u2013 DLD/Nine-Built (as-built) [39] 465\u2013485 660\u2013685 12\u201320 \u2013 These pores show smooth surfaces decorated with unmelted or partially melted powder particles suggesting that this layer was not completely remelted during the subsequent building, which was also reported by other researchers [34, 41, 42]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002206_s00202-021-01338-x-Figure4-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002206_s00202-021-01338-x-Figure4-1.png", + "caption": "Fig. 4 Magnetic flux distribution of AFPMSG models with different magnet shapes: a triangular, b trapezoidal, c skewed, and d roundedtriangular", + "texts": [ + " To provide an equal condition to compare the performance of the studied models, in addition to the same topology, the volume of the magnets remains constant for the four magnet shapes (Figs. 2 and 3). All four models are analyzed under similar nodal and rotational conditions. The rated rotational speed for all models is 450 rpm. Nonlinear 3D-FEA is employed using a solver, which is then used to compute the electromagnetic fields of the moving systems in all three dimensions. Because of the difference in magnet shapes, each model has its own unique magnetic field patterns, with different magnitudes and distributions, as shown in Fig. 4. These differences in the magnitude and distribution of the magnetic flux consequently affect the operational performance of the machine. In particular, the magnetic field distribution affects the back-EMF induced in the machine. Figure 5 shows the three-phase induced back-EMF in each generator, and also Fig. 6 provides a comparison of the phase back-EMF of the models. As shown in Fig. 6, for the significant part of the period, the back-EMF of the triangular magnet model is higher than the other three models" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000769_978-3-662-43893-0-Figure3.22-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000769_978-3-662-43893-0-Figure3.22-1.png", + "caption": "Fig. 3.22 Displacement of the center of the contact pattern on the tooth flanks of the wheel for different \u0394V and \u0394H values", + "texts": [ + " an axial shift of the pinion towards the center of the gearbox. This axial displacement is harder to absorb for a tapered roller bearing in O-configuration behind the pinion, and there is risk that the bevel gear will jam if backlash is insufficient. In general terms, the displacement behavior for the drive and coast modes may be simplified by stating that in the drive mode the offset is reduced and the mounting distance increased, while in the coast mode the offset is increased and the mounting distance reduced. Figure 3.22 shows, on the tooth flanks of the wheel, how the center of the contact pattern is displaced when varying values \u0394V and \u0394H. In each case, the arrows on the lines indicate the positive direction for the parameter change. It will be evident that on the drive side, i.e. the convex tooth flank on the wheel, with displacements \u0394V< 0 and\u0394H> 0 induced by the reactions on the teeth, the contact pattern moves towards heel and tip. On the coast side, with displacements\u0394V> 0 and \u0394H< 0, the contact pattern also moves towards heel, but with a tendency towards tooth root" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001833_tec.2020.3041658-Figure23-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001833_tec.2020.3041658-Figure23-1.png", + "caption": "Fig. 23. Rotor rated-load flux density distributions in CPM machines with different rotor PM thicknesses. (a) hPM = 5.5mm. (b) hPM = 6.0mm. (c) hPM = 6.5mm. (d) hPM = 7.0mm. (e) hPM = 7.5mm. (f) hPM = 8.0mm.", + "texts": [ + " Because the difference between the phase angles of the 5th radial and tangential air-gap flux density harmonic is more than 90 electrical degree, which contributes to the backward average torque by (13), as illustrated in Fig. 22(b). The torque proportion of the Zr order harmonic is shown in Fig. 22(c). It can be seen that torque proportion decreases with hPM and then tends to stabilize when hPM = 5.5 mm, which is mainly due to the strong FME in a higher hPM and serious magnetic saturation in the rotor teeth when hPM is larger than 5.5 mm, as shown in Fig. 23. As a consequence, the average torque contributed by the field modulated harmonics increases with hPM, and then tends to be stable, as shown in Fig. 22(d). In order to illustrate the CPM machine has stronger FME than its SPM counterpart. The torque characteristics of the SPM machine with different hPM is illustrated in Fig. 24. It can be seen that the torque increases first, and then slightly decreases with the increase of hPM, as shown in Fig. 24(a), which is mainly due to the fact that higher PM volume results in serious magnetic saturation" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001385_978-3-319-28872-7_12-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001385_978-3-319-28872-7_12-Figure3-1.png", + "caption": "Fig. 3 Automated scripts can quickly and easily apply complex modifications to the generated designs, such as the addition of lattice shaped speed holes on each solid wall of the quadrotor frame", + "texts": [ + " New designs can be formed through combinations of these underlying building blocks. Existing designs can be easily reconfigured by adjusting the composition parameters of the constituent modules. In addition to the modularity, another benefit of this method of design comes from its scripting abilities. An automated script was able to generate matching tabs and slots to attach the faces for each beam based on overall geometry. Another script was used to perforate each solid face with lattice-like speed holes, as shown in Fig. 3. Though such scripts take some time and expertise to develop initially, they can then be applied at will. In contrast to manual design tools that require repetitive placement of each feature, this process enables future designs to apply a short snippet of code to modify the entire design, regardless of complexity, as seen in Listing3. Listing 3 Scripting capabilities allow repetitive or complexmodifications to be easily encapsulated and shared 1 import quad 2 import la t t ice 3 4 q = quad.Quad( radius = 75) 5 la t t ice " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000002_j.matdes.2015.08.019-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000002_j.matdes.2015.08.019-Figure3-1.png", + "caption": "Fig. 3. Scheme of the ball flattened by a rigid target.", + "texts": [ + " The resulted collision and rebounding behavior of the balls largely depends on the initial impact velocity,while the different features are identified in relation with the following three ranges of the impact velocity. Fig. 2 shows the impact force (F) and the displacement (u) along the impact direction at different locations of the ping pong ball when the initial velocity of the ball is V0 = 0.5 m/s. Here the interaction force F0 = Kh and collision characteristic duration \u03c4 will be given below. Based on preliminary simulation results, the following assumptions can be reasonably introduced during the compression of the ball: (1) A cap portion of the ball, ACB, as schematically shown in Fig. 3, is flattened by the collision onto the rigid target and then rests on it; (2) The deformation of the ball is locally constrained around the impact end (i.e., within the cap ACB), which means that the remaining part of the ball, ADB, as shown in Fig. 3, remains undeformed and moves toward the rigid target with the initial velocity; (3) The relationship between the interaction force and the displacement due to flattening remains the same as that under a static compression. According to the solution developed by Reissner [38] for an elastic spherical shell loaded by a point force at the crown, the force F is F \u00bc 8D Rc \u03b4 \u00f02\u00de act force vs. compression ratio of the ball, when V0 = 0.5 m/s. where c = \u03b2h is the reduced thickness of the ball with \u03b2 = [12(1 \u2212 \u03c52) ]\u22121/2 which takes the Poisson's ratio of thematerial, \u03c5, into account; D = Ehc2 is the bending stiffness of the wall" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001236_j.optlastec.2019.105586-Figure5-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001236_j.optlastec.2019.105586-Figure5-1.png", + "caption": "Fig. 5. Effect of magnetic field on plasma: (a) the plasma plume without magnetic field, and (b) the plasma plume in a magnetic field.", + "texts": [ + " In a magnetic field, which is controlled by the Lorenz force, a charged particle will move in a circle around the magnetic introduction line. The gyrating radius can be calculated by = \u22a5rc mV q B| | [44], where v\u22a5 is the velocity perpendicular to B. Therefore, the motion of plasma will be regulated by the magnetic force. To prove this analysis, the plasma image during the DMD process was collected through a CCD camera. An optical filter was placed in front of it to protect the camera lens. When there was no magnetic field, the plasma plume was sharp and soaring upward (Fig. 5a). However, in the magnetic field (Fig. 5b), it became short and stout. The changed shape of the plasma caused a observable difference in the molten pool. When there was no magnetic field, the molten pool was small and confined to the laser irradiation area. This is because the affected area of plasma was restricted to a small region. As a result, the heat affected zone (plasma induced) was small, and the thermal input was low. Meanwhile, when the plasma was compressed by the magnetic force, the molten pool expanded to approximately 3 times larger than that of the former, which generated additional heat input" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000093_j.engfailanal.2013.02.003-Figure8-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000093_j.engfailanal.2013.02.003-Figure8-1.png", + "caption": "Fig. 8. Finite element model of nominal dimension.", + "texts": [ + " Six finite element models were erected to study the effects of structural parameters on stress and natural characteristics. According to Table 2, the parameters of six models were chosen as Table 3. Here we could study the influence rules of tooth back thickness by contrasting results of number 1, 2, 3, 4, and study the influence rules of tooth root fillet radius by contrasting results of number 2, 5, 6. Taking number 2 for an example, the finite element model was erected by solid45 in ANSYS containing 72,598 elements and 87,950 nodes as Fig. 8. Accurate prediction on the stress of gears related to contact mechanics which was a typical nonlinear analysis. However, linear static analysis would be enough and were simpler to study the influence rules of the two parameters. Engaging force acting on the gear tooth was the main load for the stress concentration near tooth root. While assuming engaging force as a uniformly distribution on the narrow zonal region along the pitch line, the force could be derived as follows: Ft \u00bc 2T=dm Fn \u00bc Ft= cos a Here Ft and Fn were the tangential force and normal force to be employed on finite element model" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000820_s11044-014-9445-4-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000820_s11044-014-9445-4-Figure2-1.png", + "caption": "Fig. 2 Active combined rimless wheel model", + "texts": [ + " (13), (16) and (18), the transition function, Q\u0304, is then solved as Q\u0304 = \u03b5 R\u0304 = cos\u03b1. (19) Therefore, we can find that the transition of the state error during the stance and collision phases are identical and are cos\u03b1. In the subsequent sections, we will investigate this result in more detail from the mechanical energy point of view. 3.1 Active combined rimless wheel model and linearization of motion As the realistic model of an active 1-DOF limit-cycle walker, we consider an active combined rimless wheel (CRW) shown in Fig. 2 [15]. This is composed of two identical eightlegged RWs of Fig. 1 and a body frame, and can exert a joint torque, u [N \u00b7 m], between the rear stance-leg and the body frame. We assume the following statements: \u2013 The fore and rear stance legs always contact with the ground without sliding. \u2013 The inertia moments about the CoMs of all the frames can be neglected. \u2013 The rear and fore RWs perfectly synchronize or rotate maintaining the relation \u03b8r \u2261 \u03b8f . The 3-DOF CRW with the hard ground configures a four-bar linkage, and exerting the joint torque, u, is thus equivalent to exerting that at the contact point with the ground (ankle-joint torque)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000520_s40684-016-0008-4-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000520_s40684-016-0008-4-Figure2-1.png", + "caption": "Fig. 2 Model of pitch reducer used in this study", + "texts": [ + " The power input through the sun gear was delivered to the carrier, which was an output part, after the load was shared among three planet gears. Therefore, it was possible to design a pitch reducer with a small size and light weight, and because the pitch line velocity, which influences both the noise and efficiency considerably, decreases with the gear size, it was advantageous with respect to noise and efficiency. The pitch reducer was modeled by considering the pitch bearing and all the elements of the reducer, including the gear, shaft, bearing, carrier, and housing, as shown in Fig. 2. For the gear, the parameters of the macro geometry, such as the number of teeth, module, face width, helix angle, pressure angle, center distance, and profile shift coefficient, were described, as well as those of the micro geometry, such as the profile modification and crowning. In addition, the nonlinear mesh stiffness was taken into account in the gear contact analysis. The internal bearing was modeled to have a nonlinear stiffness so that its deformation could be changed depending on the loading level" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000553_1.4034318-Figure5-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000553_1.4034318-Figure5-1.png", + "caption": "Fig. 5 (a) Angular momentum trajectories which are intersections between the energy ellipsoid and the angular momentum sphere for J1 C=J2 C 5 2:0 and J3 C=J2 C 5 0:5. For d 5 0.65, 0.8, 0.95, 1.05, 1.20, 1.35, 1.5, 1.65, 1.8, and 1.95, the projection of the trajectories: (b) onto the H2 C ; H3 C -plane, (c) onto the H1 C ; H3 C -plane, and (d) onto the plane H1 C ; H2 C -plane.", + "texts": [ + " Therefore, on the energy ellipsoid, the trajectory is closed and circulates about the H3C-axis. On the contrary, if d > 1, the vertices and focal points of the hyperbola are on the H1C-axis in the projected plane. Therefore, on the energy ellipsoid, the trajectory is closed and circulates about the H1C-axis. When d \u00bc 1, i.e., D \u00bc J2C, the trajectory of Eq. (40a) also becomes Eq. (40b). Therefore, Eq. (40b) becomes separatrices. Furthermore, H1C \u00bc H3C \u00bc 0 represents the state of permanent rotation: H2C \u00bc 1. Figure 5(a) shows the intersecting energy ellipsoid and the angular momentum sphere for J1C=J2C \u00bc 2 and J3C=J2C \u00bc 0:5. For d \u00bc 0.65, 0.8, 0.95, 1.05, 1.20, 1.35, 1.5, 1.65, 1.8, and 1.95, the projected trajectories onto the H2C; H3C-plane are shown in Fig. 5(b), those onto the H1C; H3C-plane are in Fig. 5(c), and those onto the H1C; H2C-plane are in Fig. 5(d). As Fig. 5(a) illustrates, the trajectories that form a pair of hyperbolas on the projected H1C; H3C-plane are closed on the energy ellipsoid, indicating that the rotation is periodic. We present the numerical simulations of Dzhanibekov\u2019s experiment of an unscrewing wing nut and the tennis racket experiment. For numerical solutions, we integrate Euler\u2019s equations (22a)\u2013(22c) using the fourth-order Runge\u2013Kutta method and the recovery formula for the rotation matrix (Eq. (24)), using the midpoint method (Eqs", + " The angular velocity components in the x3 - and x1-directions remain zero (blue lines). Figure 7(c) presents the angular momentum about the bodyattached coordinate axes. The magnitude of the angular momentum vector (black line total) retains the same value of 0:082 kg m2=s . The accuracy of the numerical solutions was checked both geometrically and analytically. To compare the momentum\u2019s trajectory with the geometrically exact trajectory, the angular momentum vector was plotted on the angular momentum sphere, as shown in Fig. 5(a). This geometrically exact trajectory is the intersection between the energy ellipsoid and the angular momentum sphere. For both simulations of the Dzhanibekov and the tennis racket experiments, the numerically computed trajectories of the angular momentum vector are indistinguishably close to the exact orbit. The accuracy of the numerical simulation for the tennis racket experiment was compared with the analytical solution [12,13] in Fig. 8. The relative maximum error was 0.19%. We defer the presentation of the analytical solution in a separate publication, where the sensitivity of the deviation of the initial conditions from the axis of the intermediate permanent rotation will be further investigated [23]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001744_tec.2020.2996817-Figure9-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001744_tec.2020.2996817-Figure9-1.png", + "caption": "Fig. 9. Temperature distribution of two operating modes at the rated working condition", + "texts": [ + " TABLE III THERMAL CONDUCTIVITY OF EACH PART OF THE PROTOTYPE Component Thermal conductivity W/(m\u2219K) Stator winding \ud835\udf06 = 400 Stator core \ud835\udf06\ud835\udc65 = 48, \ud835\udf06\ud835\udc66 = 48, \ud835\udf06\ud835\udc67 = 1.7 Slot wedge \ud835\udf06 = 0.8 Rotor core \ud835\udf06\ud835\udc65 = 48, \ud835\udf06\ud835\udc66 = 48, \ud835\udf06\ud835\udc67 = 1.7 Permanent magnet \ud835\udf06 = 7.5 End cover \ud835\udf06 = 96 Base \ud835\udf06 = 12 Shaft \ud835\udf06 = 50.2 According to the above analysis, the finite element method based on the steady-state temperature field is utilized to calculate solution region of the prototype, and the temperature distribution of the stator and PM are solved at the rated working condition, as shown in Fig. 9. According to the calculation data in Fig. 9, when the PMSM is switched from the healthy mode to the fault-tolerant mode, the temperature rise of the non-fault phase is the largest, increasing by 14.4\u2103. Although there is no current flowing through phase A at this time, its temperature has also increased by 8.1\u2103 due to the heat transfer of the other two-phase. The temperature rise of the PM is also obvious, increasing by 11.3\u2103. We can also notice that after switching to the fault-tolerant mode, although the stator core loss decreases by 5.5W, the temperature of the stator tooth top increases by 10" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001741_j.autcon.2020.103264-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001741_j.autcon.2020.103264-Figure1-1.png", + "caption": "Fig. 1. Configuration of blade movement by cylindrical movement.", + "texts": [ + " Unlike the previous studies, this research proposes a method for estimating the position of the blade cutting edge using a mathematical model and a Cartesian space control algorithm for controlling the height of the blade cutting edge according to the design surface with the VSD (Virtual Spring Damper) hypothesis. The study installed an automatic blade control system with the proposed algorithm in a real bulldozer to verify the performance of its leveling operation, and compared the results with the traditional leveling method. Bulldozer blade rotates on three separate axes as shown in Fig. 1 and has three separate rotation motions (lift motion, angle motion, and tilt motion) depending on the rotation of each axis. Lift motion is defined as the rotation of the blade around Y0 axis with the two lift cylinders contracting and expanding. Angle motion is defined as the rotation of the blade around Z1 axis as angle cylinders contract and expand. Lastly, tilt motion indicates the rotation of blades around the X1 axis as the tilt cylinder contracts and expands. Thus, bulldozer blade motions have a total of 3 degrees of freedom (DOF), and their posture is determined by the contraction and expansion of the lift, angle, and tilt cylinders" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001629_j.renene.2019.09.049-Figure5-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001629_j.renene.2019.09.049-Figure5-1.png", + "caption": "Fig. 5. 1/12 model of the LHPM generator.", + "texts": [ + " According to equations (10) and (11), the following expression can be derived as, \u00f01\u00fe sm\u00det bm \u00femrKcd hm \u00bc Br Bg (12) Equation (12) shows that the parameters of PM, i.e., its arc length and thickness, will influence the magnetic loading, which will be analyzed later. Since the number of poles of the LHPM generator should be typically larger, the fractional-slot concentrated (double-layer) winding is often adopted for the armature winding. 36 slots/24 poles structure of the LHPM generator with fractional-slot concentrated double-layer winding is adopted in this paper. Fig. 5 shows the typical structure of a 1/12 model of the LHPM generator. The cogging torque is caused by interaction between the PMs mounted on the rotor and the stator teeth [20,21]. It will cause noise and mechanical vibration. The influence of such torque will be more significant for the low-speed generator, especially as the starting torque. In order to reduce the cogging torque, this paper adopts the skew method for stator slot. The cogging torque with skew can be expressed as [22], Tc \u00bc X i\u00bc1;2;3::: KskTisin iZpq (13) where Ksk is the skew efficiency, Zp is the smallest common multiple between the slot number Z and the pole number 2p" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000088_j.mechrescom.2012.12.007-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000088_j.mechrescom.2012.12.007-Figure1-1.png", + "caption": "Fig. 1. Contact status between cycloid wheel and pin gears.", + "texts": [ + " In addition, an ideal profile modification curve for even load distribution can also be found by solving this equation. This method was proved as an effective way to improve the situation of heavy-duty tooth contact models. 2. Materials and methods 2.1. Single tooth meshing mathematical model Under certain assumptions, we can establish mathematical model of the single tooth meshing under considered contact area shape and deformation transfer effect of discrete contact p i l w p r n i u n t u e w i c p s i e oints. The single tooth meshing status under loads is shown n Fig. 1. As shown in Fig. 2, the pressure on single tooth meshing under oads can be shown as below: \u2212f (xi, yi) + \u0131 = k\u2211 j=1 (Aij + Bij)pj + Ei k\u2211 j=1 F(Pj) = W { Pj \u2265 0, (xj, yj) \u2208 \u02dd Pj = 0, (xj, yj) /\u2208 \u02dd \u23ab\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23ac \u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23ad (1) here F is unit center pressure, Pj is unit pressure function of indeendent variable, W is the normal total load of pin tooth, and it elates to the force of pendulum wheel, Aij is a flexibility compoent of contact deformation for unit j under unit i tooth profile, Bij s a flexibility component of pin teeth deflection for unit j under nit i tooth profile, Ei is processing and assembly errors of center ormal tooth for unit i, \u0131 is normal contact deformation between eeth, and f(xi, yi) is an imaginary shape at any point i of pin teeth nder-load without considering contact deformation" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002175_tie.2021.3084172-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002175_tie.2021.3084172-Figure1-1.png", + "caption": "Fig. 1. The five-phase 10-slot/8-pole PMSM. (a) Winding configuration of the investigated machine. (b) Winding configuration of the conventional machine. (c) Prototype. (d) Rotor.", + "texts": [ + " Current reconstruction method and the influence of winding resistance on current reconstruction are described in Sections III and IV. Experimental results are presented in Section V, and Section VI is conclusion. II. INVESTIGATED MACHINE AND SHORT-CIRCUIT FAULT ANALYSIS A. Investigated Machine The investigated machine is a five-phase 10-slot/8-pole PMSM with surface-mounted PMs and double-layer FSCW. The two coils of one phase are wound on adjacent teeth, which is different from the conventional scheme with two coils wound on the right opposite teeth, as shown in Fig. 1. The adopted winding scheme brings better magnetic isolation capability. Eccentric PMs are used to get both sinusoidal air-gap flux distribution and sinusoidal back EMF voltages. The five-phase windings are open-ended and driven by full-bridge inverter, which relieves constraint on zero-sequence current (ZSC) and enables the windings to be controlled independently. These characteristics bring more freedom for reconstructing post-fault currents and are beneficial for improving the post-fault torque performances [1]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001712_j.ymssp.2020.106823-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001712_j.ymssp.2020.106823-Figure3-1.png", + "caption": "Fig. 3. Time estimation for (a) lubricant contact; (b) solid contact.", + "texts": [ + " Because the stiffness of the solid contact is very high, there is no remarkable change of elastic deformation the engage tooth at two adjacent time-steps. Thus, a very small time-step is required for the solid contact, and the maximum time-step Dtsmax for the solid contact can be determined using the following formula Dtsmax \u00bc Dtlmax ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi Kav l =Kav s q \u00f024\u00de here, Kav l;s denote the average mesh stiffness for the lubricant and solid contact, respectively. The accuracy of the ONA mainly depends on the transition area (see Fig. 3 or Fig. 5), and the error criteria in this paper is given by x t\u00f0 \u00dej j Lj j < e \u00f025\u00de To guarantee the convergence of the numerical solution and to ensure correct results, the precision tolerant (e) is equal to 1E 15m. Although the transition area is very small, two motions for lubricant/solid contact still could occur. And the criteria for differentiating the two types of motion in the transition area are defined as sign x t\u00f0 \u00de\u00f0 \u00de\u2013sign _x t\u00f0 \u00de\u00f0 \u00de Lubricant contact sign x t\u00f0 \u00de\u00f0 \u00de \u00bc sign _x t\u00f0 \u00de\u00f0 \u00de Solid contact \u00f026\u00de It is worth noting that the computational efficiency and the solution accuracy are strongly dependent on the time-step size", + " Considering the balance between precision and computational efficiency, a variable time-step size should be used [29]. However, a problem may be encountered in calculation is that the numerical solution cannot give a convergent result with higher time-step size. For this purpose, an adaptive selection of the time-step method is used to improve the calculative efficiency for the lubricant and sold contact. After determining the current state of the motion (lubricant and solid contact), the transition time for two different motions is required. According to Fig. 3, the exact point for the state transition can be calculated using x t0 \u00fe Dtl;s \u00bc x t0 \u00fe Dtl\u00f0 \u00de \u00bc sign _x t0\u00f0 \u00de\u00f0 \u00deL Lubricant contact x t0 \u00fe Dts\u00f0 \u00de \u00bc sign _x t0\u00f0 \u00de\u00f0 \u00deL Solid contact \u00f027\u00de here, Dtl and Dts denote the time interval from lubricant to solid and vice-versa, respectively. Combining Eqs. (27) and (21) yields x t0 \u00fe Dtl;s \u00bc Cd t0;Dtl;s \u00f028\u00de Applying the power series expansion to the trigonometric function in the region near the point t0 by neglecting the trun- cation error, Eq. (28) can be arranged with Taylor\u2019s formula by neglecting the truncation error x t0 \u00fe Dtl;s x t0\u00f0 \u00de \u00fe _x t0\u00f0 \u00deDtl;s \u00fe Dt2l;s 2 Rbpx1 X1 i\u00bc1 j #i p1cos ix1t0 \u00feui\u00f0 \u00de #i p2sin ix1t0 \u00feui\u00f0 \u00de Dt2l;s 2 Rbg Xn j\u00bc1 cj _x t0\u00f0 \u00de \u00fe kjx t0\u00f0 \u00de \u00fewj Tg t0\u00f0 \u00de " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000847_1.3657269-Figure8-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000847_1.3657269-Figure8-1.png", + "caption": "Fig. 8 Phase-plane representation of transient in system", + "texts": [ + " However, since /dEg\\ /dEe\\ \\ dt )o=o, \\ dt Je=Bi there is less of a tendency to chatter at 8 = 06 than there is at 8 = Oi,. This is true even though dE, dt > dE, dt 1 = 9, because of the magnitude of the quantities involved. In the co-ordinate system used, 8 varies between \u2014180 and +180 deg. Hence, in a phase-plane representation of a transient it is convenient to confine the plot of a trajectory to \u2014180\u00b0 < 6 < + 180\u00b0. The shape of the trajectory obtained for the transient considered in Fig. 7 is shown in Fig. 8. This trajectory indicates that a considerable amount of damping can be achieved by the technique discussed. It can be seen that it is desirable to have 05 as close to \u2014180 deg as possible. The control system that has been analyzed departs from an optimum system in several respects. With optimum control of an ideal system the power element is driving the system toward equilibrium in an optimum manner at all times when this state has not been obtained. In the control system using derived rate this is not true. For example, in Fig. 8 the system is being driven away from equilibrium when \u2014180\u00b0 < 8 < 06 and no driving forces occur in the region 06 < 0 < 06. In other papers,2 the system is considered to be noise free. Since noise is present in any control system, it is necessary to account for it in the design. For the control system discussed, both dead band and hysteresis were used to minimize the effects of noise. There are disturbances which give rise to chattering in the derived-rate system. In an ideal system with optimum control this does not happen" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002432_s00170-021-07326-6-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002432_s00170-021-07326-6-Figure2-1.png", + "caption": "Fig. 2 Ti-6Al-4V rod test pieces as-manufactured and geometry of the machined tensile test samples", + "texts": [ + " In this study, the porosity of the parts was characterised on the internal and external structures by comparing Archimedes\u2019 method, metallographic analysis, and computed micro-tomography (micro-CT) methods. The Ti-6Al-4V aerospace grade 5 powder was used as feedstock to produce the test samples. The morphology of the gasatomized powder (TLS Technik GmbH, Germany) is spherical as shown in Fig. 1. The powder size distribution (D10 = 25.05 \u03bcm, D50 = 38.31 \u03bcm, and D90 = 56.69 \u03bcm) is measured by a laser diffraction sizer (Microtrac, S3500, Germany). The cylindrical specimen in Fig. 2 was manufactured with 100 mm height and 10 mm diameter in triplicate using the Aeroswift machine (CSIR, South Africa) with 5000 W fibre laser (maximum capacity) in an argon atmosphere. The preheating temperature of the built platform was 200\u00b0C to minimise the residual stresses with successive powder layer thickness of 50 \u03bcm applied during the build. The process parameters used in the current study were screened for high densities above 99% densification ratio in the as build condition by varying heat input at 51 J/mm3, 72 J/mm3, 77 J/mm3, and 79 J/mm3", + " The thermal conditions such as temperature gradient during heating and local cooling rate (\u2202T/\u2202t) during solidification were computed using the COMSOL software package for a time-dependent, 2D model. The thermal model consists of a substrate of 2mm in height with the powder layer of 50 \u03bcm thick on top of the substrate with melt pool length of 3 mm. The substrates act as the heat sink during fabrication process. The heat transfer module modelled was computed from the parameters in Table 1. The parametric sweep function was applied to compare changes in time-dependent laser intensity. The cylindrical rod test pieces on the base plate (Fig. 2) were stress-relieved at 750\u00b0C for 2 h in a vacuum furnace under an argon environment [10]. One batch was further taken for hot isostatic pressing (HIP) treatment at 900\u00b0C, 1000 bar isostatic pressure for 2 h, and then cooled under argon atmosphere. The samples were removed from the base plate then machined into a dog-bone profile. Both the cubes and tensile specimens were machined to remove about 500 \u03bcm of the rough surface prior to measuring material porosity and the mechanical tests. The rods for the tensile specimen were mechanically machined to the specification presented in Fig. 2 according to ASTM E804 conditions. The Ohaus densitometer is an automated instrument with a density resolution of 0.0001 g/cm3 that uses the Archimedes principles to determine the density of the AM samples using ethanol as a liquid medium. The porosity of the samples was calculated using Eq. 3 from the bulk density of 4.43 g/cm3 for SLM produced Ti-6Al-4V: %Porosity \u00bc Bulk Density\u2013Measured Density\u00f0 \u00de=Bulk Density*100 \u00f03\u00de Tensile tests were performed using Zwick/Roell 50-kN tensile tester with an extensometer connected to the sample neck area diameter (D) of 6\u00b10" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001703_j.jmapro.2020.02.035-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001703_j.jmapro.2020.02.035-Figure1-1.png", + "caption": "Fig. 1. Coordinate systems applied for pinion generation.", + "texts": [ + " It should be noticed that the tooth surface is very complicated that the uniform tooth flank parametric expression cannot be obtained in closed form. Actually, mathematical tooth flank modeling is a simulation process of the actual generation satisfying the theory of gearing [27]. Therefore, the equation expression of tooth surface can be obtained by simulating the cutter blade cutting process in form of the coordinate transformation from the tool coordinate system to the gear coordinate system [1]. Fig. 1 shows the coordinate system setup applying for generation of the pinion and Gleason\u2019s cradle-style machine. Coordinate systems Sm1, Sa1, Sb1 are fixed and rigidly connected to the hypoid generator. The movable coordinate systems S1 and Sc1 are rigidly connected to the pinion and the cradle, respectively. They are rotated about zb1 axis and zm1 axis, respectively, and their rotations are related with \u03c61. Coordinate systems Sp is applied for illustrating installment of the headcutter on the cradle in generation of the pinion [1\u20136]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001847_tro.2020.3038687-Figure17-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001847_tro.2020.3038687-Figure17-1.png", + "caption": "Fig. 17. Position accuracy experiment.", + "texts": [ + " Tendons from motor 2 were also longer, 391 and 377 mm for the externally actuated type compared with 187 and 173 mm for the hybrid-actuated type, respectively. The motor-joint coaxial couple structure led to a compact design with shorter tendons and reduced self-weight. However, the center of gravity was shifted toward the distal end, which increased the load on the proximal shoulder joints. We further compared the position accuracy of the two elbow joints. An aluminum pipe was connected to the joints (see Fig. 17). Three levels of weight (0 g, 300 g, and 700 g) fixed on the pipe 720 mm from the rotational axis were used as loads. Joint 2 was rotated to lift the load from 0\u00b0 to 90\u00b0. An electromagnetic motion tracking system (Liberty, Polhemus, USA) was used to measure the position of the load. Fig. 17 shows the two sensors and the origin, which is the source of the electromagnetic dipole field of the Liberty system. The experimental results in Fig. 18 show that the position accuracy of both types declined with the increase in the load, which is a characteristic of the tendon drive. However, the decline in the accuracy of the externally actuated type was significantly greater than that of the hybrid-actuated type. The errors in the joint angle of the externally actuated type were 3.8\u00b0, 5.6\u00b0, and 9" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002026_tec.2020.3035258-Figure32-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002026_tec.2020.3035258-Figure32-1.png", + "caption": "Fig. 32. Prototype of the PM-bottom HESFM with iron bridge.", + "texts": [ + " In order to balance the negative value of the induced voltage at a higher speed, only a full H-bridge converter can be implemented, since the half H-bridge converter cannot provide a reverse output voltage. Consequently, the induced voltage ripple in field windings significantly influences the control system and operating characteristics of the HEMs and it also increases the KVA ratings and thus the cost of the converter feeding it, especially at a higher speed. In order to validate the FEM predictions in previous sections, a prototype PM-bottom HESFM having 12-statorslots, 10-rotor-poles, and iron bridges is fabricated as shown in Fig. 32. The basic design parameters of the prototype HESFM are listed in Table V. The test rig illustrated in Fig. 33 is adopted for the experimental validation. It should be noted that the aim of this section is to validate the phenomenon of the no-load and on-load FW induced voltage ripples and the accuracy of FEM predictions. Therefore, the number of turns per FC is reduced to 10 in the prototype machine to reduce the difficulty of the winding process while the number of turns per phase of AW is kept the same as the FEM analysis in previous sections" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000769_978-3-662-43893-0-Figure1.8-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000769_978-3-662-43893-0-Figure1.8-1.png", + "caption": "Fig. 1.8 Bow thruster unit", + "texts": [ + " A desire for improved maneuverability first led to the development of the bow thruster where a tubular opening containing a propeller passes laterally through the entire width of the bow section of the ship, below the waterline. The propeller can move the bow to port or starboard either when the ship is stopped or moving ahead or astern at very slow speed. This is done by 1.4 Marine Drives 7 reversing the direction of rotation of the propeller or altering the pitch of the propeller blades. The propeller is powered by an electric or hydraulic motor installed in the ship. The principle of the bow thruster is illustrated in Fig. 1.8. The motor installed inside the ship drives the propeller shaft through bevel gears. An alternative concept uses a direct drive electric motor for the bow thruster. Although the simplicity of this concept is attractive, the technical problems in sealing the direct drive sustainably under water are not easily solved. The classic bevel gear drive has advantages in terms of reliability and fail-safe properties since, even if the gearbox seal fails, a bow thruster with sufficient oil pressure in the gearbox will continue to operate for several weeks whereas a direct drive will sustain immediate seawater damage" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000867_0954406214536547-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000867_0954406214536547-Figure2-1.png", + "caption": "Figure 2. Bearing 2 and 3 at sun and gear1.", + "texts": [ + " The contact force between mating gears can be obtained by multiplying the time varying gear mesh stiffness with dynamic transmission error Fpg\u00f0t\u00de \u00bc Kpg\u00f0t\u00de pg\u00f0t\u00de \u00f010\u00de where Fpg(t), Kpg(t) and pg(t) are the contact force, time varying gear mesh stiffness and dynamic transmission error of gear and pinion, respectively. In present analysis, the stiffness of the rolling element bearings that exist in the gearbox is assumed to be constant and linear. The nonlinear effect of the bearing elements is not considered. Figure 2 shows the sun and gear1 placed on two bearings. The stiffness values at SYRACUSE UNIV LIBRARY on June 9, 2014pic.sagepub.comDownloaded from of the two bearings considered in the present study are assumed to be same based on the literature data of a wind turbine. The sun and gear1 are placed in a sufficiently long shaft. Since length of the shaft is large, the lateral displacements of bearings 2 and 3 along y and z directions are different from lateral displacements of gears. The relation between lateral displacements of sun, gear1, bearing2, bearing3 can be formulated as2 yb2 yb3 \u00bc c11 c12 c21 c22 ys y1 ; zb2 zb3 \u00bc c11 c12 c21 c22 zs z1 \u00f011\u00de where c11 \u00bc l2=\u00f0l1 \u00fe l2\u00de c12 \u00bc l1=\u00f0l1 \u00fe l2\u00de c21 \u00bc l=\u00f0l1 \u00fe l2\u00de c22 \u00bc l3=\u00f0l1 \u00fe l2\u00de yb2, yb3, ys, y1 and zb2, zb3, zs, z1 are the lateral displacements of bearing2, bearing3, sun gear, gear1 in y and z directions, respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001573_j.mechmachtheory.2018.12.031-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001573_j.mechmachtheory.2018.12.031-Figure2-1.png", + "caption": "Fig. 2. End joint reaction of RPS limb.", + "texts": [ + " Then, M e axis, j can be expressed as the function on intensities f r i, 1 , f r i, 2 , . . . , f r i,n \u2212g\u22121 , namely, M e axis, j = F unction ( f r i, 1 , f r i, 2 , \u00b7 \u00b7 \u00b7 , f r i,n \u2212g\u22121 ) (6) Combing Eqs. (5) with (6) , intensities f r i, 1 , f c i, 2 , . . . , f r i,n \u2212g\u22121 of complement wrenches can be solved from the n - g -1 equi- librium equations. So, only intensity f a i of the actuation wrench and intensities f c i, 1 , f c i, 2 , . . . , f c i, 6 \u2212n of constraint wrenches are unknowns in Eq. (4) For example, for the RPS limb shown in Fig. 2 , in which R denotes a revolute pair, P denotes a prismatic pair and S denotes a spherical pair, a fixed frame is attached to the geometrical center of the R pair, with X -axis along the axis of the R pair, Z -axis perpendicular to the base plane and Y -axis determined by the right hand rule. Base on the screw theory, the unit constraint wrench $ c 1 , 1 is a force parallel to X-axis and through the geometrical center A of the S pair. If the P pair is chosen as the actuation pair, the unit actuation wrench $ a 1 is a force along the line determined by points O and A " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001846_tte.2020.3041194-Figure10-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001846_tte.2020.3041194-Figure10-1.png", + "caption": "Fig. 10. Distribution of the von Mises stress and the displacement of the rotor. (a) Mesh. (b) Stress. (c) Deformation.", + "texts": [ + " Thus, the optimizations of high sensitivity parameters and low sensitivity parameters are independent. Fig. 8 shows the comparison of torque waveform before and after optimization. It can be concluded that the torque performance is improved greatly after multi-objective optimization. Fig. 9 shows the influence of rib hb on average torque and torque pulsation in the optimization process. It can be seen that it is a good solution when the hb is chosen as 0.5 mm. After optimization, the distribution of the von Mises stress and the displacement of the rotor at 1500 r/min is obtained and shown in Fig. 10. It can be seen that the maximum stress is below 2 Mpa at the rib. The value is far less than the yield point of the iron material. Meanwhile, the PMs are affixed by the glue typed LOCTITE AA-326 on the rotor, and then the gap of 0.5 mm can be maintained. The shear strength is over 15.2 Mpa of the AA-326 which is acceptable to ensure the safety of PM at 1500 r/min. Moreover, the maximum deformations of the rib at 1500 r/min is only 5.32\u00d710-5mm. Fig. 11 shows the 1st to 5th models of the rotor, and the Campbell diagram" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000576_gt2016-57454-Figure7-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000576_gt2016-57454-Figure7-1.png", + "caption": "FIGURE 7. DM method pressure coefficient contour on symmetry of the single gear configuration at 5000 rpm", + "texts": [ + " In account of the obtained results, the DM method was exploited in order to assess its capability at grabbing the physical phenomena causing windage losses [14, 15]. Hence, an in-depth flow field investigation was carried out for 5000 rpm rotational 6 Copyright \u00a9 2016 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/conferences/asmep/89496/ on 05/29/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use speed condition, exploiting both pressure and velocity distribution along the gear and symmetry plane. In Figure 7 the pressure field distribution on symmetry surface and gear teeth is showed in terms of a pressure coefficient defined as follow: Cp = p\u2212 pREF pREF (1) Despite there is not an experimental evidence for the analyzed case, Figure 7 shows that the Dynamic Mesh method allows to detect a pressure distribution along the gear teeth similar to the one showed by the work of Hill et al. [15]. Therefore, it is clearly visible the pronounced pressure peak on the \u201dpressure side\u201d of the gear tooth due to the lubricant entering the space between two adjacent teeth as well as the lower pressure on the root diameter. The aforementioned DM model capability in windage losses phenomena detection is pointed out also by the velocity contour and vector plot showed in Figure 8", + " Here a 3D visualization of the flow field is accompanied with four 2D planes: the YZ plane (highlighted in green) at X corresponding to the gear axis, and three XY planes (highlighted in red), normal to gear axis, distant from the symmetry 0.9, 0.5 and 0.1 b/2 respectively, allowing to assess the flow field distribution inside the gear teeth space. Also Figure 8 clearly shows the characteristic flow field distribution of a single gear windage configuration. Indeed, by the plot of y and z velocity component on the YZ plane, it is possible to note the fluid mass flow driven into the gear space by the low pressure region located in the tooth valley (see Figure 7) which cause the high peak on the tooth \u201dpressure side\u201d. Moreover, exploiting the relative velocity plot on the 3 XY planes, it is possible to identify the recirculating flux into the relative frame of the gear space, as detected in other previous works [14, 15]. In the region of the teeth spacing near the wheel side, where the lubricant is entering in, its rotating velocity is lower than the gear one and the oil is impinging the tooth on the \u201dpressure side\u201d. Here static pressure increases, forming the peak shown in Figure 7, since the relative kinetic pressure contribution decreases as the gear increases the oil rotational speed. Due to the momentum transfer between tooth and fluid in such region, fluid is accelerated both in circumferential and axial direction as highlighted by the velocity plot on the plane YX of Figure 8. Moving forward into the teeth channel, the pressure gradient which arises between the two sides circumferentially accelerates the lubricant. Therefore oil rotates faster than the gear causing the above mentioned recirculating flux into the gear space, as already discussed by Hill et al", + " In the present configuration, due to the nonsymmetric geometry of the test-rig, the left part of the gear (for x > xGEAR\u2212AXIS) \u201dsees\u201d a more confined domain while the right one \u201dsees\u201d a greater region in which oil has a stronger momentum dissipation. Following the described behavior, the oil in the left side of the gearbox rotates faster than the one in the right side; consequently the relative velocity and the resulting pressure increment are lower in the left side of the gearbox rather than in right one, as confirmed in Figure 7. Fluid-dynamic losses of the gear pair Following the structure of the previous section, Figure 9 shows a comparison of the fluid-dynamic resistant torque ob- 7 Copyright \u00a9 2016 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/conferences/asmep/89496/ on 05/29/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use tained with the experiments carried out by Gorla et al. [2] and numerical results versus the rotational speed for the gear pair configuration" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000785_j.foodchem.2015.05.059-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000785_j.foodchem.2015.05.059-Figure1-1.png", + "caption": "Fig. 1. Schematic illustrating the fabrication process of a graphene\u2013CoMS hybrid paste electrode. (A) Inserting a piece of copper wire ((b) 10 cm long, 150 lm diameter) into a 3.5 cm long fused-silica capillary ((a) 320 lm I.D. 450 lm O.D.); (B) applying hot melt adhesive (c) to glue (b) in place, (C) filling the empty end of (a) with a mixture of graphene\u2013CoMS hybrid and paraffin oil (3:1, w/w) (d), and (D) Microscopic photograph of a piece of graphene\u2013CoMS hybrid paste electrode.", + "texts": [ + " After 60 min, the obtained graphene\u2013CoMS hybrid was isolated by vacuum filtration and was purified by washing with copious amounts of doubly distilled water. Graphite\u2013CoMS hybrid was also prepared for comparison purposes. The preparation procedures and conditions were the same as those of graphene\u2013CoMS hybrid except 0.15 g graphite powder was used instead of 0.3 g oxidized graphite powder. In addition, graphene was prepared by the reduction of GO sheets with hydrazine hydrate as a reducing agent (Tang et al., 2013). The fabrication process of a graphene\u2013CoMS hybrid paste electrode is shown in Fig. 1A\u2013C. Graphene\u2013CoMS hybrid powder was mixed with paraffin oil at a ratio of 3:1 (w/w). And then, a piece of copper wire (15 cm long and 150 lm diameter) was inserted into a fused silica capillary (320 lm I.D., 450 lm O.D., and 4.5 cm long) and a 3 mm long opening was left for accommodating the paste. A drop of hot melt adhesive was applied to the other end of the capillary to immobilize the copper wire. Subsequently, the empty end of the capillary was inserted into the paste on a piece of glass plate. Note that the empty end of the capillary should touch the glass plate so that the paste could be pressed into the capillary until it touched the end of the copper wire in the capillary. Prior to use, the paste electrode was smoothed on a piece of weighing paper. Fig. 1D shows the photography of a typical CoMS hybrid paste electrode. The end of the copper wire is well buried in the packed hybrid paste. In addition, graphene paste electrode and graphite\u2013CoMS hybrid paste electrode were also fabricated. The CE-AD system used in this work has been previously reported (Tang et al., 2013). A high-voltage DC power supply (30 kV, Shanghai Institute of Nuclear Research, China) was employed to apply a separation voltage between the inlet and outlet of the capillary. The outlet of the capillary was maintained at ground" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001101_j.surfcoat.2018.01.071-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001101_j.surfcoat.2018.01.071-Figure3-1.png", + "caption": "Fig. 3. The final prepared sample and three-point bending test schemes.", + "texts": [ + " 2, the sample preparation experiment was carried out by using a direct laser depositing (DLD) system which mainly consists of a 5 kW transverse-flow continuous wave CO2 laser, four axis numerical control (NC) workbench, continuous powder feeder, chiller (water-cooling unit) and coaxial powder feeding nozzle. The multi-pass deposition path was one-way orthogonal type (laser scanning pattern in Fig. 2), and the main process parameters were selected according to the factory's know-how and were listed as follows: laser power P= 1.7 kW, spot diameter d= 4mm, scanning speed V=300mm/min, powder feeding rate S= 4.2 g/min, argon shielding gas flow 10 L/min, hatch spacing H=2mm. The prepared sample with a deposited layer was shown in Fig. 3(a), the deposited thickness was more than 1mm and its size was 120mm\u00d780mm. By using an electric spark machine, several metallographic specimens were cut from the as-deposited sample and prepared according to the standard method (ASTM E3-11). After that, the 4% nitric acid alcohol solution and aqua regia were used to etch the base metal material (H13) and the deposited layer (Ni-based superalloy), respectively. Then the microstructure and composition of asdeposited specimens were examined by using an Axiovert 200 MATZESIS optical microscope and a Hitachi S3400 Cold Field Scanning electron microscope (CFESEM), which was equipped with a KevexSingma Level-4 energy dispersive X-ray spectrometry (EDS)", + " The constituent phases of the as-deposited layer were identified by X-ray diffraction (XRD) analysis, which was performed using a Phillips Rayons-X with Cu K-\u03b1 (\u03bb=1.5418 \u00c5) radiation source. Moreover, under a microhardness tester (FM-ARS 900 Vickers), the microhardness variation across the section of specimens was measured on polished surface, at a load of 500 gf (4.9 N) and a hold-time of 15 s, reference to ASTM E384. In order to determine the ability of the system (Ni-based superalloy/ H13 steel) to undergo plastic flexural deformation. As depicted in Fig. 3(a), three groups (G1, G2, G3) of specimens prepared for the three-point bending test were machined from the as-deposited sample. Besides, one control group (H13 specimens) was also prepared. In each group, three specimens (Fig. 3(b)) were ground and sprayed with white paint to reduce the interference of reflected light to video camera. The arrangement for experiments is depicted schematically in Fig. 3(c). The dimensions (h\u00d7w\u00d7L) of specimens were 5mm\u00d75mm\u00d7100mm (the deposited layer thickness T1 was 0.8mm after grinding). The loading pin and supporting pins had diameters Ds (10mm). And the bending span (Ls) of two supporting pins is 80mm. As shown in Fig. 3(d), the three types of specimens (numbered as Ni- 1, Ni-2, Ni-3) corresponding to the specimen group (G1, G2, G3), which represented three orthogonal directions considering the different orientation of the deposited layer. The specimens were tested in three configurations to place the deposited layer in tension, compression, tension-compression, respectively. The three-point bending test was performed at a constant loading rate of 1mm/min by using MTS universal electro-hydraulic servo tester, which was equipped with a realtime image data acquisition system to capture deformation process" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001761_012028-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001761_012028-Figure2-1.png", + "caption": "Figure 2. Representative cross-sections of the machine (slots in iron core are intentionally oriented as shown to assess effect of teeth with 2D FEA modelling).", + "texts": [ + " The motor housing is maintained in a partial vacuum for reducing windage friction drag at high rotational speeds. Frictionless non-contact magnetic bearings are required since mechanical bearings would not be able to operate continuously due to frictional losses and wear. Further configurational details are available in reference [16]. Initial sizing is conducted using 2D finite-element analysis (FEA) code. In this approximation the rotor poles are not offset by 90 elect.-degrees as the behavior of interest is the capability of the machine to effectively route flux within the rotor and stator. Figure 2 shows a 2D FEA model cross-section with armature coils housed between iron stator teeth. Note that the slots in the iron core (as shown in the figure) are for quantifying the effect of the iron teeth. In a real machine, the slots and teeth run parallel to the rotational axis of the machine. A double-layer, 3-phase AC winding configuration is selected. The dimensions of this cross-section suit the specifications of Table 1. The location of the field excitation coil is also shown in the figure. The 2D model of Figure 2 was calibrated with a 3D Opera FEA model shown in Figure 3. Empirical corrections were applied to the 2D model to match the results of 3D model. The 2D model is preferred because of its simplicity and ease of rapidly comparing different designs. The 3D calculated field experienced by the full-pitched stator windings is decomposed into fundamental and harmonics that are listed in Table 2. Most harmonics of concern (5th, 7th, 11th and 13th) are quite small and are not expected to be problematic in creating excessive eddy-current heating in the stator coil" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000494_s11665-015-1572-4-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000494_s11665-015-1572-4-Figure2-1.png", + "caption": "Fig. 2 Temperature Contours for symmetric step substrate shape (Color figure online)", + "texts": [ + " As the beam moves to a new location after a time t + dt, the heat flux correspondingly moved to the location, Lt + dt, with vdt being the length of deposited material over the incremental time dt. A single substrate with two level of thickness was investigated first. The idea was to place a single track across the entire substrate and get two distinct cooling rates with approximately an order of magnitude difference. The dimensions of the step substrate and the transient temperature distribution contour plot for the first deposited layer are illustrated in Fig. 2. The cooling rate along a deposited laser track is shown in Fig. 3. The average of nine measurements of the cooling rate along the laser track was 1700 and 2200 C/s in the thin and thick substrate section, respectively. This difference of the cooling rate is not deemed significant enough to make sufficient change on the microstructure and thus reveal the mechanical properties differences. Therefore, this step substrate shape was modified to create a larger effect. The second idea was to use two different substrates of differing sizes while depositing sequentially using the exact same parameters" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001710_s11071-020-05591-w-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001710_s11071-020-05591-w-Figure2-1.png", + "caption": "Fig. 2 Contact force of clearance joint", + "texts": [ + " Influences of optimization on wear characteristics, dynamics responses and nonlinear characteristics of mechanism considering clearance wear are also analyzed. 2.1 Modeling of revolute clearance Diagram of revolute clearance pair is displayed in Fig. 1. Due to existence of clearance, there are three relative motion states between shaft and bearing, namely free flight motion, impact motion and continuous contact motion. Fn and Ft represent normal contact force and tangential contact force, respectively, as displayed in Fig. 2. Relation between \u03b4 and F is as follows: F = Fn + Ft, \u23a7 \u23a8 \u23a9 \u03b4 < 0, F = 0 free flight mode \u03b4 = 0, F = 0 continuous contact mode \u03b4 > 0, F = 0 impact mode (1) where \u03b4 represents the relative penetration depth. \u03b4 = e \u2212 c, (2) where e is the clearance vector value, c is the clearance value, and c = R1 \u2212 R2. When \u03b4 < 0, there is no collision between shaft and bearing and they are in free flight mode. When \u03b4 > 0, shaft and bearing collide with each other and they are in impact mode. When \u03b4 = 0, shaft and bearing are in continuous contact mode" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000943_1350650115611155-Figure5-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000943_1350650115611155-Figure5-1.png", + "caption": "Figure 5. Angular position of balls.", + "texts": [ + " Then the resultant moment along with y-axis is X Miy \u00bc Xn i\u00bc1 Rp \u00f0Fi F\u00de sin 2 n \u00f0i 1\u00de \u00f03\u00de the resultant moment along with z-axis is X Miz \u00bc Xn i\u00bc1 Rp \u00f0Fi F\u00de cos 2 n \u00f0i 1\u00de \u00f04\u00de The 5-DOF quasi-static model When the ball bearing is working under the combined loads of an axial force Fx,, two radial forces Fy, Fz and two momentsMx,Mz, there will be a relative displacement between the outer ring and inner ring, which can be expressed as the axial displacement x, radial displacement y, z and angular displacement y, z for the inner ring (see Figure 4). And the angular position of each ball in bearing is shown in Figure 5. Assuming that the number of balls in bearing is Z and the contact force acting on ball in position angle j is Q j, then the force balance equations of inner ring are18 Fx \u00bc Xj\u00bcZ j\u00bc1 Q j sin \u00f05\u00de Fy \u00bc Xj\u00bcZ j\u00bc1 Q j cos j cos \u00f06\u00de Fz \u00bc Xj\u00bcZ j\u00bc0 Q j sin j cos \u00f07\u00de My \u00bc Xj\u00bcZ j\u00bc1 di 2 Q j sin j cos \u00f08\u00de Mz \u00bc Xj\u00bcZ j\u00bc1 di 2 Q j cos j cos \u00f09\u00de where di \u00bc dm D 2 cos In Figure 6, for a single ball with an azimuth of j under an applied static load, the action line between the inner and outer raceway groove curvature centers is OiOe" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001268_s40516-019-00099-7-Figure5-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001268_s40516-019-00099-7-Figure5-1.png", + "caption": "Fig. 5 Response surface contour plot for predicting sintering depth", + "texts": [ + "0001 Significant A-Laser Power 5.677326722 1 5.677326722 539.4902623 < 0.0001 B-Scan Speed 5.0976245 1 5.0976245 484.4038247 < 0.0001 C-Porosity 0.050033389 1 0.050033389 4.754442965 0.0518 D-Laser spot size. 3.590306722 1 3.590306722 341.1703448 < 0.0001 E-Build layer thickness 0.450933389 1 0.450933389 42.85012721 < 0.0001 Residual 0.115758519 11 0.010523502 Lack of Fit 0.115758519 6 0.019293087 Pure Error 0 5 0 Cor Total 17.24436472 31 Response Surface Contour Plot of Sintering Depth at Different Process Parameters Figure 5a shows the response surface contour plot for the prediction of sintering depth with respect to laser power and scan speed. From the plot, it is observed that, with an increase in laser power (70 W to 190 W), the sintering depth increases by keeping all other parameters constant. With an increase in laser power, more amount of heat is supplied to the powder bed at a particular instant of time. So that it will increase the sintering depth, as the laser power is directly proportional to heat flux. Similarly, the reverse phenomena are observed with varying scan speed. With an increase in scan speed, the interaction time between the powder bed and the laser beam decreases results in less amount of heat supplied at that particular time. So the sintering depth decreases. The variation of sintering depth with respect to laser power and the percentage of porosity present in the powder bed is shown in Fig. 5b. It is found that, as laser power increases from 70 W to 190 W, it leads to increase the sintering depth from 1.27 mm to 2.57 mm approximately by keeping constant porosity of the powder bed i.e. 5%. When the porosity percentage decreases from 25% to 5% with constant laser power, the sintering depth varies from 2.52 mm to 2.55 mm. It is clearly seen that change is porosity does not significantly affect the sintering depth. There is a slight change in sintering depth because of the presence of porosity which decreases the conductivity and density of the powder bed. Figure 5c shows that the sintering depth decreases from 3.24 mm to 2.45 mm as laser spot size increases from 0.2 mm to 0.6 mm at a constant laser power 190watt. The temperature of the top surface of the powder bed depends on the contact area of the laser spot size with powder bed. As the laser beam striking area increases, the energy concentration decreases, and so the surface temperature of the powder bed decreases. Hence sintering depth decreases and also it is found that the powder bed temperature inversely proportional to the laser spot size vice versa. By keeping the laser power constant, the sintering depth of the powder layer decreases from 2.63 mm to 2.43 mm with an increase in powder layer thickness from 1 mm to 3 mm as shown in Fig. 5d. From the contour plot, it is observed that as powder layer thickness increases more amount of heat loss takes place through the powder bed due to the different heat transfer mode such as conduction and convection. So energy intensity on the top surface of the powder bed decreases and causes a decrease in temperature field on the top surface, which results in a decrease in sintering depth. Optimization Analysis During the optimization of process parameters, an objective function was set to maximize the sintering depth" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000894_tro.2015.2489499-Figure13-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000894_tro.2015.2489499-Figure13-1.png", + "caption": "Fig. 13. Adept Quattro and its hidden robot model. (a) Adept Quattro: a 4\u2013 R \u2212 2 \u2212 US robot. (b) Its hidden robot model: a 2\u2013\u03a0 \u2212 2 \u2212 UU robot (when two legs are observed).", + "texts": [ + " Robots belonging to this category are probably the most numerous. They are those for which the hidden robot models have several possible assembly modes, whatever is the number of observed leg directions. Presenting an exhaustive list of robots of this category is totally impossible because it requires the analysis of the assembly modes of all hidden robot models for each robot architecture. However, some examples can be provided. Examples of such types of robots [the G\u2013S platform (see Fig. 12) and the Adept Quattro (see Fig. 13)] have been presented in [19], [21], and [22]. More specifically, in [21] and [22], it was shown (numerically but also experimentally) that the Adept Quattro [41] controlled through leg direction observation has always at least two assembly modes of the hidden robot model, whatever the number of observed legs. As a result, some areas of the robot workspace were never reachable from the initial configuration. Fig. 14 shows a desired robot configuration that was impossible to reach even if all robot legs were observed" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001644_rnc.4758-Figure5-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001644_rnc.4758-Figure5-1.png", + "caption": "FIGURE 5 The front view and top view of segmenting of parafoil [Colour figure can be viewed at wileyonlinelibrary.com]", + "texts": [ + " In Equation (20), Ar is obtained by Ar = [ msI3 03\u00d73 03\u00d73 Js ] , (21) where ms denotes the mass of the parafoil; Js denotes the matrix of rotational inertia; 03\u00d73 and I3 denote the third-order identity matrix and zero matrix, respectively. In Equation (18), the control quantity of the horizontal controller, so-called flap deflection, will lead the complex changes on F aero s . In the calculation of the aerodynamic force, the parafoil is divided into the flap and the canopy. Moreover, the canopy has eight parts geometrically along the span-wise direction, as shown in Figure 5. The aerodynamic force of these eight parts will be calculated, respectively, according to the local velocity and attack angle of each part. Then, the total aerodynamic force of the parafoil is the sum of the aerodynamic forces of these eight segments. The equations of lift and drag can be represented as follows: FLi = kiCLi0.5\ud835\udf0cSi \u221a u2 i + w2 i [ wi 0 \u2212ui ] (22) FDi = \u2212CDi0.5\ud835\udf0cSi \u221a u2 i + v2 i + w2 i [ ui vi wi ] , (23) where k denotes the product factor, C denotes the aerodynamic coefficient, S denotes the area of canopy, the subscript L and D denote the lift and drag, respectively, and i denotes the serial number of each segment of canopy", + "1 However, its lift and drag coefficient can be expressed as CL\ud835\udc53 = CL\ud835\udefc(\ud835\udefcc \u2212 \ud835\udefc0)cos2\ud835\udf00 + kArsin2(\ud835\udefcc \u2212 \ud835\udefc0) cos(\ud835\udefcc \u2212 \ud835\udefc0) (26) CD\ud835\udc53 = CD\ud835\udc530 + C2 L\ud835\udefc(1 + \ud835\udeff)(\ud835\udefcc \u2212 \ud835\udefc0)2 \ud835\udf0bAr + kArsin3(\ud835\udefc \u2212 \ud835\udefc0), (27) where the subscript Lf and Df denote the lift and drag coefficient of the flap, \ud835\udf00 denotes the anhedral angle of the parafoil. Moreover, \ud835\udefcc can be obtained by \ud835\udefcc = \ud835\udefc + \ud835\udefc\ud835\udc53 \u2212 \ud835\udefcr, (28) where \ud835\udefcr denotes the rigging angle, \ud835\udefcf denotes the deflection angle of the flap, it is also the horizontal control quantity of the parafoil. Then, the transformation matrix between the local coordinate of each segment of the canopy and the parafoil coordinate can be achieved through rotating \ud835\udefe i around the axis xi, as shown in Figure 5. The transformation matrix is presented as follows: Ti\u2212Os = [ 1 0 0 0 cos \ud835\udefei sin \ud835\udefei 0 \u2212 sin \ud835\udefei cos \ud835\udefei ] . (29) At last, the total aerodynamic force and moment of the parafoil can be shown as follows: F aero s = 8\u2211 i=1 Ti\u2212Os (FLi + FDi) + T\ud835\udc53 r\u2212Os (FL\ud835\udc53 r + FD\ud835\udc53 r) + T\ud835\udc53 l\u2212Os (FL\ud835\udc53 l + FD\ud835\udc53 l) (30) Maero s = 8\u2211 i=1 LOi \u00d7 Ti\u2212Os (FLi + FDi) + LO\ud835\udc53 r \u00d7 T\ud835\udc53 r\u2212Os (FL\ud835\udc53 r + FD\ud835\udc53 r) + LO\ud835\udc53 l \u00d7 T\ud835\udc53 l\u2212Os (FL\ud835\udc53 l + FD\ud835\udc53 l), (31) where the subscript f l and fr denote the coefficient of the left and right flap, respectively, LOi denotes the position vector of the origin of local coordinate Oi in the parafoil coordinate system, LO\ud835\udc53 r and LO\ud835\udc53 l are also designed as LOi " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001637_s11044-019-09705-0-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001637_s11044-019-09705-0-Figure2-1.png", + "caption": "Fig. 2 A 3-RRR parallel manipulator", + "texts": [ + " According to the conservation of work and energy principle, the change of the manipulator total energy (kinetic T and potential U ) should equal the total work of the actuator torques, \u03c4 , and external forces, F , within any time interval [25], T + U = \u222b t t0 \u03c4 T \u03b8\u0307 dt + \u222b t t0 F T z\u0307dt (44) where T and U are the changes in the manipulator kinetic energy and potential energy between the initial time t0 and final time t , respectively. Equation (44) is used to validate the model developed in this work. The developed scheme is used to simulate the trajectory tracking control of a 3RRR planar parallel manipulator. The manipulator geometry was presented by the author in a previous article [31] and is repeated here for the convenience of the reader. A schematic diagram of the manipulator is shown in Fig. 2. The manipulator consists of a moving equilateral triangular platform of side length h connected to a fixed equilateral triangular base of side length d by three limbs. Each limb consists of two links; the first link is connected to the ground by means of a revolute joint identified by the letter Bi and is actuated by a rotary actuator whereas the second link connects the first link to the moving platform with two passive joints at Ai and Ci . The manipulator has three degrees of freedom, therefore, three actuators, one for each limb, are needed to control the moving platform", + "5 kg, mp = 3 kg, and Ip = 0.03 kg m2. Each link of the links (BiAi ) and (AiCi) is considered as slender and its mass moment of inertia about an axis passing through the its center of mass and parallel to z-axis is mil 2 i /12, where mi and li are the mass and length of the link, respectively. The external forces, F = [Fx Fy \u03c4\u03d5]T , are assumed to be constant during the motion with the following values: Fx = 20 N, Fy = 10 N, and \u03c4\u03d5 = 0 N m. For the present manipulator, the constraint equations of the manipulator (see Fig. 2) can be written as hi(q) = hi(\u03b8 ,z) = 0, i = 1, . . . ,3 (45) For the given manipulator, hi = [ x \u2212 xBi \u2212 a cos \u03b8i \u2212 l cos(\u03b1i + \u03d5) ]2 + [ y \u2212 yBi \u2212 a sin \u03b8i \u2212 l sin(\u03b1i + \u03d5) ]2 \u2212 b2 = 0 (46) Then, from Eqs. (3) and (4), J \u03b8 (q) = \u23a1 \u23a2\u23a2\u23a3 \u2202h1 \u2202\u03b81 \u2202h1 \u2202\u03b82 \u2202h1 \u2202\u03b83 \u2202h2 \u2202\u03b81 \u2202h2 \u2202\u03b82 \u2202h2 \u2202\u03b83 \u2202h3 \u2202\u03b81 \u2202h3 \u2202\u03b82 \u2202h3 \u2202\u03b83 \u23a4 \u23a5\u23a5\u23a6 = \u23a1 \u23a3 h11 0 0 0 h22 0 0 0 h33 \u23a4 \u23a6 (47) J z(q) = \u23a1 \u23a2\u23a2\u23a3 \u2202h1 \u2202x \u2202h1 \u2202y \u2202h1 \u2202\u03d5 \u2202h2 \u2202x \u2202h2 \u2202y \u2202h2 \u2202\u03d5 \u2202h3 \u2202x \u2202h3 \u2202y \u2202h3 \u2202\u03d5 \u23a4 \u23a5\u23a5\u23a6 = \u23a1 \u23a3 h1x h1y h1\u03d5 h2x h2y h2\u03d5 h3x h3y h3\u03d5 \u23a4 \u23a6 (48) Different terms of Eqs" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000750_j.mechmachtheory.2018.11.002-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000750_j.mechmachtheory.2018.11.002-Figure2-1.png", + "caption": "Fig. 2. Coordinate systems used for generation of the reference shaper.", + "texts": [ + " If parabolic profiles are used, the corresponding parabolic coefficient is needed to define the profile in its own coordinate system S tr following derivations similar to those presented above. The origin of coordinate system S tr is at point T 1 (see Fig. 1 (a)) that is determined based on the selected toprem height h tr . The angle of toprem \u03c4 tr is referred to the tangent of the based profile (straight of parabolic) at point T 1 , and represents a discontinuity on the profile if it is selected higher than zero. Fig. 2 shows the coordinate systems used for generation of the reference shaper. It represents a conventional generation of a spur gear by a cutter defined in coordinate system S c ( x c , y c , z c ). Based on the coordinate systems represented in Fig. 2 , the family of rack cutter profiles in coordinate system fixed to the shaper is represented as r s (u, \u03d5 s ) = M sc (\u03d5 s ) \u00b7 r c (u ) (5) where M sc (\u03d5 s ) = \u23a1 \u23a2 \u23a3 cos \u03d5 s sin \u03d5 s 0 r ps ( sin \u03d5 s \u2212 \u03d5 s cos \u03d5 s ) \u2212 sin \u03d5 s cos \u03d5 s 0 r ps ( cos \u03d5 s + \u03d5 s sin \u03d5 s ) 0 0 1 0 0 0 0 1 \u23a4 \u23a5 \u23a6 (6) The profile of the reference shaper will be obtained in coordinate system S s considering simultaneously the family of rack cutter profiles in S s given by Eq. (5) and the equation of meshing, that for planar meshing in the plane z s = z c = 0 , can be represented as f sc (u, \u03d5 s ) = n s (u, \u03d5 s ) \u00b7 ( \u2202r s (u, \u03d5 s ) \u2202\u03d5 s ) = 0 (7) Here, n s ( u , \u03d5s ) is normal to the rack cutter profile represented in coordinate system S s " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000769_978-3-662-43893-0-Figure7.3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000769_978-3-662-43893-0-Figure7.3-1.png", + "caption": "Fig. 7.3 Nominal flank data", + "texts": [ + " For these reasons, with bevel gears one rather uses 3D measurement of tooth flank topography compared to that of the desired tooth flank. Using a tooth flank generator (see Sect. 3.3.1) which accounts for tooth flank modifications, it is possible to calculate the exact 3D coordinates and the corresponding normal vector at any point on a tooth flank. The coordinates of 294 7 Quality Assurance points on a desired measuring grid are calculated, along with tooth thickness which is given in the form of a tooth thickness angle (see Fig. 7.3). The resulting target grid data is fed to a 3D gear measuring machine in a specific file format. In order to obtain sufficient information to analyze the measured results, a measurement grid must be sufficiently fine and should cover most of the tooth flank area. The following measurement grid recommendations are generally used: \u2013 A measurement grid with five lines and nine columns is often adequate for machine capability tests and production checks. \u2013 More precise analyses can be performed using measurement grids with 39 39 points, naturally requiring longer measuring times", + " 296 7 Quality Assurance For quantitative analysis, it is possible to output the five parameters used to evaluate the ease-off (see Fig. 3.15), i.e. pressure and spiral angle errors, profile and lengthwise crowning, and bias, these parameters relating to the deviations measured between the desired and actual tooth flanks rather than to the ease-off. Neither DIN 3965 nor ISO 17485 defines tolerances for the tooth flank form of bevel gears. Another result obtained from tooth flank form measurements on bevel gears is tooth thickness deviation in the transverse section, which stems from the difference between the theoretical (see Fig. 7.3) and measured tooth thickness angle. This difference may be converted to give deviations in the values normally used, i.e. normal circular tooth thickness smn or chordal tooth thickness smnc. In addition to the desired measuring tooth flank data, coordinates to measure points on tooth tip, or the face cone, and tooth root, or the root cone, are calculated and supplied to the measuring device. The result of the face cone measurement does not depend on the gear cutting process and is only a check on the form of the blank, which may, however, be the cause of other deviations" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000029_rpj-04-2013-0045-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000029_rpj-04-2013-0045-Figure1-1.png", + "caption": "Figure 1 Island scan strategy", + "texts": [ + " It offers good weldability and covers a wide range of applications, from aerospace to mechanical engineering, pumps, valves and others. Table I gives the chemical composition and the particle size distribution of the used stainless steel 17-4PH powder. The powder was processed on a Concept Laser-M2 machine. This machine is equipped with an Nd:YAG solid state laser with a maximum laser power of about 190W at the powder platform. The scan strategy used to produce the samples in this study is a chess-board-like structure (Figure 1), where 5 5-mm2 squares (islands) are scanned. More details can be found in (Spierings and Levy, 2009; Badrosamay et al., 2009). The following processing parameters are known to play a significant role regarding the final material porosity: laser power PL [W], scan speed vs [mm/s] and layer thickness d [ m]. The scanning hatch distance was set to a fix value to minimize the amount of specimens to be produced. As can be expected, AM-processed materials show a specific mechanical anisotropy due to the layer-wise build process and the corresponding microstructure (Spierings et al", + " Half of the produced specimens were annealed according to the following procedure (H925): Homogenization at 1,350\u00b0C, solution annealing at 1,050\u00b0C, deep freezing and ageing at 495\u00b0C/4h. The static mechanical tests were performed on a ZWICK Z-1484 testing machine, equipped with a 20 kN load cell. Each specimen was pre-loaded with 20MPa; the strain rate was set to 0.001 s 1. A model for the SLM scanning process is developed taking into account the characteristics of the island scan strategy used (Figure 1). The number n of parallel scan lines in one island is n L/a1, resulting in a total scan length S n\u00b7L. Additionally, there is an overlap a2 between neighbouring islands, so that the total scan length Stot n\u00b7(L 2\u00b7a2). Principally, two strategies can be used for the transition of a scan line to the neighbouring one: meander or bi-directional scanning. For meander scanning, the total scanning length in one island is Stot, m n\u00b7(L 2\u00b7a2) L, for bi-directional scanning, there is a deceleration and acceleration length on each side of a scan line" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000929_j.measurement.2016.03.077-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000929_j.measurement.2016.03.077-Figure3-1.png", + "caption": "Fig. 3. The measurement error of the hypoid gear.", + "texts": [ + " The non-coincidence error between the theoretical tooth flank and the actual measured tooth flank could be represented by six parameters: Dx, Dy, Dz (translation error along X-axis, Y-axis and Z-axis, respectively) and Dhx, Dhy, Dhz (rotation error around X-axis, Y-axis and Z-axis, respectively). According to the actual situation of CNC gear measuring machine, the most influenced parameters are Dz and Dhz, which caused by the machining error of the large end and the rotation error of the rotary table, and these two errors need to be researched and compensated first, as shown in Fig. 3. Using the rigid matching method of the curved surface, this paper calculates the deviations of the parameters Dz and Dhz, and matches the actual tooth equidistant surface with the theoretical equidistant surface, as shown in Fig. 4. As shown in Fig. 4, when measuring the tooth flank, as the tooth form deviation, Dz and Dhz exist, the readings of the probe is not zero and named as d(dx, dy, dz). So the measured points Prs(xrs, yrs, zrs) on the actual tooth equidistant surface could be yielded as follows" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002436_tmag.2021.3087460-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002436_tmag.2021.3087460-Figure3-1.png", + "caption": "Fig. 3 Single-stator-dual-rotor AFPM motor topology (a) and manufactured AFPM motor components (b)", + "texts": [ + " In the first step the corresponding harmonics content of the phase current is identified after based on the cogging torque FFT analysis. This current is then used to calculate the q-axis current (\ud835\udc56\ud835\udc5e,\ud835\udc58) to compensate the corresponding cogging torque harmonic. If the desired reduction is not achieved in the first step, then the second step is followed in a similar manner. The flow diagram of this process is given in Fig. 2. An integral slot, double-rotor-single-stator AFPM motor with two 8-pole rotors is used in this paper as in Fig. 3(a). The stator includes a strip wound M250-35A core and two sets of lap type winding while the rotor includes a mild-steel disc with high relative permeability and planer surface-mounted magnets with fan shaped and triangular shaped magnets as in Fig. 3(b). The motor is specifically designed with integral slot combination with rectangular shaped slot in order to obtain high cogging torque component. Two different AFPM motor is used in order to confirm the proposed harmonic injection method. Both motors do have the same strip stator with different rotor structures and different motor parameters including cogging torque profile. TABLE I - Specifications of the Reference Axial Flux PM Motor AFPM Motor - 1 AFPM Motor - 2 Air gap 0.8 mm 2 mm Number of poles 8 8 Number of stator slots 24 24 Slot \u2013per-pole-per phase 1 1 Slot height/Slot width 16 mm/7 mm 16 mm/7 mm Stator ID/OD 100/178 100/178 Magnet shape and type Triangular shaped NdFeB Triangular shaped NdFeB Stator Resistance (Phase-Phase) 188 m\u2126 188 m\u2126 Inductance (Phase- Phase) 522" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000765_1.4966628-Figure8-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000765_1.4966628-Figure8-1.png", + "caption": "Fig. 8. An ellipse is parameterized by r\u00f0a\u00de \u00bc \u00f0Rc sin a;R cos a\u00de. Notice that u is not equal to h (the angle that r forms with the vertical), and both angles are different from a; they are related by tan h \u00bc c tan a \u00bc c2 tan u.", + "texts": [ + " Find the curvature of such functions and demonstrate that for k> 2 there is no finite initial speed that will cause the particle to immediately leave the surface (beginning at the origin). Demonstrate that any curve with an inflection point at the origin also shows this behavior. Study in detail the case of the parabola (k\u00bc 2), showing that the particle never leaves the surface, regardless of the friction coefficient and the initial speed. We can parameterize the ellipse by r\u00f0a\u00de \u00bc \u00f0Rc sin a; R cos a\u00de, with a 2 \u00bd0; p=2 and c> 0, which represents the curve shown in Fig. 8. Show that the maximum initial speed that can be given to the particle is v2 0 gR ! max\u00f0 \u00de \u00bc c2: (46) 113 Am. J. Phys., Vol. 85, No. 2, February 2017 Gonz alez-Cataldo, Guti errez, and Y a~nez 113 Another interesting, famous curve is the nephroid. A possible parameterization for this curve is r a\u00f0 \u00de \u00bc R 4 3sin a\u00f0 \u00de\u00fe sin 3a\u00f0 \u00de;3cos a\u00f0 \u00de\u00fe cos 3a\u00f0 \u00de ; (47) which is shown in Fig. 9. Obtain the speed of the particle in terms of the tangential angle u, the horizon of the curve, and the maximum initial speed (when the particle immediately leaves the surface)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002413_tie.2021.3078388-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002413_tie.2021.3078388-Figure1-1.png", + "caption": "Fig. 1. Schematic diagram of the resolver For the ideal resolver, the output signal Usin and Ucos can be expressed as", + "texts": [ + " In Section III, the field-oriented control (FOC) model and the electromagnetic finite element model (FEM) are established to simulate the influence of rotor position error on current harmonics and electromagnetic force with and without position error. Finally, the theoretical analysis is verified by noise test in Section IV. II. THEORETICAL ANALYSIS The position sensor adopted in the driven motor is the resolver, which is a variable reluctance motor. The resolver contains two orthogonal windings and an excitation winding. When the excitation winding is input with a high-frequency AC signal, sine and cosine signals will be induced in the two orthogonal windings. The schematic diagram of the resolver is shown in Fig.1. sin cos sin cos U KE U KE (1) where, KE is the excitation level of the resolver, \u03b8 is the electric angle of the resolver. The error voltage is written as sin coscos sin sin - errU U U KE (2) where, \u03c6 is the output decoded angle of the resolver. The error voltage should be as small as possible, so that the output angle \u03c6 is closer to the actual angle \u03b8. When the error voltage is equal to zero, the output angle of the ideal resolver is calculated from sin cos = arctan . U U (3) The actual working conditions of driven motors for vehicle applications are relatively poor, especially for the in-wheel motors" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002424_j.neunet.2021.06.001-Figure7-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002424_j.neunet.2021.06.001-Figure7-1.png", + "caption": "Fig. 7. Experimental platforms: Laikago and Lilibot.", + "texts": [ + " This synaptic plasticity provides the reflex network with fast nline adaptation, because the DIL can adjust the reflex gains nline (i.e., determined by w1,2(n)) depending on the sensory timulation changes (see Fig. 6(b)). For instance, when the input hange is large (small), the gains increase (decrease) for realizing daptive synaptic weights (w1,2(n)). This results in joint offset daptation. . Experiments and results In this study, we performed three main experiments on a mall-sized quadrupedal robot (Lilibot Sun et al., 2020) in simlation (Fig. 7), to evaluate the performance of the DFRL for he AQMC. The experiments consisted of (I) trotting on a level round, (II) trotting on various slopes, and (III) trotting on a omplex terrain with multiple slopes (see Fig. 8). The traditional 3 Synaptic plasticity (neuroscientific term): The ability of synapses to trengthen or weaken over time in response to increases or decreases in activity. e import this term to describe the similar abilities of the artificial neural etwork (DFFB network). v vestibular reflex was also evaluated by implementing it instead of the DFRL for comparison; this reflex utilizes a level-body posture strategy through which the robot body is maintained parallel to the horizontal surface (known as telescoping strut) (Fukuoka et al., 2003; Yu, Zhou, Qian, & Xu, 2018). This differs from the lever mechanics strategy utilized by the DFFB reflex in the DFRL. The two strategies are shown in Figs. S.7 (b) and (c) in the Supplementary material. The vestibular reflex scheme is outlined in Fig. S.8 of the Supplementary material. In addition to the comparative experiments on Lilibot, the AQMC with the DFRL was also implemented in a larger quadrupedal robot (Laikago, Fig. 7), to demonstrate the DFRL\u2019s generalizability to different platforms. Fig. 7 shows the experimental platforms: Lilibot and Laikago. These two quadruped robots have different sizes, and their specifications can be seen in Table S.2 in the Supplementary material. The robot experiments were performed on three types of terrains, as shown in Fig. 8(a), (b), and (c). The robots and terrains were simulated using CoppeliaSim4 with Vortex.5 These served as a robot operating system (ROS) node and communicated with the AQMC through certain ROS topics. The simulation platform featuring Lilibot and Laikago can be seen in https://gitlab" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002487_acsami.1c08722-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002487_acsami.1c08722-Figure1-1.png", + "caption": "Figure 1. Schematic illustration of the bistable actuator and the reconfigurable 3D electronics. (a) Casting the precursor sheet with PDMS center and silicon oil/PDMS ring. (b) Extracting the silicon oil from the silicon oil/PDMS ring by a solvent bath (e.g., dichloromethane) yields a bistable structure. UVO treatment on one surface of the bistable structure introduces a thin oxidized layer, which can absorb a polar solvent (e.g., ethanol) for expansion. (c) Bonding the flexible electronic device to the bistable actuator. (d) PDMS actuator and the device can convert configuration under the action of solvent stimuli (e.g., n-hexane and ethanol).", + "texts": [ + " The kirigami polyimide (PI) sheets adhered to the bistable actuator with different bonding modes can be driven to diverse 3D structures from the original bending configuration, and the process is reversible. This approach not only provides a potential way to reconstruct the 3D electronics but also enriches the design of multifunctional 3D electronic devices. A frequency-reconfigurable electrically small monopole antenna driven by the bistable actuator is presented as an application demonstration. It is expected that the novel method can be applied to many other areas in 3D electronics. \u25a0 RESULTS AND DISCUSSION Figure 1 illustrates the fabrication process of the bistable actuator and the reconfigurable 3D electronics. The bistable structure originates from a hybrid poly(dimethylsiloxane) (PDMS) circle sheet, whose inner circle is cast by the PDMS/curing agent mixture and the outer ring is cast by the silicon oil/PDMS/curing agent mixture (Figure 1a). The oil/ PDMS annual ring shrinks after silicon oil extraction in an organic solvent bath30 (e.g., dichloromethane) for 48 h. The shrinkage of the annual ring pushes the inner PDMS disk radially. When the load reaches a critical level, the sheet may suddenly change shape and the whole sheet buckles to a domelike structure. This PDMS structure can be flipped to reversed state and flipped back to the original state through snapping stimuli. One surface of the bistable structure is then oxidized by UV/ozone (UVO) treatment to introduce a thin film for polar solvent expansion31 (e", + "), as shown in Figures 1b and S1a. The other surface remains unoxidized, which can absorb a nonpolar solvent for expansion32 (e.g., n-hexane, ethyl acetate, etc.). The expansion of the concave surface caused by the corresponding solvent reversibly snaps the bilayer PDMS structure to another configuration. The alternative expansion of these two layers under different solvents leads to a smart bistable actuator. The kirigami PI sheets are bent and selectively bonded to one side of the bistable actuator (Figure 1c). Furthermore, the sheets can be driven to diverse 3D configurations by the snapping behavior of the actuator, shown in Figure 1d. To better understand the buckling deflection of the bistable structure, the nondimensional deflections (the ratio of the outplane deflection to PDMS sheet radius R1) with different fabrication parameters,33 shown in Figure 2a, such as the radius ratio (\u03b2) of the inner circle (R2) to the outer circle (R1), the nondimensional slenderness (\u03b1 = R1/T) of the PDMS sheet, and the linear shrinkage ratio (\u03b3) of the annular ring were systematically studied. The linear shrinkage ratio (\u03b3)30 of the annular ring can be estimated by \u03b3 = 1\u2212(1 \u2212 \u03c6)1/3, where \u03c6 indicates the silicon oil volume fraction before extraction, i" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000444_s00707-014-1145-x-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000444_s00707-014-1145-x-Figure1-1.png", + "caption": "Fig. 1 Description of the EHL problem with multiple subsurface cracks", + "texts": [ + " It can be predicted that the solution may have potential applications in studying the crack nucleation and propagation for lubricated mechanical components. 2.1 Problem description and solution approach Consider the EHL contact between an infinitely long cylinder of radius R and a half-space matrix bounded by the plane surface z = 0 with the lubricant covered in the Cartesian coordinate system at an entrainment velocity along the x-coordinate U . The half-space contains cracks \u03d5(\u03d5 = 1, 2, . . . , m) beneath the contacting surfaces (Fig. 1a). The cylindrical body has Young\u2019s modulus Eu and Poisson\u2019s ratio \u03c5u while the half-space matrix has Es and \u03c5s . When the cylinder is subjected to an external load W , the pressure generated within the lubricant causes the deformation of both the loading body and the half-space matrix, while cracks beneath the half-space would also response to the stress induced by the pressure and cause the deflection of the contacting surfaces, which in return affects the fluid pressure and the lubricant film thickness", + " The DDT is employed to model each crack of mixed modes I and II as a continuous distribution of climb and glide dislocations with unknown densities \u03c1\u22a5 and \u03c1 to be determined. As a result, the original problem concerning materials with multiple cracks is converted into a homogeneous EHL contact problem with extra elastic surface deformation induced by the edge dislocations. The solution approach is then to decompose the new problem into two subproblems: a half-space subproblem with prescribed surface loading and a homogeneous EHL contact problem (Fig. 1b). The half-space subproblem is used to determine the displacement due to subsurface cracks for a prescribed surface loading, while the homogeneous EHL contact problem determines the fluid pressure distribution and lubricant film thickness due to the external load. An iterative algorithm is developed to integrate the two subproblems, and the iteration process is conducted until the convergence of the half-space surface displacements, which involve the displacements due to the fluid pressure and subsurface cracks" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000088_j.mechrescom.2012.12.007-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000088_j.mechrescom.2012.12.007-Figure2-1.png", + "caption": "Fig. 2. Contact status of single pair of gear teeth in cycloid reducer.", + "texts": [ + " In addition, an ideal profile modification curve for even load distribution can also be found by solving this equation. This method was proved as an effective way to improve the situation of heavy-duty tooth contact models. 2. Materials and methods 2.1. Single tooth meshing mathematical model Under certain assumptions, we can establish mathematical model of the single tooth meshing under considered contact area shape and deformation transfer effect of discrete contact p i l w p r n i u n t u e w i c p s i e oints. The single tooth meshing status under loads is shown n Fig. 1. As shown in Fig. 2, the pressure on single tooth meshing under oads can be shown as below: \u2212f (xi, yi) + \u0131 = k\u2211 j=1 (Aij + Bij)pj + Ei k\u2211 j=1 F(Pj) = W { Pj \u2265 0, (xj, yj) \u2208 \u02dd Pj = 0, (xj, yj) /\u2208 \u02dd \u23ab\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23ac \u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23aa\u23ad (1) here F is unit center pressure, Pj is unit pressure function of indeendent variable, W is the normal total load of pin tooth, and it elates to the force of pendulum wheel, Aij is a flexibility compoent of contact deformation for unit j under unit i tooth profile, Bij s a flexibility component of pin teeth deflection for unit j under nit i tooth profile, Ei is processing and assembly errors of center ormal tooth for unit i, \u0131 is normal contact deformation between eeth, and f(xi, yi) is an imaginary shape at any point i of pin teeth nder-load without considering contact deformation" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000769_978-3-662-43893-0-Figure2.9-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000769_978-3-662-43893-0-Figure2.9-1.png", + "caption": "Fig. 2.9 Hypoid gear geometry definition [ISO23509]", + "texts": [ + " On bevel gears, Section A-A is defined as the transverse section and is always perpendicular to the pitch cone. Thus, Section A-A is not a plane section in itself, but rather corresponds to the complementary cone at the point under consideration. In Fig. 2.8, the complementary cone including the mean point is unrolled into an imaginary plane where the original mean pitch (cone) diameter dm is transformed into the equivalent pitch diameter dv\u00bc dm/cos \u03b4; all dimensions found in Fig. 2.8 are detailed in Table 2.3. The corresponding main dimensions for hypoid gears are given in Fig. 2.9 and described in Table 2.4. 2.2 Gear Geometry 21 22 2 Fundamentals of Bevel Gears 2.2 Gear Geometry 23 24 2 Fundamentals of Bevel Gears If a non-offset bevel gear is rolled on its fixed mating gear, one point of the rotating tooth flank will move over a spherical surface whose center is the crossing point of their axes. The tooth profile corresponding to this point is obtained from the intersection of the bevel gear tooth and the surface of the sphere [NIEM86.3] or, with sufficient accuracy, from the unrolled complementary cone (see Sect" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001717_s00170-020-05300-2-Figure5-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001717_s00170-020-05300-2-Figure5-1.png", + "caption": "Fig. 5 Thermal stresses in circular additive manufacturing product: a radial thermal stress, b circumferential thermal stress, and c vertical thermal stress", + "texts": [ + " The difference between the experimental specimen and the numerical model is caused by the melting and the material flow in the melt pool. This difference only occurs at the end of the specimen. The main part of the specimen and the measured data are not affected by this difference. So, the comparison is still valid to show the success of the proposed model. Using the validated sequentially coupled thermal-mechanical model, thermal stresses, residual stresses, and distortions of circle components in DED AM are calculated. The thermal stresses in additive-manufactured product with the radius of 125 mm at 120 s are shown in Fig. 5. The maximum radial thermal stress, circumferential thermal stress and vertical thermal stress is 495.9, 606.8, and 438.1 MPa, respectively. In DED AM of circular component by laser, the circumferential thermal stress is largest in the stress components. Especially, the asymmetrical distribution of thermal stresses is found in the additive-manufactured circle component. Larger compressive thermal stresses are found in the region near the initial manufacturing point, as shown in Fig. 5. To study the reason for the asymmetrical distribution of thermal stresses, temperature histories at different points, as shown in Fig. 6a, in the 1st layer are shown in Fig. 6b. At the initial point, different temperature histories are found comparing with that of other points. When the manufacturing process begins, the initial point is heated. Due to the fast moving laser to point A, the initial point is not fully heated and a temperature rise of 423.05 \u00b0C is found. It is different with other points" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000769_978-3-662-43893-0-Figure3.3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000769_978-3-662-43893-0-Figure3.3-1.png", + "caption": "Fig. 3.3 Profile and lengthwise crowning as parts of the ease-off", + "texts": [ + " In other cases it can be chosen as desired, but its maximum value must meet two limits: \u2013 the maximum tool edge radius, limited by the clearance c \u03c1a0, lim1,2 \u00bc c1,2 1 sin \u03b1nD,C1,2 \u00f03:10\u00de \u2013 the maximum tool edge radius, limited by the minimum slot width efn,min \u03c1a0, lim1,2 \u00bc 0, 5 efnmin cos \u03b1nD,C1,2 1 sin \u03b1nD,C1,2 \u00f03:11\u00de The calculation ismade on the virtual crown gear for identical edge radii on the concave and convex tooth flanks. To prevent meshing interference or cut-off in the tooth root, the chosen tool edge radius should be smaller than the lesser of the two above limits. Crowning This is differentiated into profile and lengthwise crowning (see Fig. 3.3). High crowning values lead to less sensitivity of the contact pattern to displacements of the gears, but also to a smaller contact pattern, a concentration of loads entailing high flank pressures as well as higher local tooth root stresses. An optimal modification of bevel gear tooth flanks can be established only if numerical simulation is available to calculate the exact tooth geometry, load-free and loaded tooth contact analysis. The latter must also include the relative displacements of the pinion and wheel induced by deflections" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000374_978-3-319-09489-2_16-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000374_978-3-319-09489-2_16-Figure3-1.png", + "caption": "Fig. 3 Schematic representations of a RRPM with m = 9 and n = 6. Source [2]", + "texts": [ + " Now these two approaches, the ICFM and the PM, get combined into the Improved Puncture Method (IPM). Thus, the objective is to obtain a method which, on the one hand, is very fast with a strictly bounded worst-case computation time and which is therefore real-time capable. On the other hand, the resulting cable force distributions shall lie on a preferably low tension level. To check, whether the IPM meets the declared objectives, it was exemplarily analyzed for a spatial parallel robot with six degrees of freedom and nine cables (as illustrated in Fig. 3). 2 See [6, 7] for more information. To measure the covered workspace, the available space inside of the robot frame (size 1.48 m\u00d71.36 m\u00d71.00 m) was discretized and the distance between the grid points was set to 0.02 m. Then it was checked for which grid points a feasible cable force distributions could be found. The cable force boundaries were defined as fmin = 10 N and fmax = 100 N. In Fig. 4a the workspace covered by the CFM is represented. For 36,648 points a feasible force distribution could be found" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001682_ffe.13199-Figure6-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001682_ffe.13199-Figure6-1.png", + "caption": "FIGURE 6 The integrated finite element model for gear rolling contact fatigue [Colour figure can be viewed at wileyonlinelibrary.com]", + "texts": [ + " The resulted surface roughness after geometric processing is presented in Figure 5B. The curvature at any point on the processed curve is equal to the curvature of the corresponding point on the tooth profile. Since this work mainly concerns the fatigue behaviour of the material points in the RVE domain, the microgeometry of roughness is only assigned at the surface of the RVE domain, and the remaining tooth profile is kept smooth. Numerical simulation of the gear rolling contact is implemented using the ABAQUS static general approach, as shown in Figure 6. The rolling direction and the depth direction are also depicted in Figure 6A. The actual meshing process of the studied gear pair is simulated by applying the rotation on driving gear and simultaneously a torque of 210 kNm on driven gear. The finite sliding surface-to-surface contact between the driving and driven gear teeth is set, and the tooth surfaces in contact is assumed to be free of friction. The generated microstructure in the RVE domain, the inclusion on the GB, and the surface roughness along the tooth profile are partially enlarged, as shown in Figures 6B to 6D" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001447_j.mechmachtheory.2019.03.002-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001447_j.mechmachtheory.2019.03.002-Figure1-1.png", + "caption": "Fig. 1. Kinematic diagram of PGT-I.", + "texts": [ + " The difficulty of planetary gear trains model lies in the combination of self-rotation and center-following rotation of the planetary gear while the carrier rotates. Bond graph method provides a rational way to model planetary gear trains [17] . The planetary gear trains has two types. First type has a single planetary gear transmission between the ring gear and the sun gear. In this paper it is called PGT-I type. The other has two planetary gears transmission between the ring and the sun, which is called PGT-II type. For PGT-I, first we need to establish equations of motion relations and set variables as shown in Fig. 1 . The parameters of gears are denoted as follows: pitch radius of the sun ( r s ), pitch radius of the planet ( r p ), pitch radius of the ring ( r r ), rotational speed of the sun ( \u03c9 s ), rotational speed of the carrier ( \u03c9 c ), rotational speed of the planet ( \u03c9 p ), rotational speed of the ring ( \u03c9 r ). The positive definition of all velocities is shown in Fig. 1 . Kinematic equations can be drawn from Fig. 1 as follows: \u03c9 r1 = \u03c9 r \u2212 \u03c9 c (1) \u03c9 p r p = \u03c9 r1 r r (2) \u03c9 s 1 = \u03c9 s \u2212 \u03c9 c (3) \u03c9 p r p = \u2212\u03c9 s 1 r s (4) The \u03c9 r 1 means relative rotational speed of the ring from the carrier. Then the bond graph structure can be established in Fig. 2 . The kinetic relations will be established after the causality definition. As shown in Fig. 2 There are three ports which connect to the external system, namely, sun gear, planet carrier and ring gear. Planetary gear is a transmission link in the system, which has no direct contact with the external system" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002140_03772063.2021.1909507-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002140_03772063.2021.1909507-Figure1-1.png", + "caption": "Figure 1: Quadcopter\u2019s configuration", + "texts": [ + " This is accomplished by a series-parallel model containing these networks and updating the weights of these using a backpropagation method, based on the error in the predicted output. In ANN-based NARMA-L2 controllers, design of controller consists of two steps. The first step is to identify the dynamics of the system to be controlled by training f(\u00b7) and g(\u00b7) submodels. Second step is to design NARMA-L2 controller and approximateu(k)using f(\u00b7) and g(\u00b7) trained at the previous step. The dynamics model of a quadcopter is described in many existent approaches [25\u201327]. The essential frames include an Earth frame, E, and body frame, B, as shown in Figure 1. Let \u03c6, \u03b8 , and\u03c8 denote the three Euler angles roll, pitch, and yaw, respectively. x, y, and z respectively represent the position of the quadcopter in the earthfixed frame {E}. Jx, Jy, and Jz denote the moments of inertia along the x, y, and z axes, respectively; m represents the mass, l the arm length of the vehicle, and g the gravitational acceleration. The quadcopter dynamics model can be described as follows: \u03c6\u0308 = 1 Ix [(Iy \u2212 Iz)\u03c8\u0307 \u03b8\u0307 \u2212 kfax\u03c6\u03072 \u2212 Jr r \u03b8\u0307 + lu2] \u03b8\u0308 = 1 Iy [(Iz \u2212 Ix)\u03c8\u0307\u03c6\u0307 \u2212 kfay\u03b8\u03072 + Jr r\u03c6\u0307 + lu3] \u03c8\u0308 = 1 Iz [(Ix \u2212 Iy)\u03c6\u0307\u03b8\u0307 \u2212 kfaz\u03c8\u03072 + lu4] (4) x\u0308 = 1 m [\u2212kftxx\u0307 + (cos\u03c6sin\u03b8cos\u03c8 + sin\u03c6sin\u03c8)u1] y\u0308 = 1 m [\u2212kftyy\u0307 + (cos\u03c6sin\u03b8sin\u03c8 + sin\u03c6cos\u03c8)u1] z\u0308 = 1 m [kftzz\u0307 + (cos\u03c6cos\u03b8)u1] \u2212 g where Ui, i = 1,2,3,4 denotes the control inputs of a quadrotor, which are computed as\u23a7\u23aa\u23aa\u23aa\u23a8 \u23aa\u23aa\u23aa\u23a9 U1 = b(\u03c92 1 + \u03c92 2 + \u03c92 3 + \u03c92 4) U2 = b(\u03c92 4 \u2212 \u03c92 2) U3 = b(\u03c92 3 \u2212 \u03c92 1) U4 = d(\u03c92 1 \u2212 \u03c92 2 + \u03c92 3 \u2212 \u03c92 4) (5) where Ui (i = 1,2,3,4) represents the angular speed of motors and g = 9" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001950_j.jfranklin.2020.07.014-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001950_j.jfranklin.2020.07.014-Figure1-1.png", + "caption": "Fig. 1. Aerodynamic forces and aerodynamic angles of HFV [27] .", + "texts": [ + " 003772 P 1 = C \u03b12 P 1 \u03b1 2 + C \u03b1 P 1 \u03b1 + C 0 P 1 P 2 = C \u03b12 P 2 \u03b1 2 + C \u03b1 P 2 \u03b1 + C 0 P 2 + C \u03b4e P 2 \u03b4e where q\u0304 = \u03c1V 2 / 2 represents aerodynamic pressure, S , m , \u03bc, g represent reference aerodynamic area, aircraft mass, gravity constant and gravity acceleration, respectively; inputs are the elevator deflection angle \u03b4e and the throttle valve opening \u03b2; C L ( \u03b1), C D ( \u03b1) are lift coefficient and drag coefficient respectively, which are mainly affected by the angle of attack \u03b1; C T ( \u03b2) is thrust coefficient, which is only related to the throttle valve opening \u03b2. Consider the elevator redundancy of the HFV \u03b4e = \u2211 2 j=1 l j \u03b4e j and l j is a positive coefficient. For introducing HFV model more conveniently, a profile of HFV is given in Fig. 1 , where some definitions of important aerodynamic forces and aerodynamic angles are shown. This paper considers two types of elevator fault modes: stuck type fault and loss of effectiveness. The elevator surface may work under different working conditions with normal operation or fault, so the working model is established as follows: \u03b4e j (t ) = b j v j (t ) + b j0 , j = 1 , 2 (7) where v j is the applied control input and different values of b j and b j 0 respectively represent several different working conditions of the actuators" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000734_icrom.2017.8466132-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000734_icrom.2017.8466132-Figure1-1.png", + "caption": "Fig. 1. Hexapod table and the vector representation of the ith pod for kinematic modeling.", + "texts": [ + " The contribution of this paper is to study the clearance on the on the kinematics accuracy of the hexapod table, formulating the kinematic error and experimentally evaluating the model. The organization of this paper is as the following: Section II describes the mechanism of a hexapod table. Section III presents the kinematic and quasi-static dynamic of the hexapod table. Then, the error model of the joint clearance on the configuration of the table is introduced in Section IV. This model is then evaluated through experimental studies in Section V. Finally, the conclusions are presented. II. DESCRIPTION OF HEXAPOD TABLE Fig. 1 illustrates the mechanism of a hexapod table. This mechanism consisted of a stationary base platform, a moving platform as an end effector, and six extendable pods. Universal joints connect pods to the base, and spherical joints connect the other side of the pods to the end effector. These actuated joints of this manipulator are prismatic and the actuators of them are located on the pods. A frame is in the center of the moving platform frame that is referred by {P} and a coordinate frame is fixed to the geometric center of the base that is referred as the based coordinate frame {W}. The location and orientation of {P} are obtained with respect to {W} through kinematic analysis. In the hexapod table, two parameters of the position of the moving platform and mechanism basis length is interrelated by mechanism kinematics. By kinematic relations, if one of the parameters is known the other can be determined [19-22]. In Fig. 1, Pai is the ith platform connection point\u2019s position vector in frame {P}; ai and bi are respectively the ith platform and base connection points\u2019 position vectors in the {W}. x is the moving platform center\u2019s position vector. li = ai - bi is the ith pod\u2019s length vector. III. KINEMATICS AND THE POD FORCES The unit vector along each pod\u2019s direction, [ ]Tziyixii nnn=n , can be obtained as [23, 24]: iii lln = (1) Taking the time derivative on li, the ith pod\u2019s linear velocity l can be determined in closed form as the following: = \u2212 x Jl 1 (2) where and x are the angular and linear velocities of the moving platform center in {W}, respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000750_j.mechmachtheory.2018.11.002-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000750_j.mechmachtheory.2018.11.002-Figure3-1.png", + "caption": "Fig. 3. Edge geometry of the circular cutter profile.", + "texts": [ + " Therefore, the profile of the reference shaper is given by { r s (u, \u03d5 s ) = M sc (\u03d5 s ) \u00b7 r c (u ) f sc (u, \u03d5 s ) = n s (u, \u03d5 s ) \u00b7 ( \u2202r s (u,\u03d5 s ) \u2202\u03d5 s ) = 0 (8) The active part of the profile of the circular cutter coincides with the section of the reference shaper and therefore it is given by Eq. (8) . However, the profiles of the circular cutters should be designed as long as possible to help mimicking the surfaces of the reference shaper as much as possible in longitudinal direction when the profile is rotated to form the generating surface of the cutter. When extending the profile of the reference shaper profile given by Eq. (8) , it has to allow for the accommodation of the edge profile of the cutter. Fig. 3 shows the extended profile of the shaper and the edge geometry defined by a circular arc of radius \u03c1e . Let point E 0 be the center of the circular arc defining the edge profile, E 1 the initial point of the edge profile, and E 2 the final point of the edge profile. The top edge circular arc and the active profile of the cutter are tangent at point E 1 . Given the edge profile radius \u03c1e , the longest active profile of the circular cutter is obtained when points E 0 and E 2 lie on the plane of symmetry of the shaper tooth. The coordinate system of the shaper has its y axis directed along the line of symmetry of the tooth (see Fig. 3 ). In order to find the coordinates of the center of the circular arc of the edge profile in coordinate system S s , r E 0 s , the point of tangency with the shaper active profile, E 1 , and the angle of extension of the edge profile, \u03be , the position vector of the center of the edge profile is represented as r E 0 (u, \u03d5 s ) = r E 1 (u, \u03d5 s ) \u2212 n E 1 (u, \u03d5 s ) \u03c1e (9) s s s Vectorial Eq. (9) provides only two scalar equations because it is being considered for any planar cross-section of the shaper" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002098_acs.jpcb.0c09314-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002098_acs.jpcb.0c09314-Figure1-1.png", + "caption": "Figure 1. Schematic of the camphor objects [(a) object I and (b) object II] used in this study. The concentration of SDS aqueous solution was 10 mM. dc is the diameter of the camphor disk. dp is the diameter of the circular plastic sheet in object I. di is the inner diameter of the plastic ring in object II.", + "texts": [ + " We used two camphor objects, that is, objects I and II, as Received: October 14, 2020 Revised: January 6, 2021 Published: January 28, 2021 \u00a9 2021 American Chemical Society 1674 https://dx.doi.org/10.1021/acs.jpcb.0c09314 J. Phys. Chem. B 2021, 125, 1674\u22121679 D ow nl oa de d vi a U N IV O F C A L IF O R N IA S A N T A B A R B A R A o n M ay 1 5, 2 02 1 at 1 4: 53 :4 1 (U T C ). Se e ht tp s: //p ub s. ac s. or g/ sh ar in gg ui de lin es f or o pt io ns o n ho w to le gi tim at el y sh ar e pu bl is he d ar tic le s. shown in Figure 1. Object I was composed of a camphor disk (diameter: 7.0 mm; thickness: 1.0 mm; mass: 38 mg) and a circular plastic sheet (thickness: 0.1 mm). Object II was composed of the same camphor disk and a plastic ring (thickness: 0.1 mm; outer diameter: 7.0 mm). The camphor disk was prepared using a pellet die set for Fourier transform infrared spectroscopy, and the plastic sheet made of polyethylene terephthalate was glued to the center bottom of the camphor disk. The contact area between the disk and the SDS aqueous phase (S in mm2) as a variable parameter was changed using a circular plastic sheet with different diameters (dp in mm) for object I or a plastic ring with different inner diameters (di in mm) for object II", + " In other words, mode bifurcation between oscillatory motion and no motion was observed around S = 18.9 mm2. The frequency of the oscillatory motion was increased at 10.2 \u2264 S \u2264 25.9 mm2 but was almost constant at \u223c0.03 s\u22121 at S \u2265 25.9 mm2. In contrast, vmax was linearly increased with an increase in S at S \u2265 18.9 mm2. The frequency and vmax as a function of dp is shown in Figure S2. To clarify the effect of the shape of the plastic plate attached to the bottom of the camphor disk, we examined the other camphor object (object II) in Figure 1b. Figure 4 shows (a) frequency and (b) the maximum speed of the oscillatory motion for the camphor object (object II) depending on S. No motion and oscillatory motion were observed at S < 19.6 mm2 and S \u2265 19.6 mm2, respectively. In other words, mode bifurcation between oscillatory motion and no motion was observed around S = 19.6 mm2. The frequency of oscillation was increased with an increase in S at S \u2265 19.6 mm2. On the other hand, the maximum speed was increased with an increase in S at S \u2265 19" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000769_978-3-662-43893-0-Figure4.35-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000769_978-3-662-43893-0-Figure4.35-1.png", + "caption": "Fig. 4.35 Tooth deflection calculation model", + "texts": [ + " The influence function, which is therefore also determined diagonally across the tooth, is established for each of the tooth segments not lying below the load application point using their compliance (q (y)) in response to loading at point yF (see also Fig. 4.36). 4.4 Stress Analysis 175 The influence of the finite face width b and the tooth boundary on the fading function E1 of the infinitely long tooth is accounted for by mirroring the deflection curve on the imaginary tooth face of the infinite tooth strip. Deflections lying outside the face width b are superimposed to the deflections within it (Fig. 4.35) [JARA50]. E \u00bc E1 \u00fe \u0394E1 \u00f04:315\u00de The result can be further improved by introducing a corrective function Wf . E \u00bc E1 x \u00bc x V x R \u00fe E1 x \u00bc x V \u00fe x R Wf x V ; x R \u00f04:316\u00de where x V \u00bc xVj j mmn ; x Vmax \u00bc 6; x R \u00bc xRj j mmn ; xV , xR according to Fig:4:35 \u00f04:317\u00de The influence of face canting on skewed and spiral gear teeth is allowed for by means of approximation Wf. Wf \u00bc 1 \u00fe f G f L \u00f04:318\u00de 176 4 Load Capacity and Efficiency An additional function, fL, represents the theoretical fading behavior of the boundary influence on spiral bevel gears, while size function fG realizes the magnification or reduction of the deflection in the vicinity of the tooth boundary at a particular spiral angle when compared to a straight bevel gear", + " Mirroring S \u00bc S1 \u00fe \u0394S1 \u00f04:331\u00de Multiple mirroring is necessary for gear widths b< 6 \u00b7mmn. If this limit goes below significantly, larger errors in the results must be anticipated. Form of the tooth faces W\u03c3 \u03b2St\u00f0 \u00de \u00bc 1 0:72 \u03b2Stj j \u00fe for the acute angled side of the tooth face for the obtuse angled side of the tooth face \u03b2 in radians \u00f04:332\u00de Boundary stress reduction R \u03be \u00f0 \u00de \u00bc tanh \u03be R \u00fe 1 \u00f04:333\u00de Influence function S\u00bc S1 \u03be \u00bc \u03be V \u03be R \u00fe S1 \u03be \u00bc \u03be V \u00fe \u03be R W\u03c3 R \u00f04:334\u00de where: \u03beV \u00bc \u03beV=mmn and \u03beR \u00bc \u03beR=mmn (see Fig. 4.35) Influence numbers sij may be determined using the stress reference value N and function value S: Influence numbers sij \u00bc S Nij sij describes the influence of the force Fnj exerted at point i on the stress at point i \u00f04:335\u00de Stress reference value Nij \u00bc 6yFj cos\u03b10 j s2Fni \u0394b Nij stress reference value for a tooth segment of width\u0394bat point i as a result of loading at point j \u00f04:336\u00de The stress reference value N includes the working point, lever arm h, force application angle \u03b1\u2019 (see Fig. 4.36) of the single force applied at point j, and the tooth thickness sFn in the normal section of the tooth root at each considered point i" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001837_j.robot.2020.103704-Figure8-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001837_j.robot.2020.103704-Figure8-1.png", + "caption": "Fig. 8. Dynamics structure diagram of LHDS.", + "texts": [ + " Dynamics analysis of LHDS model After obtaining the impedance based motion control compensation strategy of LHDS, to compensate the dynamics characteristics of LHDS, the dynamics analysis of LHDS model is carried out in this paper. Because the swing angle of HDU is very small relative to the leg component, and its center of mass slightly moves relative to the leg component, therefore, the influence bought by change of center of mass of HDU is neglected. The simplified LHDS model is a two connecting rods model, the dynamics structure diagram of LHDS is shown in Fig. 8. In Fig. 8, \u03b1 and \u03b2 are the deviation angle of the center of mass of two connecting rods relative to themselves respectively, and the direction of counterclockwise rotation is the positive direction. l1 is the distance from the knee joint to the ankle joint, p1 is the distance from the knee joint to the center of mass of the connecting rod OE. l2 is the distance from the ankle joint to t o r c d O i o f \u03c4 \u03c4 t he foot end, p2 is the distance from the ankle joint to the center f mass of the connecting rod EG" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002050_j.ijleo.2020.166055-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002050_j.ijleo.2020.166055-Figure1-1.png", + "caption": "Fig. 1. Architecture of the two-link robot manipulator.", + "texts": [ + " In general, the dynamic model of an n-link robot manipulator can be expressed as follows: M(q)q\u0308 + C(q, q\u0307)q\u0307 + G(q) + F(q, q\u0307, q\u0308) = \u03c4 (1) where q, q\u0307, and q\u0308 are the position, angular velocity, and angular acceleration of the manipulator joint angle, respectively ; M(q) is the positive definite moment of inertia of the n \u00d7 n joint; C(q, q\u0307) is the centripetal force of the n \u00d7 n order (Coriolis moment); G(q) is the inertia vector of the order n\u00d7 1; F(q, q\u0307, q\u0308) is the friction torque; \u03c4d is the unknown external disturbance; and \u03c4 is the control input torque of each joint. The architecture of the two-link manipulator is shown in Fig. 1. The dynamic characteristics of the robotic arm are as follows: (1) The inertial positive definite matrix is a bounded matrix. There are positive numbers a and b such that aI \u2264 M(q) \u2264 bI, where I is a positive definite identity matrix. (2) M(q) \u2212 2C(q, q\u0307) is an oblique symmetric matrix, xT(M(q) \u2212 2C(q, q\u0307) )x = 0, and x is a natural number. (3) Unknown friction and interference are modeled as ||(q, q\u0307, q\u0308) | | \u2264 F d, where F d is a positive real number. Assume that M(q),C(q, q\u0307), and G(q) are known and measurable state observations", + " Wu Optik 227 (2021) 166055 V\u0307(t) = 1 2 rT Mr\u0307 + 1 2 rT M\u0307r + \u2211n i=1 W\u0303 T i \u03b6i \u02d9\u0303Wi = \u2212 rT(M(q\u0308d + \u2137e\u0307) \u2212 Cr + C(q\u0307d + \u2137e) + G + Fr(q, q\u0307, q\u0308) \u2212 \u03c4 ) + \u2211n i=1 W\u0303 T i \u03b6i \u02d9\u0303Wi = \u2212 rT(Fr(q, q\u0307, q\u0308) \u2212 F\u0302r(q, q\u0307, q\u0308|W\u0302 + \u0393rr + M sgn(r) + \u2211n i=1 W\u0303 T i \u03b6i \u02d9\u0303Wi = \u2212 rT(F(q, q\u0307, q\u0308) \u2212 F\u0302 r(q, q\u0307, q\u0308|W\u0302 + F\u0302 r(q, q\u0307, q\u0308W\u2217) \u2212 F\u0302r(q, q\u0307, q\u0308W\u2217) + \u0393rr + M sgn(r) + \u2211n i=1 W\u0303 T i \u03b6i \u02d9\u0303Wi (31) By combining the adaptive law and the fuzzy approximation error in Eqs. (29)\u2013(30), we obtain V\u0307(t) = \u2212 rT \u0393rr \u2212 rT w \u2212 rTM sgn(r) + \u2211n i=1 W\u0303 T i \u03b6i \u02d9\u0303Wi \u2212 riW\u0303 T i H(q, q\u0307, q\u0308 = \u2212 rT \u0393rr (32) Since V\u0308(t) = \u2212 rT\u0393r r\u0307, V\u0308(t) is bounded, according to the Lyapunov stability theorem and Barbalat\u2019s lemma [31], it is implied that r will converge to zero as t \u2192 \u221e. This means that the sliding mode control system based on FRVFL is stable and convergent within the bounded range of the parameter estimation error. This study employed the two-link rigid manipulator shown in Fig. 1 to verify the effectiveness of the proposed algorithm. The experimental parameters are as follows [29]: M(q) = [ (m1 + m2)l2 1 + m2l2 2 + 2m2l1l2cos(q2) m2l2 2 + m2l1l2cos(q2) m2l2 2 + m2l1l2cos(q2) m2l2 2 ] G(q) = [ m2l2cos(q1 + q2) + (m1 + m2)l1gcos(q1) m2l2gcos(q1 + q2) ] C(q, q\u0307) = [ \u2212 m2l1l2sin(q2)q\u03072 \u2212 2m2l1l2sin(q2)q\u03071q\u03072 m2l1l2sin(q2)q\u03072 1 ] where m1 = m2 = 0.5 kg, l1 = 1 m, l2 = 0.8 m,and g = 9.8 m/s2. Let x = [q1, q\u03071, q2, q\u03072] T , y = [q1, q2] T ,and \u03c4U = [\u03c41, \u03c42] T; the outputs corresponding to the two joints of the control target are yd1 = 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001808_j.jsv.2020.115766-Figure11-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001808_j.jsv.2020.115766-Figure11-1.png", + "caption": "Fig. 11. Simulation results of disturbance moments about Y -axis.", + "texts": [], + "surrounding_texts": [ + "nal dimensions. Kistler 9255C can measure radial forces, i.e., in x and/or y directions, within the range of \u00b1 30 kN with a sensitivity of about \u22127 . 9 pC/N, whereas it can measure the axial force, i.e., in z direction, within the range of \u221210 to 60 kN with a sensitivity of about \u22123 . 9 pC/N. The dimensions of the table and locations of the forces sensors are shown in Fig. 13 . The reaction wheel along with Kistler 9255C, were mounted on an air bearing table to reduce the transmitted disturbances from the ground. The air bearing table is shown in Fig. 14 and it has an area of 1.8 \u00d7 1.6 m 2 . The efficiency of the vibration isolation is more than 80% at a frequency of 5 Hz. The reaction wheel was controlled to spin at different rotational speeds. It spun at speeds that range from 0 to 60 0 0 rpm with a step of 50 rpm. The force sensors\u2019 measurements at each speed were logged via the 24-bit data acquisition (LMS SC316-UTP). Using this extracted data, the total disturbance forces in three directions and the total disturbance moments about the radial axis can be determined using Eq. (35) . F x = F x 1 + F x 2 + F x 3 + F x 4 F y = F y 1 + F y 2 + F y 3 + F y 4 F z = F z1 + F z2 + F z3 + F z4 M x = l y (F z1 \u2212 F z2 \u2212 F z3 + F z4 ) M y = l x (\u2212F z1 \u2212 F z2 + F z3 + F z4 ) (35) H. Alkomy and J. Shan Journal of Sound and Vibration xxx (xxxx) xxx H. Alkomy and J. Shan Journal of Sound and Vibration xxx (xxxx) xxx H. Alkomy and J. Shan Journal of Sound and Vibration xxx (xxxx) xxx H. Alkomy and J. Shan Journal of Sound and Vibration xxx (xxxx) xxx H. Alkomy and J. Shan Journal of Sound and Vibration xxx (xxxx) xxx H. Alkomy and J. Shan Journal of Sound and Vibration xxx (xxxx) xxx" + ] + }, + { + "image_filename": "designv10_14_0000750_j.mechmachtheory.2018.11.002-Figure8-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000750_j.mechmachtheory.2018.11.002-Figure8-1.png", + "caption": "Fig. 8. Surface deviations for: (a) Case 1, and (b) Case 2.", + "texts": [ + " Five different cases of design are considered to investigate the influence of the design parameters of the circular cutters on the surface deviations of the generated face gear. The design parameters taken into account for the circular cutters for Cases 1 to 5 are listed in Table 2 . A non-listed reference geometry, Case 0, corresponds to the standard face gear generated by a shaper with the same number of teeth of the pinion and with no profile modifications. All deviations shown below for cases 1 to 5 will be referred to this geometry. Fig. 8 (a) shows the surface deviations between the tooth surface of face gear for Case 1 and the tooth surface of the reference geometry, Case 0. The reference tooth surface is always represented as a flat surface and the surface deviations of the various face gear geometries represented normal to it. Positive deviations mean that the tooth surface is outside the reference geometry (more material left during generation) and negative deviations mean that the surface of the face gear is inside the surface of the reference geometry (more material cut during generation). When the cutter profile tilt angle is 25 \u00b0 as for Case 1, top corners of the face gear tooth surface show negative deviations ( \u221252.31 \u03bcm and \u221232.28 \u03bcm). However, deviations at the bottom boundary of the active surface show positive and not desired deviations of 26.49 \u03bcm and 27.68 \u03bcm at the inner and outer section, respectively, that will cause interference with the pinion tooth surfaces. As a result, the cutter profile tilt angle needs to be increased. Fig. 8 (b) shows the surface deviations for Case 2 for which the cutter profile tilt angle has been increased to 35 \u00b0, and the cutter radius has been kept equal to that of Case 1. Considering the new cutter profile tilt angle of 35 \u00b0, the deviations at the bottom boundary of the active surface of the face gear have been reduced to \u22124.0 \u03bcm and \u22122.33 \u03bcm at the inner and outer sections, respectively. However, deviations at the top edge of the face gear tooth surfaces have been increased to \u221278.06 \u03bcm at the inner section and to \u221260", + " Case 0 represents the tooth of an standard face gear. The face gear tooth surface has standard surface and because the generating shaper has the same number of teeth as the pinion, the contact pattern covers the whole surface. Pinion and face gear are in line contact at any moment. For Case 1, in which a cutter profile tilt angle \u03b4 of 25 \u00b0 has been used, the contact pattern is not continuous because of the large positive deviations on the bottom boundary of the active surface of the face gear tooth surface at the outer section (shown in Fig. 8 (a)). A severe interference in the process of meshing is detected and therefore this design is completely unsatisfactory. When a cutter profile tilt angle of 35 \u00b0 is used for Case 2, the contact path is centered at the tooth surface and directed in profile direction. The successive contact ellipses get shorter gradually from the bottom to the top boundaries of the tooth surface. This is due to the fact that the effective longitudinal crowning is increased at the top boundary of the surface with respect to the bottom boundary" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002061_tec.2020.3048442-Figure21-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002061_tec.2020.3048442-Figure21-1.png", + "caption": "Fig. 21. Diagrams of voltage and current vectors with different field excitations. (a) Negative field excitation. (b) Positive field excitation.", + "texts": [ + " Authorized licensed use limited to: Western Sydney University. Downloaded on June 14,2021 at 20:12:10 UTC from IEEE Xplore. Restrictions apply. 0885-8969 (c) 2020 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. > REPLACE THIS LINE WITH YOUR PAPER IDENTIFICATION NUMBER (DOUBLE-CLICK HERE TO EDIT) < 9 The diagrams of voltage and current vectors with different field excitations are illustrated in Fig. 21. As can be seen, the angle between current vector and voltage vector varies with different field excitations. With positive field excitation, the magnetic field is enhanced and the angle between current vector and voltage vector is reduced. Therefore, the power factor is improved with flux-enhancing field current, and reduced under flux-weakening field current. Based on the voltage vector and current vector diagram, the power factor of the DSHE machines are calculated as shown in Fig. 16. As predicted, the power factor can be improved with positive field excitation" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000917_imece2015-52386-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000917_imece2015-52386-Figure1-1.png", + "caption": "Figure 1. (a) Schematic of the geometry of the Vickers indentation tests, and (b) Schematic of the nominal projected contact area Anom", + "texts": [ + " Then, the residual stress and microstructure in microstructure scale of the samples were investigated. Based on theoretically and numerically analysis, Carlsson and Larsson [13, 14] indicated that residual strain fields can be accurately correlated with the hardness value, while residual stresses can be related to the size of the contact area. In the Vickers test, it is assumed that elastic recovery does not occur once the load is removed. However, elastic recovery does occur, and sometimes its influence is quite pronounced. As shown in the Figure 1, the impression shape is distorted due to elastic recovery, which is very common in anisotropic materials. The hardness value is accurately given by the formula \ud835\udc3b\ud835\udc3b = \ud835\udc36\ud835\udc36\ud835\udf0e\ud835\udf0e\ud835\udc4c\ud835\udc4c (1) where C is a constant, dependent on the geometry of the sharp indenter only, and \ud835\udf0e\ud835\udf0e\ud835\udc4c\ud835\udc4c is the yield strength of the material. For a strain-hardening material \ud835\udf0e\ud835\udf0e\ud835\udc4c\ud835\udc4c is replaced by \ud835\udf0e\ud835\udf0e\ud835\udc5f\ud835\udc5f\ud835\udc5f\ud835\udc5f\ud835\udc5d\ud835\udc5d\ud835\udc5f\ud835\udc5f being the flow stress at a representative value of plastic strain, \ud835\udf00\ud835\udf00\ud835\udc5f\ud835\udc5f\ud835\udc5f\ud835\udc5f\ud835\udc5d\ud835\udc5d\ud835\udc5f\ud835\udc5f. \ud835\udc3b\ud835\udc3b = \ud835\udc36\ud835\udc36\ud835\udf0e\ud835\udf0e\ufffd\ud835\udf00\ud835\udf00\ud835\udc5f\ud835\udc5f\ud835\udc5f\ud835\udc5f\ud835\udc5d\ud835\udc5d\ud835\udc5f\ud835\udc5f\ufffd (2) It could be expected that the change of the hardness value due to a residual plastic strain can be accurately described by the equation above if the representative strain in those 2 Copyright \u00a9 2015 by ASME Downloaded From: http://proceedings" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002357_j.jfranklin.2021.02.011-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002357_j.jfranklin.2021.02.011-Figure1-1.png", + "caption": "Fig. 1. Definitions of vectors and frames.", + "texts": [ + " The skew-symmetric operator converts arbitrary real vector a = [ a 1 a 2 a 3 ] to a skew- ymmetric matrix as a \u00d7 = [ 0 \u2212 a 3 a 2 ; a 3 0 \u2212 a 1 ; \u2212a 2 a 1 0 ] . The diagonal operator diag n i=1 (c i ) = diag (c 1 , \u00b7 \u00b7 \u00b7 , c n ) arranges the generalized elements c i nto a generalized diagonal matrix, where c i can be a scalar or a matrix; the column-vector perator vc n i=1 (d i ) = [ d 1 \u00b7 \u00b7 \u00b7 d n ] arranges the generalized elements d i into a column vector, here d i can be a scalar or a row-vector. The definition of all frames and vectors in Fig. 1 can be referred in [40] , so can be the irection cosine matrix (DCM) R NT , which describes a transformation from the frame T OXYZ o the frame N . OXYZ 2 2 o\u23a7\u23a8 \u23a9 w l R s E R t{ A t\u23a7\u23a8 \u23a9 w E .2. Kinematics and dynamics model .2.1. Attitude maneuver model By considering the active spacecraft as a rigid body, the kinematics and dynamics model f the active spacecraft for attitude control can be described as \u02d9 qb = E ( q b ) \u03c9 b \u00a8\u0302 qb = \u2032 at t i + E b2 ( J \u22121 b \u03c4 + a ) + E b2 J \u22121 b0 \u03c4 \u02d9 \u03c9 b = \u2212J \u22121 b0 \u03c9 \u00d7 b J b0 \u03c9 b + a + J \u22121 b \u03c4 + J \u22121 b0 \u03c4 (1) here q b \u2208 R 4 is a quaternion defined by Section 2" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000194_s12206-015-0833-3-Figure7-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000194_s12206-015-0833-3-Figure7-1.png", + "caption": "Fig. 7. Solid model of a four contact-point ball bearing connection.", + "texts": [ + " The elastoplastic performance of material is described by RambergOsgood model as Eq. (14). The material parameters of each layer are listed in Table 2 [12]. Young's modulus of elasticity, Poisson's ratio of ball\u2019s material is 2.10\u00d7105 MPa and 0.3, respectively. 1/ ' , ' n a a a E K s ee \u00e6 \u00f6= + \u00e7 \u00f7 \u00e8 \u00f8 (14) where E is the modulus of elasticity, K' is the cyclic strength coefficient, n' is the cyclic strain hardening exponent. The modeling and calculation in this study are done by means of the commercial finite element software ANSYS. Fig. 7 shows the complete solid model of a four contact-point ball bearing connection. The complete solid model includes a support structure, a four contact-point ball bearing, a load structure and some bolts. The support structure is connected with the bearing\u2019s outer ring by many bolts. The load structure is connected with the bearing\u2019s inner ring by many bolts. The end face of the support structure is fixed. The end face of the load structure is applied with combined loads (radial force, axial force and overturning moment)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001179_j.ijsolstr.2018.11.031-Figure5-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001179_j.ijsolstr.2018.11.031-Figure5-1.png", + "caption": "Figure 5: (Color online) The 3D bifurcation bifurcation diagrams based on the semi-analytical solution.", + "texts": [ + "15 are 14, 6, 4, 3, respectively. It can be seen that the critical growth factor varies in a relatively small region, which agrees well with the asymptotic analysis by Jin et al. (2018). However, a slight increase of g\u0302 would cause a large increase of |A1| for larger H\u0302. Therefore, a thicker inner layer produces a larger wrinkled amplitude, and further a large fold. After verifying the accuracy of the proposed semi-analytical approach, we further show the three-dimensional (3D) bifurcation diagrams in Fig. 5. The evolution of the amplitude with same geometry but different modulus ratios are depicted in Fig. 5(a) where \u03be varies from 2 to 80. Meanwhile, we plot the evolution surface with the same modulus ratio (\u03be = 10) but different inner layer thicknesses in Fig. 5(b) where H\u0302 is from 0.02 to 0.14. Different critical mode number corresponds to different color. Since the mode number is always an integer, a stair-like profile is observed if g\u0302 is specified, and the step has a big variation when nc changes. This implies that the critical mode is the main factor enlarging the amplitude. Similarly, the intersection curve in \u03be \u2212 g\u0302 or \u03be \u2212 H\u0302 plane denotes the critical growth factor. It can be seen that |A1| is also an increasing function of \u03be and H\u0302 if the growth factor g\u0302 is fixed. In order to illustrate the conclusion clearer, we plot the post-buckling patterns for three typical cases in Fig. 5(a) and Fig. 5(b) in Fig. 6. The corresponding geometrical and material parameters together with the critical mode numbers are also shown. It can be seen that a lower critical mode corresponds to a higher amplitude. ACCEPTED M ANUSCRIP T (a) \u03be = 5, nc = 16. (b) \u03be = 20, nc = 12. (c) \u03be = 40, nc = 10. (d) H\u0302 = 0.04, nc = 7. (e) H\u0302 = 0.08, nc = 4. (f) H\u0302 = 0.12, nc = 3. Figure 6: Post-buckling patterns based on the semi-analytical solution when the total growth g\u0302 is fixed by 1.2. The solid lines are the inner, outer and interfacial radii" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001637_s11044-019-09705-0-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001637_s11044-019-09705-0-Figure3-1.png", + "caption": "Fig. 3 Definition of the moving plate parameters", + "texts": [ + " The manipulator consists of a moving equilateral triangular platform of side length h connected to a fixed equilateral triangular base of side length d by three limbs. Each limb consists of two links; the first link is connected to the ground by means of a revolute joint identified by the letter Bi and is actuated by a rotary actuator whereas the second link connects the first link to the moving platform with two passive joints at Ai and Ci . The manipulator has three degrees of freedom, therefore, three actuators, one for each limb, are needed to control the moving platform. The parameters of the moving plate are shown in Fig. 3. To describe the motion of the moving platform, two coordinate systems are defined. A fixed coordinate system attached to fixed platform with origin O and axes x and y, and is called the reference frame, the origin O is located at the centroid of the fixed triangular base. The second coordinate system is attached to the moving platform (with origin O \u2032 and axes x \u2032 and y \u2032), the origin O \u2032 is located at the centroid of the moving triangle. The coordinates of the points of connection of the manipulator with the fixed base are: B1(\u2212300,\u2212173" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000820_s11044-014-9445-4-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000820_s11044-014-9445-4-Figure1-1.png", + "caption": "Fig. 1 Passive rimless wheel model", + "texts": [ + " Section 3 introduces a 1-DOF active RW model for analysis and describes the linearized equation of motion and quadratic approximation of mechanical energy. Section 4 derives the analytical transition function of the state error for the stance phase and discusses its accuracy through numerical simulations. Section 5 extends the analysis to an underactuated RW model and describes the method in general cases. Finally, Sect. 6 concludes this paper and describes future research direction. 2.1 Passive rimless wheel model In this section, we consider the RW model shown in Fig. 1. This model is planar and eightlegged. The radius or leg length is l [m] and the total mass is m [kg]. The angle between two adjacent leg frames is \u03b1 and is \u03c0/4 [rad]. We assume that the model does not have inertia moment and the contact point with the ground does not slip during motion. Let \u03b8 [rad] be the angular position of the stance leg with respect to vertical. The kinetic energy, K [J], and the potential energy, P [J], then become K(\u03b8\u0307) = 1 2 ml2\u03b8\u03072, P (\u03b8) = mgl cos \u03b8. (1) Following Lagrange\u2019s method, the equation of motion becomes ml2\u03b8\u0308 \u2212 mgl sin \u03b8 = 0", + " (19) Therefore, we can find that the transition of the state error during the stance and collision phases are identical and are cos\u03b1. In the subsequent sections, we will investigate this result in more detail from the mechanical energy point of view. 3.1 Active combined rimless wheel model and linearization of motion As the realistic model of an active 1-DOF limit-cycle walker, we consider an active combined rimless wheel (CRW) shown in Fig. 2 [15]. This is composed of two identical eightlegged RWs of Fig. 1 and a body frame, and can exert a joint torque, u [N \u00b7 m], between the rear stance-leg and the body frame. We assume the following statements: \u2013 The fore and rear stance legs always contact with the ground without sliding. \u2013 The inertia moments about the CoMs of all the frames can be neglected. \u2013 The rear and fore RWs perfectly synchronize or rotate maintaining the relation \u03b8r \u2261 \u03b8f . The 3-DOF CRW with the hard ground configures a four-bar linkage, and exerting the joint torque, u, is thus equivalent to exerting that at the contact point with the ground (ankle-joint torque)", + " This problem on numerical calculation arises in a fast convergent gait and is discussed again in the next section. This section discusses the discrete behavior of HZD of an underactuated rimless wheel (URW) with a torso in the same manner as the previous sections. The main purpose is to specify the analysis method with general formulae. 5.1 Model of underactuated rimless wheel and its linear approximate equation of motion Figure 11 shows the model of an URW with a torso [16]. This URW consists of an eightlegged symmetrically-shaped RW of Fig. 1 and a torso link. The torso link is connected to the RW at the central position and the moment of inertia about the joint is I [kg \u00b7 m2]. Let \u03b8 = [\u03b81 \u03b82]T be the generalized coordinate vector. The equation of motion then becomes [ Ml2 0 0 I ][ \u03b8\u03081 \u03b8\u03082 ] + [\u2212Mgl sin \u03b81 0 ] = [ 1 \u22121 ] u. (72) By linearizing Eq. (72) about \u03b8 = \u03b8\u0307 = 02\u00d71, we get[ Ml2 0 0 I ][ \u03b8\u03081 \u03b8\u03082 ] + [\u2212Mgl 0 0 0 ][ \u03b81 \u03b82 ] = [ 1 \u22121 ] u. (73) We denote Eq. (73) as M0\u03b8\u0308 + G0\u03b8 = Su. (74) Next, we outline the collision dynamics. We assume the following: \u2013 The URW falls down as a 1-DOF rigid body or achieves the condition of \u03b8\u03071 = \u03b8\u03072 immediately before the next impact" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002361_j.matchar.2021.111019-Figure5-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002361_j.matchar.2021.111019-Figure5-1.png", + "caption": "Fig. 5. Schematic diagram illustrating the evolution of precipitates morphology during long term thermal cycling.", + "texts": [ + " The content of precipitates gradually increased from the top to the bottom regions for the sample under the same parameters. The morphology and size of precipitates varied significantly between different regions suggesting that there was a significant difference in thermal history between different layers. Driven by capillary forces, the morphology of precipitates upon elevated temperature exposure will change and eventually lead to spherical dispersoid shape [39]. The evolution in precipitates morphology was schematically depicted in Fig. 5. For rod-like precipitates with a high aspect ratio, the shape was unstable with respect to the development and growth of perturbations. When the system was heated sufficiently to allow atomic migration, thermal grooves were formed in order to establish static equilibrium among surface tensions [39,40]. The formed grooves were surrounded by two ridges, and the curvature of these ridges promoted the mass transport away from the groove, further promoting the deepening of the grooves to achieve the re-establishment of tension balance" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000728_j.ijleo.2018.09.079-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000728_j.ijleo.2018.09.079-Figure3-1.png", + "caption": "Fig. 3. 3D solid model of the porthole performance die.", + "texts": [ + " Surfaces of the profiles were examined after extrusion, anodizing and static powder painting. Samples from profiles were cut, cold molded, polished, and etched to reveal the profile microstructure and anodic coating cross section. Characterization was carried out using optical microscope and scanning electron microscope. In the second stage of this study, design, production, and performance tests of the porthole die for a standard handrail profile was carried out. Solid model of the die is given in Fig. 3. The main purpose of this stage was to demonstrate the performance of the AM die under industrial conditions of use. Production of the porthole die was completed in 107 h. Following production, aging and stress reduction was carried out by annealing the production platform and the die at 490 \u00b0C for 8 h. The surface hardness of the die was measured both after production and stress relieving (Table 6). Once again, no conventional finishing and surface hardening post-processes were applied to the die" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001872_tmag.2019.2947611-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001872_tmag.2019.2947611-Figure2-1.png", + "caption": "Fig. 2. Rotor structure for improving the flux weakening ability. (a) tapered flux barrier, (b) segmented magnet pole", + "texts": [ + " The PM volume is optimized to satisfy the condition that the minimum d-axis flux linkage is close to zero, then the PMSM can offer a wide constant-power operation range as well as high-efficiency operation in the constant-power operation region [5-6]. The PM volume reduction means low output torque in the constant torque region. It is more applicable to optimize the rotor structure to adjust the inductance of d-axis. The angle and iron bridge between two pieces of magnets in the V shape rotor structure are investigated to obtain the higher inductance of d-axis [7]. As shown in Fig.2(a), the multiple flux barriers PMSM has been evaluated and the outer layer barrier is tapered to improve the flux-weakening ability [8]. The tapered flux barriers provide more paths for the armature field to weaken the field of PM. The segmented magnet pole technic shown in Fig.2(b) is adopted and there exists an iron bridge between every two PM segments [9-10]. As a result, the paths for the armature field are provided and the wide constant power speed ratio is achievable. Meanwhile, the paths for the armature field result in excessive PM flux leakage and decreasing saliency, and the ability generating output torque is decreased. T Corresponding author: Mingyao Lin (E-mail: mylin@seu.edu.cn). Color versions of one or more of the figures in this paper are available online at http://ieeexplore" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001285_s12541-019-00226-6-Figure13-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001285_s12541-019-00226-6-Figure13-1.png", + "caption": "Fig. 13 Contour plot of the temperature field with addition of the bead element at different times", + "texts": [ + " These results indicate that the bead height is significantly equalized with the controlled scanning-speed. To investigate the variations in the temperature and residual stress due to scanning speed changes during the DED process, the FE analysis was carried out. In the FE simulation, two scanning speed conditions were considered: one corresponds to the estimated scanning speed condition of Fig.\u00a011 in which scanning speed decreases from 900 to 500\u00a0mm/s at the corner section, the other is a constant scanning speed condition at 900\u00a0mm/s for both straight and corner sections. Figure\u00a013 shows the contour plot of the temperature field. Seven beads are added on the substrate according to the scanning speed. The color was set as grey for the 1 3 temperatures higher than the melting point temperature (1725\u00a0K) of H13 steel; in other words, the grey region represents the molten pool area. The maximum temperature was observed at the corner, and the temperature history was measured at the measuring point, as shown in Fig.\u00a013. The temperature history at the first and fourth beads is presented in Fig.\u00a014. The temperature increases sharply and decreases after heat source passing the scanning track. The maximum temperature of the fourth bead is larger (2484\u00a0K) than that of the first bead (2289\u00a0K). This result occurs because as the scanning speed decreases at the corner section corresponding with the fourth bead, a higher heat input is induced, which leads to a higher maximum temperature of the fourth bead. When the scanning speed decreases at the corner section, the maximum temperature is 2489\u00a0K, but the maximum temperature decreases to 2369\u00a0K at a constant scanning speed, as shown in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000769_978-3-662-43893-0-Figure6.1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000769_978-3-662-43893-0-Figure6.1-1.png", + "caption": "Fig. 6.1 Principle of straight bevel gear milling using intermeshing cutters", + "texts": [ + " Typical for all three methods are: tooth shape: tapered depth manufacturing: single-indexing method; both pinion and wheel are generated tool: two intermeshing cutters with radial blades profile crowning: by machine kinematics lengthwise crowning: curved blade paths The axes of two disc-shaped cutters, one for the left tooth flank, one for the right tooth flank, are at an angle to each other, causing the blades to intermesh alternately in such a way that their primary cutting edges form a trapezoidal profile. Since the cutting edges do not lie precisely in the plane of rotation, but at a small inclination angle (see Fig. 6.1), the teeth receive a fixed lengthwise crowning and the bottom of the tooth slot is not straight but arc-shaped. The angle between the cutters is given by the flank angles of the blades and the tapered slot width of the bevel gear. 234 6 Manufacturing Process The broaching process known as Revacycle\u00ae is the most productive cutting process to manufacture straight bevel gears. As a specialized tool is needed for each transmission ratio, it is suited only to mass production. The tool is disc-shaped in diameters of approximately 530\u2013635 mm" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000983_s11071-016-3155-9-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000983_s11071-016-3155-9-Figure2-1.png", + "caption": "Fig. 2 System diagram", + "texts": [ + " The main factors affected the value of Lyapunov exponent can be derived by Fig. 1, which includes structural parameters (e.g., L), the mass of system (m), rotational inertia (Ix , Iy, Iz) and the initial value (x0). Therefore, the main factors affected the stability of quadrotor unmanned aerial vehicles are L ,m, Ix , Iy, Iz and x0. 3 Methodology 3.1 Dynamics model Assuming the quadrotor unmanned aerial vehicle is a rigid body and four propeller axes are all perpendicular to the surface of vehicle body (see Fig. 2). B(x, y, z) represents body-fixed coordinate frame and E(X,Y , Z) indicates the earth-fixed coordinate frame. Four situations exist with the different value of F1, F2, F3 and F4, as follows. When F1 = F2 = F3 = F4, the quadrotor unmanned aerial vehicle stays the state of rising, descending or hovering; When F2 = F4 and F1 = F3, the system stays the state of pitching; When F1 = F3 and F2 = F4, the system stays the state of rolling; When F1 = F3 = F2 = F4, the system stays the state of yawing. The dynamics model of quadrotor unmanned aerial vehicles based on Euler\u2013Poincare equation is written as q\u0307 = V (q)p (2) M(q) p\u0307 + C(q, p)p + F(p, q, u) = 0 (3) where V (q), M(q) andC(q, p) are kinematics matrix, inertia matrix and gyroscopic matrix, respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001304_j.mechmachtheory.2019.103697-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001304_j.mechmachtheory.2019.103697-Figure3-1.png", + "caption": "Fig. 3. Mechanism diagram of the force-sensitive element.", + "texts": [ + " o 1 denotes the projection of point o in the plane formed by A O i ( i =4 , 5 , 6 ) and a o i ( i =4 , 5 , 6 ) . The distance from A O i ( i =1 , 2 , 3 ) and a o i ( i =1 , 2 , 3 ) to the axis Z is r 1 . The distance from A O i ( i =1 , 2 , 3 ) to the plane XOY is l 2 . The distance from A O i ( i =4 , 5 , 6 ) to the axis Z is r 2 . The distance from a o i ( i =4 , 5 , 6 ) to the axis z is r 3 . The angle between the line A O 4 o 1 and the plane XOZ is \u03b1. The angle between the line a o 4 o 1 and the plane xoz is \u03b2 . A o 4 a o 4 o 1 denotes a right triangle and cos ( \u03b1 \u2212 \u03b2) is equal to r 3 / r 2 . Fig. 3 is the mechanism diagram of the force-sensitive element. The cross-section of the force-sensitive element is a square with side length a . The length of the flexible rod between two flexible spherical joints is l 1 . The arc radius of the flexible spherical joints is r . The minimum cross-section diameter of flexible spherical joints is t . The distance between the end of the flexible spherical joints and the platform is l 4 . The arc center angle of the flexible spherical joints is 2 \u03b8m . When the configuration is used as a sensor structure, the strain gauges are attached to the middle surface of the force sensitive element in order to obtaining the axial stress of the six flexible rods" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000423_0954405414553979-Figure12-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000423_0954405414553979-Figure12-1.png", + "caption": "Figure 12. Centrifugal impeller and blade surface.", + "texts": [ + " Compute dpi ,SAE(w 0), i=1, . . . , n3 Step1: 1. Solve the linear programming problem LP to determine the differential increment of the surface shape and tool shape parameters Dw 2. Update wk+1 =wk +Dw 3. Compute dpi ,SAE(w k+1), i=1, . . . , n3 If j1 (rk0=r k+1 0 )j\\ z or j1 (uk=uk+1)j\\ z or dpi ,SAE(w k+1)\\ 0, exit and report w* =wk+1, else set k= k+1 and return to Step1 of (1). In order to demonstrate the validity of the proposed method, we give examples of flank milling of free-form impeller blades. As illustrated in Figure 12, the centrifugal impeller has 30 blades. Two cutters are applied. One is a ball-end cylindrical cutter, and its initial radius r0 is chosen as 3.75mm. The other is a ball-end conical cutter, and its initial bottom radius r0 and half taper angle u are chosen as 3mm and 58, respectively. The height of the two cutters is H=20mm. According to the accuracy requirement, the prescribed tolerance d is 0.02mm. Besides, the ball end of the cutter should be tangential to the hub surface during the machining process" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000553_1.4034318-Figure7-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000553_1.4034318-Figure7-1.png", + "caption": "Fig. 7 (a) A sequence of snap shots of the tennis racket: t 5 0.0, 1.4, 1.8, 3.0, and 4.0 s. (b) The components of the angular momentum vector about the inertial coordinate axes and its magnitude, and (c) about the body-attached coordinate axes and its magnitude.", + "texts": [ + "org/about-asme/terms-of-use m\u00bc 0.375 kg, J1C \u00bc 0.0185 kg/m2, J2C \u00bc 0.0164 kg/m2, and J3C \u00bc 0.00121 kg/m2. We only present the result when the initial spin is applied with the e2-axis in Fig. 4(d). The initial angular velocities are: x2\u00f00\u00de \u00bc 5.0 rad/s and x1\u00f00\u00de \u00bc x3\u00f00\u00de \u00bc 0:001 rad/s. The vertical velocity of 24.5 m/s is applied at the center of mass, which gives the racket air time of 5 s. The initial value gives d \u00bc D=J2C ffi 1. An inertial coordinate frame is selected to be the body-attached frame at t\u00bc 0: eI \u00bc e\u00f00\u00de. Figure 7(a) shows a sequence of snap shots of the tennis racket in the first row at t\u00bc 0, and in the second row from the leftmost figure at t\u00bc 1.4, 1.8, 3.0, and 4.0 s [22]. Figure 7(b) illustrates the inertial components of the angular velocity as well as the magnitude of the angular velocity vector. The x2 -component and the magnitude of the vector display the common value of 0.082 kg m2/s (black lines). The angular velocity components in the x3 - and x1-directions remain zero (blue lines). Figure 7(c) presents the angular momentum about the bodyattached coordinate axes. The magnitude of the angular momentum vector (black line total) retains the same value of 0:082 kg m2=s . The accuracy of the numerical solutions was checked both geometrically and analytically. To compare the momentum\u2019s trajectory with the geometrically exact trajectory, the angular momentum vector was plotted on the angular momentum sphere, as shown in Fig. 5(a). This geometrically exact trajectory is the intersection between the energy ellipsoid and the angular momentum sphere" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000925_s11044-016-9500-4-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000925_s11044-016-9500-4-Figure1-1.png", + "caption": "Fig. 1 Schematic diagram of a 6-legged walking machine", + "texts": [ + " 4, we discuss the commonalities and differences between applications of the weighted Moore\u2013Penrose generalized inverse in the force analysis of redundantly actuated PMs and in the analysis of passive overconstrained PMs. Finally, some concluding remarks are summarized in Sect. 5. In this section, the explicit relationship between the weighted Moore\u2013Penrose generalized inverse and the optimal driving force/torque distribution of redundantly actuated PMs is revealed on the basis of [13, 14]. A 6-legged walking machine [13], as shown in Fig. 1, whose main body is supported by at least three legs while moving, can be regarded as a PM composed of three or more supporting limbs (legs), a moving platform (the main body) and a base (the ground). Each leg consists of three links connected to each other and to the main body by revolute joints. The contacting relation between each foot and the ground can be regarded as a spherical joint. Then, the joints of each leg from the ground to the main body are denoted as joints 1, 2, 3, and 4, respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000846_tie.2014.2300060-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000846_tie.2014.2300060-Figure3-1.png", + "caption": "Fig. 3. (a) Kinematic structure of the robot in the sagittal plane. Associated joint frames are indicated. \u03b8f denotes the pitch axis body tray orientation. (b) Actual robot representation in the sagittal plane. Only pitch axis ankle and knee joints have passive compliant elements (green joints). (c) Whole system can be considered as a 3-DOF robot with elastic joints.", + "texts": [ + " In order to gain insights regarding the base resonance frequency of the system, a computational analysis is conducted, prior to the experimental identification procedure. Even though we utilize the experimentally identified base resonance frequency value during locomotion generation, a computational analysis is of importance for two reasons: 1) to double-confirm the experimental identification results; and 2) to see whether the elasticity properties of the system would yield a physically realizable base resonance frequency, prior to the identification procedure. Keeping these in mind, the robot may be modeled as a 3-DOF system with elastic joints. Fig. 3 provides an illustration for this model. Considering this case, the system dynamics can be expressed using Spong\u2019s formulation [28] \u03c4 = Ir(q)q\u0308 + Cr(q, q\u0307)q\u0307 +Gr(q)\u2212 \u03c1q\u0307 (3) = Im\u03b8\u0308 +Bm\u03b8\u0307 + \u03ba(\u03b8 \u2212 q). (4) In (3) and (4), \u03c4 is the generalized joint torque vector, while q and \u03b8, respectively, denote link side and motor side angular position vectors. Ir(q), Cr(q, q\u0307), Gr(q) are inertia, coriolis/centrifugal, and gravitational force terms, regarding the reduced model presented in Fig. 3(c). \u03c1 is a diagonal matrix that represents link side friction terms, and it may be identified via the method proposed in [29]. Im and Bm are diagonal matrices that store rotor inertia and friction constants, which are appropriately scaled via respective gear ratios. \u03ba refers to the diagonal joint stiffness matrix. Referring to [30], the natural frequencies of this system are determined, while generalized coordinates q converge to the equilibrium configuration that is defined as qe = \u03b8. That being said, we utilize Jacobian of the gravitational force term \u03a8Gr(q) that can be computed via partial differentiation with respect to joint motion (\u03a8Gr(q) = \u2202Gr(q)/\u2202q)", + " Finally, the horizontal foot trajectory xf can be computed as follows to make sure that necessary foot placement planning can be achieved: xf = x\u2212 (z \u2212 zf ) tan\u03c8. (34) Having computed all of the necessary components of CoM trajectories, we now proceed to the joint computation scheme. As we stated previously, only the joints in the sagittal plane are equipped with passive compliant elements. Therefore, we focus on jumping motion in which the pitch axis ankle, knee, and hip joints are actuated, i.e., 3 DOF. Fig. 3(a) displays the kinematic structure and joint frames. At this point, we can easily derive an analytical inverse kinematics solution that maps CoM position and upper body pitch axis orientation (x, z, \u03b8f ) with joint positions (q1, q2, q3). On the other hand, this approach cannot characterize the rotational inertia, an important characteristic of the robot dynamics [5]. Considering this fact, we utilized x-axis and z-axis CoM translational momenta (Px, Pz) and pitch axis CoM angular momentum (Ly) for the joint motion computation task, in accordance with the method proposed in [27]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001624_icra.2019.8793606-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001624_icra.2019.8793606-Figure1-1.png", + "caption": "Fig. 1. The Vogi platform, a passively-coupled tilt-rotor VTOL aircraft", + "texts": [ + " Many different designs have been proposed, including tailsitters [2][3][4], tilt rotors [5][6], and tilt wings [7][8]. The concept of a passively coupled tilt-rotor [9] provides several advantages over other types of VTOL aircraft: this vehicle is more immune to cross winds, which have a large impact on tilt-wing or tailsitter aircraft. As well, it does not require tilting actuators that are needed in tiltrotors or tilt-wings [10]. Furthermore, conventional tilt rotors can suffer from rotor to wing aerodynamic interference [11], which is avoided on the current platform. The prototype called Vogi, shown in Fig. 1, is composed of a quadrotor passively coupled to an aircraft about the pitch axis via a parallelogram mechanism. The aircraft is equipped with a bidirectional rotor at the rear to control its pitch angle at low speed as well as elevator and ailerons to control its pitch and roll at high speed. Due to the passively coupled mechanism, the transition is smooth and continuous. The aircraft has a wingspan of about a meter and a takeoff weight of about a kilogram. Because this vehicle is an articulated multibody system, the kinematics is challenging to derive" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002018_s40815-020-00949-z-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002018_s40815-020-00949-z-Figure2-1.png", + "caption": "Fig. 2 UWR\u2019s leg model", + "texts": [ + " 4, we take simulation studies to demonstrate the effectiveness of the proposed methods. The conclusion of this research is given in Sects. 5. Notations used in this study are defined in Table 1 Since the UWR needs to work on the rough underwater terrains, structural reliability and hydraulic property should be considered in the design process of UWR. Therefore, an UWR inspired by the crab is presented in Fig. 1. To walk underwater, each leg of UWR is required to track its own desired trajectory. One effective design of the UWR\u2019s legs is shown in Fig. 2. The dynamics of UWR\u2019s leg with p links can be described by Euler\u2013Lagrange equation as follow M q\u00f0 \u00de\u20acq\u00fe C q; _q\u00f0 \u00de _q\u00fe g q\u00f0 \u00de \u00bc s\u00fe x \u00f01\u00de where q; _q; \u20acq 2 Rp denote the generalized coordinate, velocity, and acceleration, respectively, M q\u00f0 \u00de 2 Rp p represents the symmetric positive definite inertia matrix, C q; _q\u00f0 \u00de 2 Rp p represents the Coriolis and centrifugal matrix, g q\u00f0 \u00de 2 Rp represents the gravitational force, s 2 Rp is the control input, and x 2 Rp is the external disturbance (including environment disturbance and unmodeled dynamics)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001850_lra.2020.3045646-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001850_lra.2020.3045646-Figure1-1.png", + "caption": "Fig. 1. Representation of the redundancy as the arm angle \u03c8 between the real elbowE and the virtual elbowEv in the reference plane, with the line connecting the shoulder S and wrist W as the rotation axis.", + "texts": [ + " These methods generalize well to different kinematic structures and variable degrees of redundancy. They are also able to enforce limits on joint position, velocity and acceleration. On the contrary, closed-form IK solutions only exist for few redundant robot types, like the popular non-offset 7-DoF anthropomorphic manipulator structure. The analytic approaches for this manipulator type generally parametrize the elbow redundancy with an arm angle\u03c8 that defines the relative rotation between the shoulder-elbow-wrist plane and a reference plane (see Fig. 1). Notable advantages of closed-form solutions for robots with a single degree of redundancy compared to classical differential IK solutions are as follows: An exact solution at the position level is returned in one call. Computation times are much faster and real-time requirements can be met more easily. This is especially beneficial for online motion planning. A desired arm angle can be directly specified if needed. Also, continuous switching between such a particular solution and a general solution with optimized arm angle is possible" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001925_978-3-030-46817-0_30-Figure19-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001925_978-3-030-46817-0_30-Figure19-1.png", + "caption": "Fig. 19. FEM model - displacement (magnitude).", + "texts": [], + "surrounding_texts": [ + "Control device MR 96 is the horizontal testing machine suitable for tensile test of long materials such as steering tie rods. The tie rod is loaded in tension. The tensile test on three tie rod sample, of 21 mm inner diameter and 26 mm outer diameter and 329 mm of length, have provided the following results presented in Table 4. During the tensile test, all results were recorded. The tensile force is calculated using: Ft = A\u2219E\u2219et, where:" + ] + }, + { + "image_filename": "designv10_14_0001998_j.engfailanal.2020.104942-Figure4-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001998_j.engfailanal.2020.104942-Figure4-1.png", + "caption": "Fig. 4. Planet gear bore-rim crack to simulate the latter stage of the H-225 planet gear failure as shown in Fig. 1, where the simulated crack (an EDM notch) size is 10-mm \u2013 half way through the gear rim width of 20 mm.", + "texts": [ + " We will compare the C-TSA with the less commonly used composite planet TSA (PTSA) as the first step in our analysis to illustrate the need for using a planet based TSA rather than carrier based TSA for detecting planet gear bore-rim cracks. We then discuss the results under a new HT-TSA resampling scheme, i.e. BSRS, that enables the simultaneous removal of multiple sets of shaft harmonics and their intermodulation sideband components. Lastly, we will show how effective the EPTSA technique is in extracting the fault information related to the planet gear bore-rim crack. Using the vibration data from the test rig with the cracked gear shown in Fig. 4, we obtained the spectra of the time synchronous resampled (TSR) vibration signal with respect to planet gear rotation and its corresponding time synchronous average (TSA), i.e. the composite planet-TSR (P-TSR) and P-TSA (where Navg = 7 \u00d7 80 = 560 and Nppr = 10557) respectively. From the P-TSR spectrum in the blue curve in Fig. 5, we can see that the dominant spectral content relates to the 53rd planet pass harmonic or 159th carrier shaft order, W. Wang et al. Engineering Failure Analysis 117 (2020) 104942 i" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000025_1.4024783-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000025_1.4024783-Figure3-1.png", + "caption": "Fig. 3 (a) Definition of roughness lay direction angle hi for surface i (i \u00bc 1; 2) and (b) example roughness profile (Eq. (21a) with hi \u00bc 90 and t \u00bc 0) to show sharp valleys and relatively smooth peaks", + "texts": [ + " For a complete scuffing simulation, the normal load is increased incrementally from 18 N (corresponding to Hertzian pressure of ph \u00bc 0:76 GPa) to 623 N (corresponding to Hertzian pressure of ph \u00bc 2:47 GPa) in 30 constant loading steps with one minute of operating time for each step as shown in Fig. 2. As Li et al. [1] showed that the critical scuffing temperature for the steellubricant pair considered is in the range of 450\u2013500 oC, the critical scuffing temperature is therefore set at Tc \u00bc 500 oC in this study and the simulation is stopped whenever the maximum surface temperature reaches Tc. Defining the roughness lay direction angle (h1 for surface 1 and h2 for surface 2) as the angle between the roughness lay direction and the sliding direction x as shown in Fig. 3(a), ten different combinations of roughness lay directions of the contacting surfaces as defined in Table 1 are investigated to quantify the influences of roughness directionality on scuffing failure. The roughness profile of surface i (i \u00bc 1; 2) takes the form of Si \u00bc 2A ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi sin p x u1t y cot\u00f0hi\u00de\u00bd B s (21a) when hi > 0 . Here, the amplitude A \u00bc 0:5 lm and wavelength B \u00bc 35 lm", + " When hi \u00bc 0 Si \u00bc 2A ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi cos py B n o r (21b) for the case that the roughness profiles of the two surfaces are inphase (peaks versus peaks; valleys versus valleys) and S1 \u00bc 2A ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi cos py B n o r (21c) S2 \u00bc 2A ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi sin py B n o r (21d) for the case that the mating roughness profiles are out-of-phase (peaks versus valleys). The roughness profiles taking such forms have the attributes of sharp valleys while relatively smooth peaks as shown in Fig. 3(b) in order to simulate the after run-in engineering surface roughness profiles such as those measured in Ref. [20]. A typical turbine oil Mil-PRF-23699 is used as the lubricant with its temperature set at Tf 0 \u00bc 120 C, at which, the lubricant has an ambient density of q0 \u00bc 935:2 kg=m3, viscosity of g0 \u00bc 0:00273 Pa s and two-slope pressure-viscosity coefficients of a1 \u00bc 11:3 GPa 1 and a2 \u00bc 5:29 GPa 1. The transition pressure for the two-slope pressure-viscosity relationship of Eq. (6) is curve-fitted as pt \u00bc 79:0163\u00fe 0:91401 Tf 0 with the units of pt in MPa and temperature Tf 0 in degree Celsius" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001739_tec.2020.2995880-Figure23-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001739_tec.2020.2995880-Figure23-1.png", + "caption": "Fig. 23 Prototype stator and rotor (a) rotor (b) stator", + "texts": [ + " Personal use is permitted, but republication/redistribution requires IEEE permission. See http://www.ieee.org/publications_standards/publications/rights/index.html for more information. in two air gaps are identical, so the measured back EMF half value of the prototype is taken to compare with the results by the analytical model and FEA. The structure parameters of the prototype are given in Tab. 1. The magnet length of the prototype is 8mm owing to the symmetry, and the prototype and the test platform are shown in Fig. 23, and Fig. 24, respectively. The line back EMF results of the analytical model, FEA and test are illustrated in Fig. 25. In Fig. 25, the analytical model on the back EMF calculation is verified indirectly by the good agreement of the AFPMM back EMF results between the analytical model, FEA and test. This paper presents an analytical model for calculating the electromagnetic performance of AFPMMs with the mixed eccentricity including both the rotor radial deviation and the angular eccentricity without using the time-consuming 3-D FEA" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000160_pime_proc_1962_176_050_02-Figure71-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000160_pime_proc_1962_176_050_02-Figure71-1.png", + "caption": "Fig. 71. Sign convention diagram", + "texts": [ + " 70) Sweep velocity on outer track (actual) 7r -- 720 DONc fi/sec (84) Sweep velocity on roller surface (actual) 71 -- dn fi/sec (85) 720 Sweep velocity on inner track (actual) 7r -- Di(N-Nc) fi/sec (86) 720 For theoretical (no-slip) conditions (84) = (85) = (86) i~ DoDiN - 720 Di + Do ft/sec (87) Sliding velocity is the algebraic difference between sweep Sliding velocity (actual) between inner track and roller velocities on the mating surfaces. 7f = (86)-(85) = - [Di(N-Nc)-d,] fi/sec (88) 720 Sliding velocity (actual) between outer track and roller 7r = (85)-(84) = - [d,,-D,Nc] ft/sec (89) 720 Sign conventions (Fig. 71) Both the sliding velocities and the cage slip shown in Fig. 71 are termed positive since the sweep velocities progressively diminish from the source of power (inner race A in diagram) through the intermediate member (roller B) to the h a l driven member (cage C). This is the natural condition considering transmission against motion-opposing drag forces. Note: Negative sliding conditions are quite feasible, for heavy drag between the inner race and cage bore of an inner-centred cage could result in negative cage slip and negative sliding at the outer race under steady running conditions" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001646_s11771-019-4207-3-Figure9-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001646_s11771-019-4207-3-Figure9-1.png", + "caption": "Figure 9 Schematic structure of face gear hobbing machine (1\u2212Axial slide(x); 2\u2212Radial slide(y); 3\u2212Additional slide(z); 4\u2212Work piece axis(B); 5\u2212Spindle axis(C))", + "texts": [ + " Figure 7 shows the difference among each row of teeth on the hob. The angle \u03b5 between the centerline of each gear and the bottom edge is different. The difference value between adjacent blades is: 1 1 2 2\u03c0 / | | N n (9) The cutter teeth are assembled in a certain order on the cutter body groove, and the axial position is fixed using the wedge, both ends are J. Cent. South Univ. (2019) 26: 2704\u22122716 2709 fixed with the end cover, the end cover can be pressed on the cutter body by the compression nuts, as shown in Figure 8. Figure 9 shows the structure of the face gear hobbing machine, including three moving shafts (x, y, z) and two rotating axes (B, C). Motion in the x direction is used to achieve radial feed motion. The motion in the y direction is used to achieve axial feed motion. The z-direction motion is used to realize the additional motion of the cutter relative to the face gear. The rotational motion in the B direction is used to realize the rotation of the face gear along its own axis. The rotational motion in the C direction is used to realize the rotation of the cutter around its own axis", + " South Univ. (2019) 26: 2704\u22122716 2713 Figure 16 shows the tooth profile errors of face gear affected by a similar offset error of \u03b4. The error trends of inner and outer diameter tooth surface under front and back displacement errors influence are the same as that of tool radial errors. There are differences between the error values and the error range. According to the generation hobbing principle of face gear, we develop the five-axis face gear hoobing machine, and the schematic structure is shown in Figure 9. The straightness of (x, y, z) direction guide are 0.02 mm/m, 0.03 mm/m and 0.03 mm/m, respectively. The dividing accuracy of B axis is 7 ms, the run out of C axis is 0.01 mm, and the power of spindle motor is 18 kW, as well as the maximum speed of spindle is 1500 r/min. The assembled spherical hob machining process and the assembled hob are shown in Figures 17 and 18, respectively. The face gear hobbing experiment is carried out on the developed face gear hobbing machine tool. In the hobbing process, the workpiece speed is 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001910_j.ijnaoe.2019.11.004-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001910_j.ijnaoe.2019.11.004-Figure2-1.png", + "caption": "Fig. 2. The motion items of the surface vessel in Body-fixed Coordinate and Earth-fixed Coordinate.", + "texts": [ + " Measurement system: The measurements are accomplished by using the sensors, including digital compass, gyroscope and accelerometer. The measured data are processed and logged in myRIO for further analysis. 5) Host computer: Host computer installed the LabVIEW is used to deploy the program and control the FRMSS remotely. The motions of the ship contain three components: the position and orientation vector h defined in eartg-fixed coordinate, the linear and angular velocity vector n as well as the force andmoment vector t defined in body-fixed coordinate. The relevant items to describe Hoorn's motion are illustrated in Fig. 2. Normally, the three DOF mathematical model is sufficient to describe the surface vessel's motion. However, the experimental reading of the yaw rate transformed from the body-fixed coordinate to earth-fixed coordinate is influenced by huge roll motion. Thus, the ship's four DOF dynamic model is employed (Fossen, 1994) as Eq. (1):8><>: \u00f0m\u00femx\u00de _u m\u00femy vr \u00bc X m\u00femy _n\u00fe \u00f0m\u00femx\u00deur \u00femyay _r myly _p \u00bc Y \u00f0Ix \u00fe Jx\u00de _p myly _v mxlxur \u00feWHf \u00bc K \u00f0Iz \u00fe Jz\u00de _r \u00fe myay _v \u00bc N (1) where the involved items are outlined in Table 2" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001847_tro.2020.3038687-Figure8-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001847_tro.2020.3038687-Figure8-1.png", + "caption": "Fig. 8. Module design of anthropomorphic robotic arm. The movements of a human arm are divided into three joint modules.", + "texts": [ + " In previously reported anthropomorphic robotic arms with tendon-driven mechanism [14], [15], the entire arm was designed by locating heavy motors on the base or the proximal part, whereas in this article, an anthropomorphic robotic arm is designed using the proposed 2M2D and 3M3D joint modules to illustrate the practicality and usability of the modularization of couple tendon-driven joints. It has 7 DoFs, the same as in a human arm, and incorporates all the motors and mechanical parts inside the arm. A human arm has seven pairs of motions for seven DoFs, i.e., three DoF in the shoulder, one in the elbow, one in the forearm, and two in the wrist. The robotic arm in Fig. 8 was designed to enable the coupling of multiple motors and realize Authorized licensed use limited to: University of Gothenburg. Downloaded on December 21,2020 at 07:20:41 UTC from IEEE Xplore. Restrictions apply. the corresponding motions of the human arm. The wrist adduction/abduction and extension/flexion, and the elbow extension/flexion and forearm supination/pronation were obtained by two 2M2D joint modules. The shoulder adduction/abduction, extension/flexion, and internal rotation/external rotation were obtained by a 3M3D joint module" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002369_tmag.2021.3064402-Figure22-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002369_tmag.2021.3064402-Figure22-1.png", + "caption": "Fig. 22. Equivalent winding for rotor PMs.", + "texts": [ + " The torque fluctuations also lead to the fluctuation of rotor speed. However, the maximum fluctuation of rotor speed is about 1.5 rpm under TL = 25 N\u00b7m. The introduced SPMSM in [30] and [33] is modeled by IWFT while considering the rated parameters introduced in Table V. The iron material of stator and rotor cores is the silicon steel nonoriented grain, which has the B\u2013H curve, as shown in Fig. 21 [35]. 1) Calculation of Inductances: For modeling the SPMSM by using IWFT, it is necessary to replace the rotor PMs with an equivalent winding. Fig. 22 shows this process for analyzed SPMSM with radial magnetized PMs. Authorized licensed use limited to: Carleton University. Downloaded on June 16,2021 at 09:17:18 UTC from IEEE Xplore. Restrictions apply. where lm and M are, respectively, the radial thickness and magnetization vector of PMs. Fm is also the MMF of each PM pole. Assuming Im = 1 (A), equivalent winding includes four PM coils; each of them has Fm turns. In the case of large air-gap electric motors, both radial and tangential components of winding functions should be considered to calculate the magnetizing component and air-gap leakage component of self-inductance" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000175_j.mechmachtheory.2015.11.014-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000175_j.mechmachtheory.2015.11.014-Figure3-1.png", + "caption": "Fig. 3. A meshing gear pair and its coordinate systems.", + "texts": [ + " The tooth flank of work gear surface is longitudinal crowned by using Eq. (7) and free twist by applying the VPA honing cutter with a suitable VPA coefficient of rack cutter b, as shown in Fig. 1(a). A meshing gear pair is composed of the proposed crowned helical pinion (Gear A) and the standard helical gear (Gear B). The surfaces of the Gear A are longitudinal crowned with twisted tooth flank (Case 1) and twist-free tooth flank (Case 2). Coordinate systems for the meshing of a gear pair are shown in Fig. 3, wherein coordinate systems S1(x1, y1, z1), S2(x2, y2, z2) and Sf(xf, yf, zf) are rigidly connected to the Gear A, Gear B, and the fixed frame, respectively. Si(xi, yi, zi) is the auxiliary coordinate system. Eog is the operating center distance between Gear A and Gear B, ro1 and ro2 are the radiuses of operating pitch circle of the Gear A and Gear B, respectively. By applying the coordinate transformation method, the tooth surface equations and surface unit normal vectors of the Gear A and Gear B can be represented in the fixed coordinate system Sf" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002178_s11665-021-05944-5-Figure13-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002178_s11665-021-05944-5-Figure13-1.png", + "caption": "Fig. 13. Schematic diagram of the cross section of a deposited sample", + "texts": [ + " 11(a) are Journal of Materials Engineering and Performance approximately 20-50 lm in size, and the pores in Fig. 11(b) are approximately 10-30 lm in size. It is consistent with the increase in elongation after adding laser. The reason is the pulse laser can stir the molten pool, which increases the flow of the molten pool and promotes the escape of gas. It effectively reduces the internal porosity of samples (Ref 31, 32). Figure 12 shows the micro-hardness distribution of the samples with different laser power along the deposition direction. The red dotted lines in Fig. 13 represent the measuring position (A test point is every 0.1 mm). The hardness values of the sample change periodically along the deposition direction, which is related to the periodic distribu- tion of the microstructure along the deposition direction (Ref 33, 34). Table 3 is the average hardness of different laser power. And micro-hardness of the sample with 300 W laser is 88.8 HV0.2, with an increase of 11.2% compared to that without laser. It is concluded that the refined grains is the main reason for the increase in microhardness" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002005_tmag.2020.3027816-Figure6-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002005_tmag.2020.3027816-Figure6-1.png", + "caption": "Fig. 6. Magnetic field strength distributions. (a) inclined eccentricity and (b) parallel eccentricity. Red lines show eccentric direction of the rotor.", + "texts": [ + " (The rotor cross section of inclined eccentricity at z = 28.5mm is shifted by the length 28.5tan\u03b8i \u2248 0.4mm.) However, the cross section at z = 28.5mm in Fig. 4 (a) has more asymmetric distribution of the magnetic flux density than Fig. 4 (b), because the difference in the magnetic flux density between the point (A) and (B) is greater. Fig. 5 shows the magnetic flux density distributions at the gap center in the section of z = 28.5mm. The variation of the amplitude value for inclined eccentricity is larger than that for parallel eccentricity. Fig. 6 shows the distribution of magnetic field strength of magnets. In the case of (a)inclined eccentricity, the magnetic field strength of each magnet is almost the same. In the case of (b)parallel eccentricity, the magnetic field strength of the leftside magnets is larger than that of the right-side magnets. Hence, the expression 1 \u2013 \u03bbcos\u03b8 need to be considered in the expression of electromagnetic force. The value of \u03bb for parallel eccentricity is larger than that for inclined eccentricity. The cause that the value of \u03bb for parallel eccentricity is smaller than that for inclined eccentricity can be explained by using the gap magnetic circuit" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002134_lra.2021.3066961-Figure4-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002134_lra.2021.3066961-Figure4-1.png", + "caption": "Fig. 4. (a) The split belt treadmill test setup and safety harness used by participants. (b) A Simulink model for data analysis, programming, and controlling the SR-AFO. (c) The control box which houses all the eletro-pneumatic hardware, pressure gauge, and pressure regulation hardware. (d) The portable compressor used to provide fluidic actuation to the SR-AFO.", + "texts": [ + " While effective in many ways, these methods are often extremely costly and complex in terms of computation. By instilling protocol that relies on human entrainment to assist in restoring natural gait patterns, computationally heavy algorithms can be replaced with simple periodic pulses, which are tailored to each individual and contain no closed feedback loop once active. An open-loop control strategy is used to apply fixed, controlled perturbations to the user by utilizing a portable compressor (Model 8010 A, California Air Tools, USA) (Fig. 4(d)) to pressurize the cells of the soft actuators for a specified inflation time. The control strategy, designed in Simulink (Mathworks, MA, USA), contains a modifiable algorithm for updating the frequency of applied perturbations during each trial. The open-loop control strategy involves no feedback of any kind, and provides gentle, heavily-damped, soft perturbations via the pneumatic actuators, which differ from the rigid and abrupt robotic torque pulse used in previous entrainment studies [3]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001925_978-3-030-46817-0_30-Figure26-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001925_978-3-030-46817-0_30-Figure26-1.png", + "caption": "Fig. 26. Max warp angle validation of FEA model.", + "texts": [], + "surrounding_texts": [ + "Figs. 23, 24, 25 and 26. 142 I. Alagi\u0107 In case of ball joint, the results of stress distribution and allowed displacement are presented in Figs. 27, 28, 29, 30, 31 and 32. Finite Element Analysis (FEA) of Automotive Parts Design as Important Issue 143 The displacement achieved as a result of finite element analysis doesn\u2019t much differ from results of the laboratory-test performed by control device MR 96. The maximum displacement appeared into Z direction 0,145 [mm]. On the basis of conducted simulations were possible to affirm that the magnitude of deformations depends on model geometry. The largest concentration of stresses appeared in places near of cover of ball joint. Through change of model geometry it was possible to influence on expansion of stresses and displacement distribution. To analysis of maximum warp angle of ball joint using FEA solver was very low compared to its allowed value 58 \u00b1 6\u00b0." + ] + }, + { + "image_filename": "designv10_14_0000898_s12206-015-0417-2-Figure8-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000898_s12206-015-0417-2-Figure8-1.png", + "caption": "Fig. 8. Blade coordinate system for pitch bearings design.", + "texts": [ + " The bearing was designed to satisfy the criteria required for pitch bearings in wind turbines, such as fatigue life against fatigue loads, static load factor against extreme loads, and safety factors against surface durability and root breakage of gear teeth, which are regulated and calculated by the related guidelines [7, 8]: larger fatigue life than the required life, more than unity for static load factor, more than unity for safety factor against gear teeth surface durability under fatigue and extreme loads, and more than 1.15 and 1.10 for safety factor against gear teeth root breakage under fatigue and extreme loads, respectively. The specifications and photograph of the bearing are seen in Table 2 and Fig. 7, respectively. Among several coordinate systems for wind turbine com- ponents design, the blade coordinate system is applied for pitch bearings, as shown in Fig. 8 [7]. In this coordinate system the z axis indicates axial direction parallel to the blade, the x axis indicates the radial direction coinciding with a wind turbine rotor axis, and the y axis indicates the radial direction perpendicular to the x axis. The loads applied to the pitch bearing under a field environment can be divided into extreme and fatigue loads [7]. Extreme load is the static load during a wind gust event. Fatigue load, however, is the continuous load applied during the service life and can be expressed as the load duration distribution or damage equivalent load" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002152_13506501211010030-Figure10-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002152_13506501211010030-Figure10-1.png", + "caption": "Figure 10. Splash lubrication process: (a) initial stage, (b) free-flow stage, (c) ejection flow stage and (d) splashing flow stage.", + "texts": [ + " It showed that the CFD simulation can well predict the churning power loss of the spiral bevel gears in the gearbox under splash lubrication. Churning power loss of the intermediate gearbox Figure 9 shows the curve of churning power loss of the intermediate gearbox calculated by CFD simulation. We can obtain the churning power loss of the helicopter intermediate gearbox as \u223c2.6051 kW. The rated input power of the gearbox is 135 kW, and the churning power loss accounts for \u223c1.93% of the rated input power. Splash lubrication process of the intermediate gearbox As shown in Figure 10, the splash lubrication process of the spiral bevel gears in the intermediate gearbox can be divided into three stages: free flow (Figure 10(b)), ejection flow (Figure 10(c)) and splashing flow (Figure 10(d)). The characteristic of free flow is that in the initial stage, the oil is affected by the rotating gears, flows along the circumference of the gears under the action of its inertia and viscous force, and falls vertically under the influence of gravity. The characteristic of ejection flow is that the oil splashed by gears is thrown out along the rotating direction of the gears. The characteristic of splashing flow is that the inertia force of the oil is larger than the resultant force of gravity and viscous force" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000160_pime_proc_1962_176_050_02-Figure103-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000160_pime_proc_1962_176_050_02-Figure103-1.png", + "caption": "Fig. 103. Hz&-r vibration frequency (non-dimensional) against non-dimensional rotor spin", + "texts": [ + " Proc Instn Mech Engrs The approximation that cos + = c was not made here. This quartic equation had been solved by Mr S. Michaelson of the Mathematics Department of Imperial College on the London University \u2018Mercury\u2019 computer. The lower, stiff rotor, frequency (taking n = 0-45q) was plotted in Fig. 102. The undamped curve from Fig. 4 of the paper was shown dotted for comparison, both for LID = 1. The higher frequency (considered by Morrison) also appeared in the solution of the equation. This was shown in Fig. 103. The values of this frequency were very sensitive to the magnitude of the velocity coefficient. The vibration was very heavily damped. If the velocity coefficients were estimated from Pinkus and Sternlicht (Tables 8.1 and 8.2) the damping was reduced and at low values of E turned into an amplification, i.e. the vibration became unstable. This might explain why bearings vibrated at low eccentricity ratios, a fact that had never been satisfactorily cleared up. In reply to Mr Hughes, they had not considered the uncavitated film at all but were extremely interested to note that the bearing did not automatically vibrate as soon as the Sommerfeld conditions were established, as the simple application of Routh\u2019s conditions might suggest" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000769_978-3-662-43893-0-Figure4.33-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000769_978-3-662-43893-0-Figure4.33-1.png", + "caption": "Fig. 4.33 Profile lines and contact line on the wheel tooth", + "texts": [ + " The following points should be considered when calculating the load-free contact pattern: \u2013 the tooth contour limiting the contact pattern, \u2013 the simultaneous engagement of several teeth, \u2013 specified or calculated deviations in relative position Other parameters which may be considered are: \u2013 pre-mesh or edge contact, \u2013 specified pitch deviations To determine the contact pattern, the bevel gear set rotates in such a way that at each moment one tooth flank pair is in contact. All meshing positions are therefore considered in predefined steps, taking the transmission error into account, with the currently active flank pair contacting at one point. As load increases, the tooth flanks deform creating a narrow contact ellipse in place of the load-free contact point. A profile line is formed by the points with the smallest contact distance which can be determined for each contact position (see Fig. 4.33). Points on a potential contact line which fall below a certain contact distance (usually 3 to 6 \u03bcm, depending on the thickness of the marking compound) at one meshing position at least, constitute the effective contact line and are part of the contact pattern. Those points with the smallest contact distance on each contact line form what is termed the path of contact. Figure 3.14 shows a contact pattern in which the contact lines\u2014the major axes of the contact pattern\u2014are represented. In Fig. 3", + "4 Stress Analysis 173 Determining tooth deflection Solving the contact problem by means of the influence numbers method supposes that pinion and wheel tooth are sub-divided into several segments along the face width at which different deflections may occur. The load along the contact line is dissected into individual forces corresponding to these sub-divisions, with the load acting in the middle of each considered segment. These segments are interlinked by means of imaginary springs, which is true for all form alterations with the exception of Hertzian contact flattening (deformation). Besides, the segments are created by means of spherical sections of radius r through the bevel gear tooth (Fig. 4.33). An arbitrary line of intersection between the sphere and the tooth flank/fillet yields a profile line. A tooth segment is accordingly formed by two spheres and the profile lines. A point of intersection between a profile line in the middle of a segment and a contact line (calculatedwith the load-free tooth contact analysis) is the working point of an individual force and is described by polar coordinates r and \u03d1. The grid thus created on the wheel tooth is transformed to the pinion tooth in correspondence with the contact line" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001515_j.addma.2019.100892-Figure10-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001515_j.addma.2019.100892-Figure10-1.png", + "caption": "Fig. 10: Preliminary studies concerning the parameters split draw and extrusion.", + "texts": [ + " the insufficient modeling of a smooth surface by a finite element mesh which may become apparent in very slender and thin structures. However, with prescribed values for the parameter mindim clearly defined structures evolve which do not differ significantly, regardless of the prescribed value. For all further considerations, we consider the parameter mindim with an assigned value of 15mm. Results concerning the study of the influence of the parameters split draw and extrusion are presented in fig. 10. Both parameters lead to solutions that can potentially be manufactured by milling. From the manufacturing and postprocessing point of view, the extrusion solution would be easier to handle than the split draw solution since by definition all parts have the same width. However, the extrusion solution leads to structural parts of equal width even in regions where the complete width of the design space is not required to fulfil the requirements Jo ur na l P re -p ro of for the considered C-frame. Further, the extrusion solution does not allow for transitions between wide and potentially more slender regions so that it can be anticipated that such designs will not necessarily be optimal in terms of lightweight design so that we will consider both solution types further for the milling solution" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001645_j.snb.2019.127413-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001645_j.snb.2019.127413-Figure1-1.png", + "caption": "Fig. 1 Schematic presentation of the immunosensor.", + "texts": [], + "surrounding_texts": [ + "The preparation process of G/Co9S8-Pd and assembly of the immunosensor are shown in the supporting information. The fabricated label-free immunosensor was tested using an amperometric i\u2013t curve method. First, PBS (10 mL, pH=7.4) was prepared as base solution. Under a scanned potential of -0.4 V and after a stable background signal was established, 10 \u03bcL of H2O2 (5 mol/L) was injected into the PBS solution under mild stirring. Finally, the variation of response currents of the tested Jo na immunosensor were recorded." + ] + }, + { + "image_filename": "designv10_14_0000123_j.sna.2014.10.039-Figure6-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000123_j.sna.2014.10.039-Figure6-1.png", + "caption": "Fig. 6. Structure of the SPR sensor.", + "texts": [ + " Used the same method, the frequency shift was 621 Hz, and the thickness was ca. 8.5 nm for 12 bilayer film. After assembly, as shown in Fig. 5, we can see the different morphology of the surfaces using the atomic force microscopy (AFM). In addition, the surface roughness (Rq) for the bare sensor and assembled with 12 layers polymers are 5.19 nm and 4.10 nm, respectively. 3. Measurement system for glucose concentration 3.1. SPR sensor An SPR sensor in the Krestchman structure (ICx Technologies) (Fig. 6) was used in this study, and the exterior and interior structures included a LED, an aperture and polarizing film, a gold film, a reflecting mirror, a photodiode array, and a memory chip. When polarized light of a given wavelength is reflected from the sensing surface over a range of incident angles, surface plasmon resonance The intensity of the reflected light decreases markedly, which can be detected by a photodiode array. The software can convert the SPR angle to a refractive index to obtain the concentration of the analyte [19,20]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000080_j.engfailanal.2013.03.008-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000080_j.engfailanal.2013.03.008-Figure2-1.png", + "caption": "Fig. 2. Model I of crack.", + "texts": [ + " This constant shows that the ability of a material to resist fracture. Usually this constant is called the fracture toughness, as expressed by KIC. That is: KIC \u00bc rc ffiffiffi a p Y \u00f01\u00de KIC is the measurement of the material to prevent unstable crack growth capacity, and is the material characteristic parameter. It is irrelevant to the crack size, geometry, and loading, and is only relevant to material composition, heat treatment, and machining process. Obviously, the greater KIC is, the greater the capacity of the material to resist fracture. As shown in Fig. 2, it is possible to test the KIC values of various materials by experimental methods. The penetrating crack with length 2a is in the center of the sample. External stress is perpendicular to the crack plane. The coordinate system xoy1 is used in Fig. 2, and the following stress distribution near the crack tip can be proved [22]: rx \u00bc K1ffiffiffiffiffi 2pr p cos h 2 1 sin h 2 sin 3h 2 ry \u00bc K1ffiffiffiffiffi 2pr p fcos h 2 1\u00fe sin h 2 sin 3h 2 g sxy \u00bc K1ffiffiffiffiffi 2pr p sin h 2 cos h 2 cos 3h 2 9>= >>; \u00f02\u00de From Eq. (2) it can be seen that the stress components all contain KI which is relevant to crack size, shape, and stress; subscript I indicates that the open form of the crack is under external force. KI is called the SIF of the stress field in fracture mechanics" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000943_1350650115611155-Figure10-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000943_1350650115611155-Figure10-1.png", + "caption": "Figure 10. Coordinate system of force on bearing.", + "texts": [], + "surrounding_texts": [ + "Ball bearing, quasi-static model, nonuniform, preload, heat generation Date received: 16 September 2014; accepted: 22 July 2015" + ] + }, + { + "image_filename": "designv10_14_0001573_j.mechmachtheory.2018.12.031-Figure4-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001573_j.mechmachtheory.2018.12.031-Figure4-1.png", + "caption": "Fig. 4. The local frame of substructure i j .", + "texts": [ + " End reaction W stru, ij of substructure i j is just the reaction of joint i j + 1 , namely, W stru,i j = ( F stru,i j M stru,i j ) = ( F i, j+1 M i, j+1 ) (13) where F stru, ij and M stru, ij denote the end reaction force and end reaction moment of substructure i j . 3.2. Strain energy of a substructure For convenience, strain energy substructure i j is solved with respect to a local frame o ij -x ij y ij z ij , in which the origin o ij is at the geometrical center of joint i j , x ij axis is along the link centerline, z ij axis along the axis of joint i j , and y ij axis is determined by the right hand rule, shown in Fig. 4 . With respect to the local frame, W stru, ij can be expressed as W L stru,i j = ( F stru,i j x i j , F stru,i j y i j , F stru,i j z i j , M stru,i j x i j , M stru,i j y i j , M stru,i j z i j )T (14) and gravity G ij of substructure i j can be expressed as G L i j = ( G i j x i j , G i j y i j G i j z i j ) (15) Based on the Clapeyron\u2019s principle [37] , strain energy of substructure i j under its gravity and end reaction can be ex- pressed as E i, j = 1 2 ( W L stru,i j )T ( C e \u2192 e i j W L stru,i j + C g\u2192 e i j G L i j ) + 1 2 ( G L i j )T ( C e \u2192 g i j W L stru,i j + C g\u2192 g i j G L i j ) (16) where C e \u2192 e i j and C g\u2192 e i j denote compliance matrices of substructure i j at the end point with respect to applications of W L stru,i j and G L i j , respectively, and C e \u2192 g i j and C g\u2192 g i j denote compliance matrices of substructure i j at the gravity center with respect to applications of W L stru,i j and G L i j , respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000175_j.mechmachtheory.2015.11.014-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000175_j.mechmachtheory.2015.11.014-Figure1-1.png", + "caption": "Fig. 1. Generation for honing cutter: (a) surface parameters of VPA rack cutter; (b) coordinate systems of the schematic process.", + "texts": [ + " In addition, 3-D models of the meshing gear pairs composed by a crowned pinion and a stand helical gear are constructed and analyzed to illustrate their transmission and load characteristics by using KISSsoft software. The effects working conditions of the meshing gear pair (contact temperature and power loss) have significantly advances when the proposed method with twist-free tooth flanks of the crowned work gear is applied in external gear honing process. To generate the honing cutter's profile with a variable pressure angle, the tooth profile of the standard rack cutter is modified with the pressure angle changing in the lead direction, as shown in Fig. 1(a). Wherein son is the normal circular tooth thickness of the rack cutter. According to Fig. 1(a), the position vector and normal vector of the rack cutter's right-hand side profile can be expressed in the coordinate system Sr(xr, yr, zr) as follows: rr u1; v1\u00f0 \u00de \u00bc xr u1; v1\u00f0 \u00de yr u1; v1\u00f0 \u00de zr u1; v1\u00f0 \u00de 1 2 664 3 775 \u00bc u1 cos\u03b1on u1 \u2212 sin\u03b1on \u00fe bv1\u00f0 \u00de v1 1 2 664 3 775; \u00f01\u00de and Nr u1; v1\u00f0 \u00de \u00bc nxr u1; v1\u00f0 \u00de nyr u1; v1\u00f0 \u00de nzr u1; v1\u00f0 \u00de 2 4 3 5 \u00bc sin\u03b1on\u2212bv1 cos\u03b1on \u2212bu1 cos\u03b1on 2 4 3 5; \u00f02\u00de where u1 and v1 are the rack cutter surface parameters, \u03b1on is the pressure angle of the rack cutter, and b is VPA coefficient of rack cutter, as shown in Fig. 1(a). The schematic generation mechanism of these honing cutters is shown in Fig. 1(b), where coordinate systems Sr(xr, yr, zr), Sh(xh, yh, zh), and S4(x4, y4, z4) are rigidly connected to the rack cutter, honing cutter, and frame, respectively. And roh is the radius of the operating pitch circle of the honing cutter, \u03b2oh is the helix angle of the honing cutter. The generated honing cutter rotates through an angle \u03c81 about the z4-axis, while the rack cutter translates a distance roh\u03c81. By applying the homogenous coordinate transformation matrix equation, the locus and normal vectors of rack cutter surface represented in the coordinate system Sh(xh, yh, zh) are attained by rh u1; v1;\u03c81\u00f0 \u00de \u00bc Mhr \u03c81\u00f0 \u00de rr u1; v1\u00f0 \u00de \u00bc cos\u03c81 \u2212 cos \u03b2oh sin \u03c81 \u2212 sin \u03b2oh sin \u03c81 roh cos \u03c81 \u00fe \u03c81 sin \u03c81\u00f0 \u00de sin \u03c81 cos \u03b2oh cos\u03c81 \u2212 sin \u03b2oh cos\u03c81 roh sin \u03c81 \u00fe \u03c81 cos\u03c81\u00f0 \u00de 0 \u2212 sin \u03b2oh cos \u03b2oh 0 0 0 0 1 2 664 3 775 xr yr zr 1 2 664 3 775; \u00f03\u00de and Nh u1; v1;\u03c81\u00f0 \u00de \u00bc Lhr \u03c81\u00f0 \u00de Nr u1; v1\u00f0 \u00de; \u00f04\u00de where Lhr(\u03c81) is the upper-left (3 \u00d7 3) submatrix of the (4 \u00d7 4) homogeneous coordinate transformation matrix Mhr(\u03c81)", + " The rotational relationship between the work gear and honing cutter is defined as \u03d51 \u03d5h; l\u00f0 \u00de \u00bc Nh N1 \u03d5h \u00fe tan \u03b2o1 ro1 l \u00f012\u00de where symbols Nh and N1 indicate the number of teeth of the honing cutter and gear, respectively. The tooth surface of honed work gear can be defined by solving Eqs. (6), (8), (9), (10), (11) and (12), simultaneously. The tooth flank of work gear surface is longitudinal crowned by using Eq. (7) and free twist by applying the VPA honing cutter with a suitable VPA coefficient of rack cutter b, as shown in Fig. 1(a). A meshing gear pair is composed of the proposed crowned helical pinion (Gear A) and the standard helical gear (Gear B). The surfaces of the Gear A are longitudinal crowned with twisted tooth flank (Case 1) and twist-free tooth flank (Case 2). Coordinate systems for the meshing of a gear pair are shown in Fig. 3, wherein coordinate systems S1(x1, y1, z1), S2(x2, y2, z2) and Sf(xf, yf, zf) are rigidly connected to the Gear A, Gear B, and the fixed frame, respectively. Si(xi, yi, zi) is the auxiliary coordinate system" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000740_s00170-018-2841-9-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000740_s00170-018-2841-9-Figure2-1.png", + "caption": "Fig. 2 Coordinate systems of grinding wheel and screw rotor", + "texts": [ + " Given the shaft section of the grinding wheel is composed of multiple discrete points (Zc, Rt), the revolution surface equation of grinding wheel can be expressed as follows [5]: X c \u00bc Rtcos\u03d5 Yc \u00bc Rtsin\u03d5 Zc \u00bc f Rt\u00f0 \u00de 8< : \u00f01\u00de where Xc, Yc, and Zc are revolution surface equation of grinding wheel; Rt is the radius when the width of the grinding wheel is Zc; \u03d5 is the parameter variable that is angle between Rt and XcOcZc plane; Xc to Yc direction is defined as positive direction, as shown in Fig. 1. Then, Eq. (1) can be rewritten in vector form as: R !\u00bc X c Rt;\u03d5\u00f0 \u00de; Yc Rt;\u03d5\u00f0 \u00de; Zc Rt;\u03d5\u00f0 \u00de\u00f0 \u00de The axes of the grinding wheel and the screw rotor are crossed, forming a mounting angle \u03c9. T is the distance between the axes of the screw rotor and the grinding wheel. The point M is a point on the contact line between grinding wheel and screw rotor. O \u2212 XYZ is the screw rotor coordinate system; O \u2212 XcYcZc is the grinding wheel coordinate system, as shown in Fig. 2. The contact line equation can be expressed as [5]: k ! r!\u00fe p k ! \u2022 n!\u00bc 0 \u00f02\u00de where p is the screw parameter of rotor calculated by p = S/2\u03c0; S is the lead of the screw rotor; r! is the vector OM ! in the screw rotor coordinate systems O \u2212 XYZ; k ! is the unit vector in the Z axis of the screw rotor coordinate system O \u2212 XYZ; and n! is the normal vector of contact point M. In the grinding wheel coordinate system O \u2212 XcYcZc, the vector r! can be expressed: r!\u00bc R !\u00fe T j ! c \u00bc Rtcos\u03d5 i ! c \u00fe Rtsin\u03d5 j ! c \u00fe f Rt\u00f0 \u00de k!c \u00fe T j ! c \u00f03\u00de M M Oc Oc -Yc Xc Zt Yc Zc Rt YcFig. 1 Schematic diagram of grinding wheel structure where i ! c, jc ! , and kc ! are the unit vectors in the grinding wheel coordinate system O \u2212 XcYcZc; i ! , j ! , and k ! are the unit vectors in the screw rotor coordinate system O \u2212 XYZ. Figure 2 shows the geometry relationship between grinding wheel coordinate system O \u2212 XcYcZc and screw rotor coordinate system O \u2212 XYZ. X \u00bc X ccos\u03c9\u00fe Zcsin\u03c9 Y \u00bc Y c \u00fe T Z \u00bc Zccos\u03c9\u2212X csin\u03c9 8< : \u00f04\u00de i ! c \u00bc cos\u03c9 i !\u2212sin\u03c9 k ! j ! c \u00bc j ! k ! c \u00bc cos\u03c9 k !\u00fe sin\u03c9 i ! 8>< >: \u00f05\u00de Substituting Eq. (5) into Eq. (3), the vector r! can be expressed in the screw rotor coordinate system O \u2212 XYZ: r!\u00bc Rtcos\u03d5cos\u03c9\u00fe f Rt\u00f0 \u00desin\u03c9\u00bd i!\u00fe Rtsin\u03d5\u00fe T\u00f0 \u00de j ! \u00fe f Rt\u00f0 \u00decos\u03c9\u2212Rtcos\u03d5sin\u03c9\u00bd k! \u00f06\u00de The term k ! r!\u00fe p k ! can be calculated as follows: k " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001805_s12206-020-0909-6-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001805_s12206-020-0909-6-Figure1-1.png", + "caption": "Fig. 1. Inner view of a gear box.", + "texts": [ + " Depending on the number of teeth present in a gear, the ratio of the angular velocities known as gear ratio is determined among the input and output of a gear. If more than two gears are made to function in a sequence, then it results in the gear transmission or gear train [2]. There are various types of gears are available like external and internal gear, helical gear, bevel gear, spiral bevel gear, spur gear, worm gear, epicyclic gear, magnetic etc. and many. Similarly there is different types of teeth are also available like straight teeth, helical teeth, chevron teeth, etc. The Fig. 1 gives the inner view of a gear box which possess 5 gears [3]. Almost all the vehicles that are manufactured nowadays comprise of 5 gears gear box mechanism. As seen, all the gears have been mounted on a drive shaft and then the driven gears are mounted on a counter shaft. Nowadays, new innovation like miniature type of gears have been manufactured as these gears helps in reducing the gear noise and improving the accuracy of the transmission and the ability to carry the load [4]. Different types of miniature gears the gear which has the outer diameter value which is very less than 1 mm and hence known with the name of micro gears" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001369_tmag.2016.2517598-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001369_tmag.2016.2517598-Figure2-1.png", + "caption": "Fig. 2. Single-axis square Helmholtz coils", + "texts": [ + " Thus, the magnetic flux density at the center can represent the magnetic flux densities within the whole uniform region. According to the Biot-Savart law, the relationship between the magnetic flux density at the center of the single-axis square Helmholtz coils and the current is iii IKB (3) where Ki (i = 1, 2, 3) are the structural coefficients of inner, middle, and outer coils respectively, Ii (i = 1, 2, 3) are the currents fed to the inner, middle, and outer coils respectively. The structure of the single-axis square Helmholtz coils is shown as Fig. 2. The structural coefficient of each single-axis square Helmholtz coils is derived respectively as [30] 2 2 2 2 2 0 2 2 2 \u03c0 2 4 i i ii i i i ii i d t ad t a t aN K \u03bc i =1, 2, 3 where 2a denotes length of the coil, t denotes the thickness of the coil, 2d denotes the distance between the central points of two coils, l denotes the width of the coil, \u03bc0 = 4 107 H/m denotes the permeability of vacuum, Ni denotes the coil turns. The phase difference between current and voltage caused by inductance in each of three Helmholtz coils is described respectively as i i i R L arctan i = 1, 2, 3 (4) 0018-9464 (c) 2015 IEEE" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002049_1464419320972870-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002049_1464419320972870-Figure1-1.png", + "caption": "Figure 1. (a) shaft-bearing-pedestal system (b) simplified model of rolling ball bearing. Figure 2. Contact stiffness and oil film stiffness.", + "texts": [ + " Harmonics frequencies of fBPFI are modulated mainly by 2fs instead of fs in frequency domain for multiple defects on inner raceway. Then the effects of the angle between two defects (hI/OAD), the number of defects (NDI/O) and the location of defects on outer raceway on dynamic response have been studied. Moreover, a series of formulae describ- ing relationship between amplitude of fBPFO/fBPFI and hOAD/hIAD are shown in discussion section. In order to simplify the shaft-bearing-pedestal system shown as Figure 1(a), some necessary assumptions are given as below: 1. The inner race rotates with the shaft and outer race is fixed on pedestal. 2. The inertial force and gyroscopic moment of roll- ing elements are neglected. 3. Relative slip only occurred in loading zone without uneven lubrication. 4. Thermal effects due to friction are neglected. 5. Contact friction is assumed negligible. 6. Isothermal EHL condition is considered. When the inner race rotates with the axis and outer race is fixed with pedestal, the shaft-bearing-pedestal system could be simplified as 4-DOF dynamic model shown in Figure 1(b). Kp, Ks, Cp and Cs are the stiffness and damping coefficient of outer and inner race, respectively. Comprehensive stiffness and damping coefficient As the Hertz contact theory is adopted to express the contact between raceways and rolling element and the mass of rolling element is neglected, the stiffness of rolling element consists of contact stiffness and oil film stiffness,23 which is shown in Figure 2. When the ball rolls over raceway, the contact stiffness could be formulated as37 Ki=o \u00bc 2 ffiffiffi 2 p E 1 l2 3 X qin=ou 1=2 1 d1in=ou " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000969_j.triboint.2016.08.034-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000969_j.triboint.2016.08.034-Figure1-1.png", + "caption": "Fig. 1. Motions of balls and geometry of ball-race contact [21].", + "texts": [ + " [22] investigated the heat dissipation characteristic of ball bearing cage and inside cavity by numerical model based on the motions and heat generation obtained via the quasi-dynamic model of ball bearing, and found that cage parameters are significant to air\u2013oil flow and thermal dissipation inside bearing cavity at ultra high rotation speed. In this paper, the influence of race conformities on the internal contact characteristics and ball motions in angular contact ball bearings is investigated based on the developed dynamic model [21], which is coupled with the EHD theory at each contact spot, the strategy for optimally determining the race conformities of ball bearing is further proposed. Fig. 1 shows the coordinate systems used in this study and velocities of a ball. The assumptions of the mathematical model are made as follows: The outer ring is fixed in housing; the mass center of each bearing element is coincident with its geometric center; the global deformations of the inner/outer ring are neglected, only the local deformations between balls and raceways are considered; the inner ring rotates around the bearing axis; the ball rotates around its center and bearing axis; translational motions of the centers of balls and inner ring are considered, except for the cage", + " The present dynamic model of angular contact ball bearings is similar to the one developed in previous research [21], with further considering the rolling friction of balls and thermal effect on film thickness, thus the entire model will not be repeated here, but the main equations to be solved will be recalled for clarity, besides the calculation of traction forces and moments. As shown in Fig. 2, the balls in the angular contact ball bearing are subjected to contact forces from outer and inner raceways, traction forces and moments, centrifugal forces due to orbital motion, interaction forces between balls and cage and the drag force from lubricant. As shown in Fig. 1, a ball in angular contact ball bearing not only rotates around the bearing axis with speed \u03c9c, but also rotates about its own center with speeds \u03c9x\u2032, \u03c9y\u2032, \u03c9z\u2032 around x\u2032, y\u2032 and z\u2032 axes in O-x\u2032y\u2032z\u2032 frame, respectively. According to Euler equations of motion and Newton equations, we can derive the differential equations governing the motion of the jth ball as follows: \u03c9\u0307 = \u2032\u2032 \u2212 \u2032\u2032 + \u2212 ( )I F R F R F d F d 2 2 1c jc x i x o m d m i o cage j j \u03c9 \u03c9 \u03c9( \u0307 \u2032 \u2212 \u2032) = ( \u2033 \u2212 \u2033) ( )I r F F 2j j y i y x jc y o he rolling elements" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001872_tmag.2019.2947611-Figure18-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001872_tmag.2019.2947611-Figure18-1.png", + "caption": "Fig. 18. Fields of the conventional HPDPMSM. (a) Flux line at rated torque. (b) Flux density at rated torque. (c) Flux density at maximum speed. (d) Flux line at maximum speed.", + "texts": [], + "surrounding_texts": [ + "As shown in Fig.3, The proposed bypass-rib is composed of two unconnected ribs and located in the front end of the flux barrier. The main parameters of the bypass-rib include the width wrib, length hrib and distance lrib. The influences of the bypass-rib on the average torque are investigated using 2D FEA method. The results are shown in Fig.4, 5 and 6. The average torque is slightly decreased when the width wrib and length hrib increasing. It can be seen, the length hrib has greater influence on the average torque than the width wrib, especially when the width wrib is greater than 0.5mm. According to the FEA results, the width wrib and length hrib should not be bigger than 0.7 mm and 1.4 mm respectively, about 25% and 50% respectively of the thickness of the PM. The average torque will decrease dramatically when the width wrib and length hrib exceed the recommended values. Fig.6 shows the average torque with the distance lrib of the bypass-rib. The average torque waveform in Fig.6 is plotted with the fixed width wrib and length hrib that are 0.6mm and 1.2mm respectively. The average torque is slightly changed when the distance lrib increasing. It is convenient to add the bypass-ribs to the flux barriers after the geometry design of the rotor." + ] + }, + { + "image_filename": "designv10_14_0000304_j.mechmachtheory.2018.05.006-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000304_j.mechmachtheory.2018.05.006-Figure2-1.png", + "caption": "Fig. 2. The 3-RFR class of SPMs: (a) the kinematically equivalent leg topologies (b) the virtual model.", + "texts": [ + " The main advantage of isostatic architectures is that they do not need the strict dimensional and geometric tolerances of overconstrained machines during manufacturing and assembly; they can thus work even when precise geometrical conditions are not accurately met, if at the expense of parasitic motions. Moreover, modular solutions characterized by three identical legs (symmetrical PKMs) are usually preferred for economic reasons. Three-legged isostatic SPMs include the class of mechanisms shown in Fig. 2 (b), whereby each leg is attached to the fixed and mobile bases by means of two revolute (R) pairs, whose axes intersect at the center of the spherical motion; in between, three lower pairs are equivalent to a planar joint (F) whose plane contains the axes of the revolute joints and the center of the motion, thus coinciding with the leg plane. This class of mechanisms was already studied by Karouia and Herv\u00e9 [16,17] , Di Gregorio [18] , Kong and Gosselin [19,20] and Zlatanov and Gosselin [21] among others. An interesting architecture of this class is sketched in Fig. 2 (b): it will be shown that, after optimization, this class can yield a large orientational workspace, characterized by a good dexterity around the home configuration, i.e., when the reference frame F 1 attached to the mobile platform (MP) has the same orientation as the reference frame F 0 fixed to the ground. This is one of the robot postures that allows the aligning of the unit vector e h normal to the MP plane with the vertical vector e v . The planar joint in the middle of each limb can be equivalently substituted by any of the PRR, RPR, PPR, PRP and RRR topologies, provided that some rules are observed: the directions of the prismatic joints must be parallel to the plane of motion, while the axes of the revolute pairs must be perpendicular to the plane, as illustrated in Fig. 2 (a). Finally, such joints may be merged to obtain cylindrical (C) and universal (U) joints to give a variety of leg topologies: ( RP )R( RR ) = CRU , ( RR )P( RR ) = UPU , ( RP )P( RR ) = CPU , ( RP )R( PR ) = CRC and ( RR )R( RR ) = URU . The synthesis of the kinematic chain has been driven by the goal of maximizing accuracy: To this end, the actuated joints have been chosen so as to obtain the coincidence of the singularity loci. Wrist configurations that cause singularities are identified and interpreted by means of screw theory [22] ", + " The study of these architectures is discussed in the subsection below. If the moments of the previous section are denoted by n i , i = 1 , 2 , 3 , whose directions are normal to the leg planes, then the singular configurations of the 3-RFR wrists with prismatic actuated joints are given by: n 1 \u00d7 n 2 \u00b7 n 3 = 0 (3) The vectors involved depend on the orientation of the MP and can be expressed as functions of the rotation matrix Q that describes the orientation of the MP, once a proper parametrization of the matrix has been chosen. To this end, by looking at Fig. 2 (b) it is noted that the unit vector of a i , the position vectors of A i denoted \u02c6 ai , for i = 1 , 2 , 3 , define the axes of a reference frame F 0 fixed to the ground, while the directions of the vectors d i , for i = 1 , 2 , 3 , define a reference frame F 1 fixed to the MP. The two frames share the same origin O , their relative attitude being given by the rotation matrix Q that maps a vector from the mobile frame F 1 to the fixed frame F 0 . The three vectors n i can then be expressed as n 1 = \u2212d 1 \u00d7 \u02c6 a1 = i 0 \u00d7 d j 1 \u21d2 [ n 1 ] 0 = d[ i 0 ] 0 \u00d7 [ Q ] 0 [ j 1 ] 1 n 2 = \u2212d 2 \u00d7 \u02c6 a2 = j 0 \u00d7 d k 1 \u21d2 [ n 2 ] 0 = d[ j 0 ] 0 \u00d7 [ Q ] 0 [ k 1 ] 1 n 3 = \u2212d 3 \u00d7 \u02c6 a3 = k 0 \u00d7 d i 1 \u21d2 [ n 3 ] 0 = d[ k 0 ] 0 \u00d7 [ Q ] 0 [ i 1 ] 1 (4) where d i is the vector connecting the origin of the reference frame F 0 with a point of the force line of action; among all possible choices for vector d i , the one of interest goes from the center of the spherical manipulator to the point B ", + " A more intuitive representation can be obtained as proposed by Corinaldi [27] upon expressing the four factors in Eq. (6a) in terms of the pointing direction e h of the mobile platform after the mapping of the rotation, i.e., in the form [ q ] 0 = [ e h ] 0 sin ( \u03b8/ 2 ) = Q ( r , r 0 )[ e h ] 1 sin ( \u03b8/ 2 ) (7) where \u03b8 is the angle of rotation of Q . A MP pose is indicated as a point on the surface of a three-dimensional sphere of unit radius by the vector q , having chosen [ e h ] 1 = [ e v ] 0 = \u2212\u221a 3 / 3[1 1 1] T as a reference vector fixed to the mobile reference frame, as illustrated in Fig. 2 (b). In the home configuration the two frames are coincident and [ e h ] 0 = \u2212\u221a 3 / 3[1 1 1] T , indicating the vertical direction also in the fixed frame F 0 . Vector q collects information of both the pointing direction e h and the angle of rotation \u03b8 . For example, the origin of frame F 0 represents the pose of the manipulator in the reference relative position between the mobile and the fixed frames, because it corresponds to a rotation through an angle equal to 0 around any direction. Vector [ q ] 0 = \u221a 3 / 3[1 1 1] T represents, instead, a rotation of \u03c0 around e v " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001087_s00332-018-9447-0-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001087_s00332-018-9447-0-Figure1-1.png", + "caption": "Fig. 1 A. E. Green\u2019s sketch of a twisted steel ribbon. First appeared in Green (1937); now in the public domain", + "texts": [ + "eywords Thin elastic sheets \u00b7 Energy scaling laws \u00b7 Wrinkling \u00b7 Microstructure Mathematics Subject Classification 74G65 \u00b7 74B15 \u00b7 49S05 \u00b7 49K20 \u00b7 74K20 Communicated by Felix Otto. This work was partially supported by the National Science Foundation through Grants OISE-0967140 and DMS-1311833. B Ethan O\u2019Brien eobrien2@andrew.cmu.edu Robert V. Kohn kohn@cims.nyu.edu 1 Courant Institute of Mathematical Sciences, New York University, New York, NY 10012, USA 2 Department of Mathematical Sciences, Carnegie Mellon University, Pittsburgh, PA 15213, USA Consider a thin elastic ribbon, clamped firmly on both ends and twisted into a helicoid, as shown in Fig. 1. A flat sheet is not related to a helicoid by an isometry, so we need to apply some force on the clamps in order to maintain the helicoidal shape. It is a curious fact that, for not too extreme values of the force, the ribbon develops smallscale wrinkles down the center, perpendicular to the axis of rotation. In fact, a twisted ribbon exhibits a range of morphologies: for example, if stretched with small enough force the ribbon forms triangular facets separated by creases, and with large enough force it might not wrinkle at all" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001824_j.yofte.2020.102379-Figure17-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001824_j.yofte.2020.102379-Figure17-1.png", + "caption": "Fig. 17. Schematic showing that (a) a normal tooth and (b) a faulty tooth of sun gear pass the strain measurement point of ring gear.", + "texts": [ + " 16 shows abundant details of the strain signal. Compared with the results in the previous researches [2,5,25,26], the result obtained by the proposed FBG interrogator has a higher quality. To verify the performance of the proposed FBG interrogation approach on the fault diagnosis of planetary gearbox, the tooth root strain signals of ring gear are measured under two conditions: a normal tooth of sun gear passes the measurement point; a faulty tooth of sun gear passes the measurement point, as shown in Fig. 17. The averaged strain signals with a length of four mesh cycles are shown in Fig. 18 where the fault feature can be clearly seen. An improved TFP-based edge-filter FBG interrogation approach has been proposed to measure the tooth root strain of ring gear in the planetary gearbox. First, the interrogation principle of the proposed approach is analyzed by comparing with the conventional edge-filter approaches. The influence of the creep of piezoelectric actuator on the interrogation results is discussed" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000194_s12206-015-0833-3-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000194_s12206-015-0833-3-Figure1-1.png", + "caption": "Fig. 1. Simplification method of ball-race contact.", + "texts": [ + " Thus, the stiffness of the bearing and its supporting structure can be taken into account relatively easily. When using the numerical method, the accuracy of the load distribution in the bearing is better. Thus, the numerical method is adopted in this study. The FE method is utilized to build the mechanical model of large rolling bearing. The contact between the rolling element and raceway is simplified by a nonlinear spring element with the same contact stiffness in order to save the calculation time. As an example, Fig. 1 shows the simplification method of ball-race contact in a four contact-point ball bearing. Each raceway has two arcs which contact the ball. In Fig. 1(a), arcs 1 and 2 are in the outer raceway. Arcs 3 and 4 are in the inner raceway. C1, C2, C3, C4 are the centers of curvature of arcs 1- 4, respectively. Cb is the ball center. When the bearing is in static equilibrium, the contact forces between the ball and arcs 1 and 4 are through C1, C4 and Cb. The contact forces be- tween the ball and arcs 2 and 3 are through C2, C3 and Cb. Thus, as illustrated in Fig. 1(b), contacts between the ball and arcs can be simulated by two traction springs with the same contact stiffness. Spring C1C4 simulates the contact between the ball and arcs 1, 4. Spring C2C3 simulates the contact between the ball and arcs 2, 3. To make spring forces directed along the normal to the arcs all the time, the arc center must move with its arc. Hence, two rigid beams are used to join centers with the corresponding arcs, which are shown in Fig. 1(c). In this way, the contact force always passes through the curvature center of each arc. The intersection angle between the spring and the bearing radial plane is the contact angle [28]. Fig. 2 shows an actual FE model of a four contactpoint ball bearing. In the FE model, the ball-race contact is simplified by this method. The ends of rigid beams are glued to the nodes of raceway. The simplification is used for all of the balls in the bearing. When the loads and constraints applied on the bearing are determined, the distributions of contact forces, including the maximum contact force Qmax, and angles in the bearing, can be calculated" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001524_j.wear.2019.203106-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001524_j.wear.2019.203106-Figure1-1.png", + "caption": "Fig. 1. Gear test device.", + "texts": [ + " Several gear test devices have been developed to understand the behaviours of gears under operating conditions. These devices enable monitoring of gear performances under variable torques and rotation speeds, and also allow precautions to be taken against gear damages that may occur. In this study, a new test device has been developed for gears of 1 module. The test device comprises of an AC servomotor, a torque sensor, a gearbox, couplings, a brake mechanism, and various machinery parts (see Fig. 1). The movement is transferred from the AC servomotor to the input torque meter via a flexible coupling, and then to the gearbox with another flexible coupling. The connection to the output torque meter is established with a flexible coupling at the end of the gearbox. The brake mechanism, which represents the load, is connected to the output torque meter again with a flexible coupling. The desired movement is provided to the motor via the servomotor driver and the desired load level is adjusted using the brake mechanism" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001573_j.mechmachtheory.2018.12.031-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001573_j.mechmachtheory.2018.12.031-Figure1-1.png", + "caption": "Fig. 1. Structural diagrams: (a) diagram of a parallel mechanism, and (b) diagram of a limb.", + "texts": [ + " When h is an infinite value, the corresponding screw is expressed as \u03c1$ = \u03c1 ( S S 0 ) = \u03c1 ( 0 S ) for infinite h (2) which can represent a couple or a translation. For a twist of a rigid body $ 1 = ( S 1 ; S 0 1 ) T and a wrench exerted on the body $ 2 = ($ 2 ;$ 0 2 ) T , if they satisfy the condition $ 1 \u25e6 $ 2 = S 1 S 0 2 + S 2 S 0 1 = 0 (3) where \u201c\u25e6\u201d denotes the reciprocal product, $ 2 is a constraint wrench. Otherwise, $ 2 is an actuation wrench [36] . 2.2. Expression of the end joint reaction A parallel mechanism consists of several limbs connecting a moving platform to a fixed base, shown in Fig. 1 (a). For limb i shown in Fig. 1 (b), if the limb can output n degrees of freedom (DOFs), it has 6- n constraint wrenches $ c i, 1 , \u00b7 \u00b7 \u00b7 , $ c i, 6 \u2212n based on screw theory [5] . Without loss of generality, limb i includes an actuation pair, so it has only an actuation wrench $ a i . Here, a joint connecting a limb to the moving platform is called the end joint of the limb. If the end joint of limb i has g DOFs, the reaction of the joint has 6- g independent components. First, $ a i and $ c i, 1 , . . . , $ c i, 6 \u2212n can be chosen as independent components" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000769_978-3-662-43893-0-Figure6.7-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000769_978-3-662-43893-0-Figure6.7-1.png", + "caption": "Fig. 6.7 Geometry of a stick blade. \u03b3K top relief angle \u03b7u hook angle \u03b4 inclination angle of the stick blade in the cutter head", + "texts": [ + " 2.1). On the other hand, machines with a two-part tool spindle for two-part cutter heads were also developed. On these machines one spindle centre could be offset relative to the other, allowing an appropriate difference in tool radii for lengthwise crowning. One of these manufacturing methods is termed Zyklo-Palloid\u00ae and was introduced in 1955 (see Sect. 2.5). The other is known under the name of Kurvex. The first stick blade cutter heads for high volume production were introduced in 1967 (see Fig. 6.7). 6.2 Cutting of Spiral Bevel Gears 235 For decades, all these gear cutting machines had extremely complex mechanics and mechanical drive chains. Since the middle of the 1980s, it has been possible to replace these high-precision drive chains by numerically controlled drives, known as \u201celectronic gearboxes\u201d. Later, a coordinate transformation was used to convert the up to ten setup- andmotion-axes of the mechanical machines into the kinematics of a CNC machine with six axes. Generally speaking, all six axes move simultaneously in a timed relationship when cutting spiral bevel gears (see Sect", + " The avoidance of a common divider between the number of blade groups and the number of teeth of the work piece may also be the deciding factor. The width of the blade shank is chosen on the basis of the required profile height and the stiffness of its cross-section. Blade shanks which are too wide have an unused shoulder and increase the cost of the blank, incurring additional grinding costs and unnecessary wear on the grinding wheel. On a stick blade, the surfaces which form the blade profile are ground to match the blade shank (see Fig. 6.7). For regrinding, the stick blades are removed from the cutter head and the entire blade profile is reground in the direction of the blade shank. These stick blades are mounted in the cutter head again and adjusted to the defined point height using a cutter head setup device. In contrast to profile blades, no compensation of the cutting depth is necessary. The cutting edge geometries of stick blades are differentiated according to the number of profile surfaces to be reground: \u201ctwo-face grinding\u201d (2F) or \u201cthree-face grinding\u201d (3F)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000423_0954405414553979-Figure13-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000423_0954405414553979-Figure13-1.png", + "caption": "Figure 13. Initial tool path for conical cutter.", + "texts": [ + " Besides, the ball end of the cutter should be tangential to the hub surface during the machining process. The tolerance e is set to be 0.005mm. The ball-end center of the cutter can be determined with the dichotomy method, and the CL can be calculated using the methods in Bedi et al.4 and Tang et al.23 Thirty tool axes are generated as the initial CLs, and the initial tool axis trajectory surface is obtained with the interpolation method. When the conical cutter is applied, the initial undercut and overcut are 0.0256 and 0.00754mm, respectively. The initial tool path is shown in Figure 13. The SLP method is applied to solve the constraint optimization problem. The convergence process is depicted in Figures 14\u201317. As shown in Figures 15 and 16, the interference of the cutter with the adjacent blade occurs when the cutter bottom radius is 3.14mm and half taper angle is 6:128. Therefore, a conical cutter with a bottom radius of 3mm and half taper angle of 68 is selected as the optimal cutter. Its stiffness is 2:43e+4 N=mm. The tool path is then generated by solving problem P4. After optimization, the undercut and overcut become 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001644_rnc.4758-Figure7-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001644_rnc.4758-Figure7-1.png", + "caption": "FIGURE 7 Platform of flexible wing unmanned aerial vehicle (UAV). A, Powered plant; B, System mechanical structure [Colour figure can be viewed at wileyonlinelibrary.com]", + "texts": [ + " In the HITL simulation, the rotation speed of the propeller can be received by the optoelectronic switch. It will also be sent to Matlab software by the serial port. In this way, it has the feedback of the horizontal and vertical control quantities in the HITL simulation. The real-time actual control quantity is obtained by the physical device. It not only can test the performance of the hardware but also can simulate the flight state of the system more realistically. The powered plant and the system are shown in Figure 7. In the platform, the horizontal control quantity is acquired by the optical electricity encoder and the vertical control quantity is obtained by the photoelectric inductive switch. During the real-time experiment, all the control quantities will be sent to the Matlab software and embedded control unit. By the HITL simulation, the platform can maximally simulate the flight state of the flexible wing UAV. The process of the HITL simulation is shown in Figure 8. In the model, the parameters of the flexible wing UAV are designed as a certain type of the parafoil, the significant physical parameters are all shown in Table 1" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000894_tro.2015.2489499-Figure21-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000894_tro.2015.2489499-Figure21-1.png", + "caption": "Fig. 21. Hidden robots involved in the tested visual servoings of the 3\u2013PRS robot (projection in the yz plane \u2013 R and S joints at Ai and Bi , respectively, are drawn with the same symbol for the sake of clarity of the drawing). (a) When all leg directions ui are observed (Case 1): a 3\u2013PRS robot. (b) When all leg directions ui and the Plu\u0308cker coordinates of the line passing through the legs 1 and 2 are observed (Case 2): a {2\u2013PRS}\u2013PRS robot. (c) When all leg directions ui and the Plu\u0308cker coordinates of the lines passing through the legs 1 and 3 are observed (Case 3): a 3\u2013PRS robot.", + "texts": [ + " These singularities represent some workspace boundaries. Type 2 singularities are more complex and are studied in [49]. 2) Analysis of the Possible Hidden Robot Models: Case 1: Let us now assume that we want to control the 3\u2013PRS robot depicted in Fig. 20 by using the observation of its leg directions ui (see Section II). From Section III, we know that using such a control approach involves the appearance of a hidden robot model. This hidden robot model can be found by straightforwardly using the results of Section III and is a 3\u2013PRS robot shown in Fig. 21(a). This robot is known to be architecturally singular (it can freely move along the z-axis) and cannot be controlled by using only the observation of its leg directions ui . Case 2: As a result, one would logically wonder what should be the necessary information to retain in the controller to servo the robot. For instance, let us use the Plu\u0308cker coordinates of the line passing through the axis of the cylinder (see Section IVD), i.e., the direction and location in space of this line. Let us consider that we add this information for the estimation of the legs 1 and 2 positions. Modifying the hidden robot model according to Fig. 16(a), the corresponding robot model hidden in the controller is depicted in Fig. 21(b): This is a {2\u2013PRS}\u2013 PRS robot, which is not architecturally singular. In other words, using the Plu\u0308cker coordinates of the line for legs 1 and 2 involves to actuate both the first P and R joints of the corresponding legs, i.e., the virtual legs are PRS legs. For the {2\u2013PRS}\u2013PRS robot, it is possible to prove that two assembly modes exist. Indeed, for this robot, when fixing the position of points B1 and B2 (which is the case when actuating the P and R joints of the legs 1 and 2), the platform can freely rotate around (B1B2)", + " Thus, the robot is not fully controllable in its whole workspace. Case 3: From the result that, using the Plu\u0308cker coordinates of the line passing through the axis of the cylinder, the leg of the virtual robot becomes a PRS leg, it is possible to understand what is the minimal set of information to provide to the controller to fully control the robot in the whole workspace: we need to use all the Plu\u0308cker coordinates of the lines passing through legs 1 to 3. In such a case, the hidden robot model is a 3\u2013PRS robot depicted in Fig. 21(c). It is possible to prove that this robot has no Type 2 singularity and can freely access to its whole workspace. 3) Simulation Results: Simulations are performed on an Adams mockup of the 3\u2013PRS robot with the following values for the geometric parameters: l = 0.5 m, d = 0.4 m, R = 0.1 m. This virtual mockup is connected to MATLAB/Simulink via the module Adams/Controls. The controller presented in Section II is applied with a value of \u03bb assigned to 20. The initial configuration of the robot end-effector is z0 = 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000894_tro.2015.2489499-Figure10-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000894_tro.2015.2489499-Figure10-1.png", + "caption": "Fig. 10. UraneSX robot and its hidden robot leg. (a) Schematics of the architecture: a 3\u2013PUU robot with the three actuated P joints in parallel. (b) Its hidden robot leg: a PUU leg; thus, the hidden robot is a 3\u2013PUU robot with the three passive P joints in parallel leading to an uncontrollable translation along the P joints direction.", + "texts": [ + " It is obvious that for robots with legs whose directions are constant in the whole workspace, it is not possible to estimate the platform pose from the leg directions only; 6[PP] means an active planar chain able to achieve two dof of translation, such as PP or RR chains. 2) robots with legs whose directions are constant for an infinity of (but not all) robot configurations: this is the case of PRRRP robots with all P parallel [see Fig. 9(a)] and of Delta-like robots actuated via P joints for which all P are parallel (such as the UraneSX (see Fig. 10) or the I4L [39], [40]). It was shown in [16] through the analysis of the rank deficiency of the interaction matrix that it was not possible to control such types of robots using leg direction observation. Considering this problem with the hidden robot concept is very easy. For example, in the case of the PRRRP robot with parallel P joints, the hidden robot has a PRRRP architecture [see Fig. 9(b)], where the parallel P joints are passive. This robot is well known to be architecturally singular as there is no way to control the translation along the axis of the parallel P joints. This result can be easily extended to the cases of the hidden robots of the UraneSX and the I4L (see Fig. 10). 3) robots with legs whose directions vary with the robot configurations but for which all hidden robot legs contain active R joints but only passive P joints: the most known robot of this category will be the planar 3\u2013PRP robot for which the hidden robot model is a 3\u2013PRP, which is known to be uncontrollable [25], [37] (see Fig. 11). The hidden robot model can be used to analyze and understand the singularities of the mapping and to study if a global diffeomorphism exists between the space of the observed element and the Cartesian space" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001687_j.engfailanal.2020.104411-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001687_j.engfailanal.2020.104411-Figure3-1.png", + "caption": "Fig. 3. Gear Wear Test Machine [9]", + "texts": [ + " One drawback of additive manufacturing is limits on the maximum size of parts that can be produced, which is also applicable to the EOS device. Yet because the gear size examined for this study was small, no problems were encountered. Table 2 shows the expected properties of 420 steel materials according to the device\u2019s production parameters. Wear testers for a module of 1 mm gears consisted of AC servomotors, torque sensors, a gearbox, couplings, and brake mechanisms, along with various other mechanical parts. Components of the construction are shown in Fig. 3. An AC servomotor is controlled by its driver and provides control at the desired speed. The movement from the AC servomotor is transmitted first to the input torque meter via the flexible coupling and then to the gearbox by a second flexible coupling. At the output of the gear unit, the movement is transmitted to the output torque meter by a third flexible coupling, and then to the brake mechanism representing the load by the fourth flexible coupling. The servomotor drive allows the motor to move at the desired speed while the brake mechanism adjusts the load level" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001925_978-3-030-46817-0_30-Figure20-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001925_978-3-030-46817-0_30-Figure20-1.png", + "caption": "Fig. 20. FEM model - stress distribution.", + "texts": [], + "surrounding_texts": [ + "Control device MR 96 is the horizontal testing machine suitable for tensile test of long materials such as steering tie rods. The tie rod is loaded in tension. The tensile test on three tie rod sample, of 21 mm inner diameter and 26 mm outer diameter and 329 mm of length, have provided the following results presented in Table 4. During the tensile test, all results were recorded. The tensile force is calculated using: Ft = A\u2219E\u2219et, where:" + ] + }, + { + "image_filename": "designv10_14_0000018_cca.2015.7320797-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000018_cca.2015.7320797-Figure2-1.png", + "caption": "Fig. 2. 3D representation of the fixed-wing UAV on XFLR5", + "texts": [ + " They are separated for the sake of clarity in the equations, even if in practice they are mixed such that the left and right elevons deflections are respectively: \u03b4l = \u03b4a \u2212 \u03b4e, \u03b4r = \u2212\u03b4a \u2212 \u03b4e (12) The lift, lateral and drag coefficients are respectively given by: CL = CL0 + CL\u03b1\u03b1+ CLq q Va + CL\u03b4e \u03b4e (13) CY = CY\u03b2\u03b2 + CYp p Va + CYr r Va + CY\u03b4a \u03b4a (14) CD = CD0 + CDCL C2 L (15) Similarly, the moment coefficients are given by: Cl = Cl\u03b2\u03b2 + La Va ( Clpp+ Clrr ) + Cl\u03b4a \u03b4a (16) Cm = Cm0 + Cm\u03b1\u03b1+ Lo Va Cmqq + Cm\u03b4e \u03b4e (17) Cn = Cn\u03b2\u03b2 + La Va ( Cnpp+ Cnrr ) + Cn\u03b4a \u03b4a (18) where CL0 , CL\u03b1 , CLq , CL\u03b4e , CY\u03b2 , CYp , CYr , CY\u03b4a , CD0 , CDCL , Cl\u03b2 , Clp , Clr , Cl\u03b4a , Cm0 , Cm\u03b1 , Cmq , Cm\u03b4e , Cn\u03b2 , Cnr , Cnp and Cn\u03b4a are the stability derivatives that can be determined using computational fluid dynamics tools. One of the most commonly used is the open-source software XFLR53, which works at low Reynolds number. It requires to create a 3D representation of the aircraft, as depicted in Fig. 2. The wing is designed using the geometric characteristics of the UAV provided by the manufacturer, and the different airfoil profiles coordinates are available in very complete databases4. Once the 3D model of the UAV is obtained, XFLR5 then allows to use different algorithms to estimate the stability derivatives, such as the Lifting Line Theory, the Vortex Lattice Method or the 3D Panel Method. Remark 2: Note that other open-source softwares exist to estimate the stability derivatives, such as AVL5 or Tornado6 for Matlab" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000896_j.talanta.2015.12.037-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000896_j.talanta.2015.12.037-Figure3-1.png", + "caption": "Fig. 3. CL signal intensities obtained for six PCs using the optimized conditions for", + "texts": [ + " The proton NMR spectrum of R6G in the presence of SDS compared to that of R6G in D2O (with no SDS) showed that some of the aromatic protons of R6G were shifted upfield in the former, indicating that these protons were positioned in the hydrophobic cavity of SDS (Fig. S4). It is worth mentioning here that CL-based affinity methods have also been used for direct determination of host\u2013guest inclusion interactions and has great potential towards becoming a reliable method for quantitatively studying host\u2013guest binding [18]. Using SDS as a medium for dissolving R6G, the other parameters that were previously described for the Ce(IV)\u2013RB CL system were also optimized. Table 1 shows the optimized conditions for both systems, and Fig. 3 shows the CL signal intensities obtained for six PCs using the optimized conditions for the Ce(IV)\u2013RB system in aqueous media and the Ce(IV)\u2013R6G system in the presence of SDS. The results indicated that the Ce(IV)\u2013R6G CL system exhibited a better sensitivity. The figure (Fig. 3) also shows the CL signal intensity of the PCs with Ce(IV)\u2013R6G in aqueous media. The CL signal enhancement for the five PCs varied between 30% and 300% in the presence of SDS compared to that in water (with no surfactant), while the signal for GA was suppressed by only 30%. Using the optimum conditions shown in Table 1, the calibration curves for GA, CA, SA, RUT, QRC and CAT were obtained. These curves were established with a series of standard solutions that contained each analyte. Ten solutions with different concentrations of each analyte were used, and linear regression curves were obtained for all of the studied analytes" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000346_ceit.2018.8751858-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000346_ceit.2018.8751858-Figure3-1.png", + "caption": "Fig. 3. Mechanical structure of LTS with two gear motors on UAV.", + "texts": [ + " In this study, six-rotor UAV is used and the design of the LTS is carried out in accordance with the six-rotor UAV. The LTS used in this study is designed to leave three payloads to real-time specified targets. Also, the UAV used in this study moves autonomously, and the LTS's payload transporting and leaving to real-time specified targets is also performed autonomously. The front and side view of the designed LTS in simulated environment are shown in Fig. 1, and the front view of the realized LTS on the six-rotor UAV is shown in Fig. 2. The LTS which has mechanical structure shown in Fig. 3 is mounted on the bottom surface of the six-rotor UAV body so that the swing of payloads is reduced during payload transferring and leaving. The two DC geared motors specified in Table I are used in LTS. The mathematical model of the DC geared motor given the parameters in Table I is used as the reference model in MRAC. This section describes the control strategy design for LTS on six-rotor UAV. In the MRAC which has been shown block diagram in Fig. 4, the reference output is obtained using the mathematical or system identification model (reference model) of the system to be controlled" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002440_s11771-020-4537-1-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002440_s11771-020-4537-1-Figure1-1.png", + "caption": "Figure 1 BSM global-contact coordinate system and transformation relationship between coordinates O-XYZ, Ob-TNB, S\u2212xsyszs and N\u2212xnynzn", + "texts": [ + " In order to update the study of precision loss, the BSM sliding-rolling mixed motion behavior considering given axial loading and rotational speed working conditions was studied in this paper. Furthermore, the BSM precision degradation and BSM precision loss rate were also gained in the light of the sliding-rolling mixed motion behavior in this study. 2 Ball screw mechanisms creep rate Creep analysis is the basis for resolving the problem of the BSM rolling contact, including the BSM rolling friction [13] and the BSM rolling contact wear [32, 33]. This paper uses the BSM coordinate system shown in Figure 1 [13, 22]. O-XYZ is the global coordinate system, and Ob-TNB is the free coordinate system. Both S\u2212xsyszs and N\u2212xnynzn are local coordinate systems. For the meaning of other parameters in the figure, refer to the Nomenclature. The creep rate of ball and nut along the xn and yn direction, i.e., \u03c2xn and \u03c2yn, according to KALKER\u2019s linear creep [31], is expressed as follows: n b n N n b Nx x x (1a) n b n N n b Ny y y (1b) where b nx and b ny are the components of ball spin angular velocity \u03c9br in the xn and yn directions, respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001148_s0373463318000668-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001148_s0373463318000668-Figure1-1.png", + "caption": "Figure 1. The coordinate system.", + "texts": [ + " Simulation results are given in Section 4 and discussed. The work is summarised in Section 5. 2. THE MULTI-AUV SYSTEM PROBLEM. An underwater vehicle has six degrees of freedom of motion in space. The kinematics problem is described by setting up a coordinate system. 2.1. The coordinate system. In order to make the motion analysis of the vehicle clearer and simpler, this paper uses two coordinate systems to describe the kinematics problems of an underwater vehicle, an inertial coordinate system and a vehicle coordinate system. Figure 1 depicts the coordinate systems (Sun et al., 2016). Inertial coordinate system E \u2212 \u03be\u03b7\u03b6 is also called the Earth coordinate system or fixed coordinate system, and its origin is a certain point on Earth. Vehicle coordinate system O\u2013xyz is fixed on the AUV and moves with the AUV. The vector \u03b7 = [ x y z \u03c6 \u03b8 \u03c8 ]T is used to describe the position and attitude of the AUV in inertial coordinates, and the vector \u03bd = [ u v w p q r ]T describes the linear and angular velocity of the AUV in the vehicle coordinates" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001348_iet-smt.2015.0146-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001348_iet-smt.2015.0146-Figure1-1.png", + "caption": "Fig. 1 Experimental rig used in this study", + "texts": [ + " When a ball bearing is installed in a motor, most of the load carried is axial, but there is also a radial load and inner ring radial thrust. Speed should also be considered. All four factors contribute to vibration signals generated during operation. This paper 1 a Schematic b Actual test rig analyses and optimises these signals by combining the four elements within the rig. A schematic diagram of the rig and an illustration of the actual rig used in the experimental work carried out in this study are shown in Fig. 1. Table 1 shows the forces that can be investigated. To avoid interaction between the motor and the bearing under test, a chain drive is used to connect the motor to the test-rig shaft. The rig has a brake and a pneumatic cylinder to apply an axial load, a radial load and radial thrust. Detailed rig specifications are shown in Table 2. The test bearing used in this paper was an SKF 6204 bearing. Round thru-holes of diameters 1.5, 2 and 2.5 mm in the inner and outer rings were made using an electrical discharge machine" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000731_s11661-018-4914-7-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000731_s11661-018-4914-7-Figure2-1.png", + "caption": "Fig. 2\u2014Schematic representation of the molten pool morphology in the (a) transverse section and (b) longitudinal section with five geometrical parameters.", + "texts": [ + "5 mm below the nozzle tip, and scans along the positive X axial direction at a constant speed Vb. Metal powders are fed into the laser spot region through the annular conical channel of coaxial nozzle. In the powder stream, the particles travel through the laser irradiation zone in which they are rapidly heated up. Three horizontal planes in different distance from coaxial nozzle tip are later used to discuss the laser-powder interaction process. In this coaxial LPD mathematical model, the transient evolution the molten pool morphology corresponds to the temperature field. As shown in Figure 2, five geometrical parameters (W, D, H, L, a) measured from the simulation results are used to describe the molten pool morphology, with W being the maximum width, D being the maximum melting depth,H being the maximum height, L being the maximum length of the molten pool and a being the angle between the horizontal line and the tangent line of the fusion interface. The distribution of powder flow affects laser-powder interaction process and the resultant temperature of powder particles, which influences the laser energy balance between powder flow and substrate", + " Therefore, like the variation of peak temperature with the tracks, peak temperature of each layer increases with the layer numbers and finally reaches steady state with more layers. The molten pool morphology directly relates to the distribution of temperature field. The molten pool morphology, especially the lifted portion in the overlapping region, has a predominant effect on the thermal gradient distribution which directly affects the crystal growth in the molten pool. The asymmetric extent of molten pool can be described by the angle a shown in Figure 2. Figure 7 shows the molten pool morphology under different overlapping ratios. The dotted dark line in Figure 7(a) represents the fusion boundary of simulated molten pool, and the red-dotted line in Figure 7(b) represents the measured fusion boundary of experimental results. As shown in Figure 7(a), the simulated molten pool morphology of the first track in the first layer is symmetric in the transverse section with the angle a = 29 deg, and keeps stable with the variation of the overlapping ratios" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001939_0142331220928887-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001939_0142331220928887-Figure1-1.png", + "caption": "Figure 1. The motion variable of ship.", + "texts": [ + " Simulation results are exhibited in Section 5 to verify the theoretical results obtained in the previous section. Finally, the conclusion is remarked in Section 6. Kinematics. In this paper, considering the small roll angle (f) and pitch angle (u) of the DP ship, the kinematics model is simplified into three-degree-of-freedom expressions including surge, sway and yaw (Bian, 2011). Its equation form is _h=R(c)y \u00f01\u00de where R(c)=RZ,c = \u00bdcos (c), sin (c), 0 ; sin (c), cos (c), 0 ; 0, 0, 1 represents rotation matrix obtained from rotation of Z axis and y = \u00bdu, v, r T , h= \u00bdx, y,c T as seen in Figure 1. hd is defined as the expected location information of the system, so the tracking error he can be expressed by expression (2). he =h hd \u00f02\u00de whereh,hd ,he 2 R2 3 S: Dynamics. According to the momentum theorem, we can get the dynamic vector expression of the DP ship. Its equation form is MRB _y +CRBy = tRB \u00f03\u00de where MRB represents the inertia matrix of the rigid body system, y = \u00bdu, v, r T represents the component of the velocity vector of each degree from the body-fixed coordinate frame. CRB represents the Coriolis matrix of the rigid body, and tRB = \u00bdX ,Y ,N T represents the vector of the external force and moment of the DP ship" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002206_s00202-021-01338-x-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002206_s00202-021-01338-x-Figure3-1.png", + "caption": "Fig. 3 AFPM models with different magnet shapes: a trapezoidal, b skewed, c triangular, and d rounded-triangular", + "texts": [ + " The output power rating of the machine is 3 kW, with a rated speed of 450 rpm. The machine operates at a 60 Hz supply frequency. The design parameters of the machine are calculated using the improved D2L process, which results in 16 rotor magnetic poles and 12 stator coils. Using the design parameters defined in Table 1 and Table 2, threeAFPMSGmodels are designed,with all dimensions and material properties identical except for the magnet shape. The material used for the rotor is steel, and each rotor houses 16 surface-mounted Nd-Fe-B permanent magnets. Figure 3 presents the investigated AFPMSG models for the different rotor magnet shapes. These machine models employ a coreless stator consisting of twelve concentrated and non-overlapping coilswith a trapezoidal shape characteristic. Although a variety of coil topologies are possible, such as trapezoidal, circular, hexagonal, and rhomboidal, previous research results have shown that trapezoidal coils make the most effective use of the magnetic field due to their large active area and, for this reason, trapezoidal coils are selected for this study", + " 6 provides a comparison of the phase back-EMF of the models. As shown in Fig. 6, for the significant part of the period, the back-EMF of the triangular magnet model is higher than the other three models. Also, the model with trapezoidal magnets has a higher peak. The waveforms of the back-EMFs are not purely sinusoidal; this is due to the uneven magnet flux distribution around the machine. The higher back-EMF for the triangular magnets is because theflux leakage in thismodel is lower than the others. Figure 3 shows that there is a difference between thewidths of the magnets around the inner radius of the machine. The triangular magnets have a larger distance between each other, which reduces the leakage flux between adjacent magnets, thus increasing the back-EMF. To have a better comparison, the RMS of the back-EMF for all four analyzed models are presented in Table 3. A comparison of the harmonic components of the back EMF for the three magnet designs is shown in Fig. 7. It can be seen that the fundamental harmonic of the induced EMF in the model with triangular magnets is the maximum (173" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000302_b978-0-12-813691-1.00006-3-Figure6.22-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000302_b978-0-12-813691-1.00006-3-Figure6.22-1.png", + "caption": "FIGURE 6.22", + "texts": [ + " Modification of these electrodes with CNMs (Fig. 6.21) can enhance their conductivity and electrocatalytic property toward the electrochemical behavior of different compounds and pharmaceutical compounds. In some of the electrochemical techniques, movement of the electrode with respect to the solution is required. In such systems, the electrode is itself in motion (e.g., rotating disks, rotating wires, vibrating electrodes, etc.) or the solution forcedly flows past a stationary electrode. The rotating disk electrode (RDE) (Fig. 6.22) is made of a disk of electrode material (e.g., glassy carbon or platinum) embedded in a rod of an insulating material (e.g., glass, Teflon, epoxy resin, or another plastic). The electrode is attached to a motor and rotated at a certain frequency, f5\u03c9/2\u03c0, where \u03c9 is the angular velocity (s21). Electrical contact is made to the electrode by means of a brush contact (e.g., carbon-silver) (Bard and Faulkner, 2001). RDEs are mostly used in hydrodynamic techniques, for example, hydrodynamic amperometry or voltammetry" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001683_tia.2020.2968036-Figure33-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001683_tia.2020.2968036-Figure33-1.png", + "caption": "Fig. 33. Prototype SPMSM. (a) Stator core. (b) Rotor. (c) Windings. (d) Motor.", + "texts": [ + " However, the SPMSM suffers from higher demagnetization risks with the increasing temperature because Although medium to large size SPMSMs have much smaller resistance than the small size ones, the rules in the small size SPMSM are also applicable in medium to large size machines. In order to verify the theoretical analysis by FE method coupled with the motion equation, experiments of 3PSC with a GLT load are conducted on a 12-slot/10-pole prototype SPMSM. The details of the prototype SPMSM are shown in Fig. 33 and the parameters are listed Table III. In order to conduct the experiments in a controllable manner, Sm2Co17 is selected as the PM material, which has a strong demagnetization withstand capability. The back-EMF is shown in Fig. 34 and the measured results matches well with the FE results. The experiments are simplified as a weight driving the shortcircuited prototype SPMSM from the static state. The test rig is shown in Fig. 35. When the 3PSC starts, the weight is released and drives the short-circuited SPMSM through a pully and string" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001597_s11012-019-00996-3-Figure5-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001597_s11012-019-00996-3-Figure5-1.png", + "caption": "Fig. 5 Hammering test of the bearing system. a the experiment setup, b the structure sketch", + "texts": [ + " According to Refs [19, 31], it can be assumed that the damping c is in proportional to the stiffness k\u00f0x; t\u00de. Therefore, the damping c is given as:c \u00bc n k\u00f0x; t\u00de, in which n is the damping ratio. The exact expression of the damping ratio n is given as: [32] where wa is the resonance frequency of the acceleration response,w1 and w2 are the frequencies at the half-power bandwidth points. The instant hammering method has been employed to measure the system frequency response. The experiment setup and structure sketch of the bearing system are shown in Fig. 5a, b, respectively. In Fig. 5b, the bearing pedestal 1 installed with two 7206B bearings was regarded as the front-end support bearing unit and the bearing pedestal 2 installed with a 6206ZZ bearing was regarded as the rear-end support bearing unit. The 6206ZZ bearing had no effect on the system axial stiffness as a result of the shaft that was fixed at frontend and freed at rear-end. The vibration signals extracted from the acceleration sensor (PCB 352C04) were processed by the signal analysis system (DH 5956). Then, as shown in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001974_j.triboint.2020.106606-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001974_j.triboint.2020.106606-Figure1-1.png", + "caption": "Fig. 1. A dynamic model of a gear systems.", + "texts": [ + " The reliability of lubrication is evaluated by the advanced first-order second-moment method. Lastly, an example of spur gear pair under random excitation with different intensity is simulated, the effect of noise intensity on numerical characteristics of the dynamics response and lubrication reliability are discussed. The lubrication reliability of the gear pair under different rotation speed and roughness are also discussed. To predict the meshing force on the gear tooth, a torsional dynamic model is established. As shown in Fig. 1, the gear pair is simplified as two discs and the interaction of the gear teeth is simplified as a spring and a damping. The dynamics equations, derived by Newton\u2019s laws, are as follows: { I1\u03b8\u03081 + Rb1c(Rb1\u03b8\u03071 \u2212 Rb2\u03b8\u03072 \u2212 e\u0307) + Rb1k(t)(Rb1\u03b81 \u2212 Rb2\u03b82 \u2212 e) = T1 I2\u03b8\u03082 + Rb2c(Rb1\u03b8\u03071 \u2212 Rb2\u03b8\u03072 \u2212 e\u0307) \u2212 Rb2k(t)(Rb1\u03b81 \u2212 Rb2\u03b82 \u2212 e) = T2 (1) where Ii is the rotary inertia, Rbi denotes the base radius, \u03b8i is the torsional angular displacement, i = 1, 2 denote pinion and gear, respectively, T1 and T2 represent the input and output torque" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000765_1.4966628-Figure6-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000765_1.4966628-Figure6-1.png", + "caption": "Fig. 6. A catenary parameterized by r\u00f0a\u00de \u00bc \u00f0Ra;R\u00f02 cosh a\u00de\u00de. Here, a0 \u00bc cosh 1\u00f02\u00de 1:317.", + "texts": [ + " 2, February 2017 Gonz alez-Cataldo, Guti errez, and Y a~nez 111 In this case, the point of departure is satisfied when v2 0 gR \u00fe 2 sin2u \u00bc 2 cos2u; (37) and therefore cos 2u\u00f0 \u00de \u00bc v2 0 2gR : (38) Notice that v0\u00bc 0 implies a departure point of u0 \u00bc p=4 \u00f0a0 \u00bc p=2\u00de; the slope at this point (u0 \u00bc 45 ) is slightly smaller than in the case of the frictionless circle, where u0 48 . Figure 5 shows the evolution of v2\u00f0u\u00de=gR as it approaches the horizon of the cycloid, H\u00f0u\u00de \u00bc 2 cos2u (compare Fig. 3). We can parameterize the catenary by r\u00f0a\u00de \u00bc \u00f0Ra;R\u00f02 cosh a\u00de\u00de, where a is a real, positive number, which represents the curve shown in Fig. 6. Under this parameterization, t\u0302 \u00bc sech a x\u0302 tanha y\u0302; which leads to tan u \u00bc sinh a: (39) From this relationship, the curvature is found to be j u\u00f0 \u00de \u00bc 1 R cos2u; (40) and using this curvature in Eq. (20), the speed of the particle sliding on this surface becomes v2 u\u00f0 \u00de gR \u00bc v2 0 gR e2lu \u00fe 2e2lu \u00f0u 0 sin u0 l cos u0\u00bd e 2lu0 cos2u0 du0: (41) Unfortunately, the integral here cannot be performed in terms of elementary functions. Meanwhile, Eq. (21) gives the horizon as H\u00f0u\u00de \u00bc sec u; (42) which bounds the initial speed at v2 0 gR " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002175_tie.2021.3084172-Figure14-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002175_tie.2021.3084172-Figure14-1.png", + "caption": "Fig. 14. Prototype and test bench. (a) Diagram of the test bench system. (b) Test bench.", + "texts": [ + " The post-fault torque capability indicator, i.e. k1, increases with the increase of current amplitude, which means higher post-fault torque coefficient can be obtained with larger amplitude of reconstructed currents under SC fault. k2 is less than zero for all the post-fault conditions, which means that the reconstructed armature MMF contains demagnetized MMF with the proposed FTC strategy. V. EXPERIMENTS A. Test Bench Description Experiments are conducted to verify the analysis results. Prototype and test bench are shown in Fig. 14. A five-phase H-bridge inverter is used to drive the machine, and the load machine which is controlled by commercial driver operates with speed control. Voltages and currents are measured by power analyzer with bandwidth of 50 kHz. First, no-load back EMF voltage waveforms are measured, as shown in Fig. 15. The measured waveforms agree well with FEM calculated ones, and THD of the measured back EMF voltages is 3.52%. Then experiments are carried out in the following conditions: 1) Fault mode. 2) CWC and CWoC" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000175_j.mechmachtheory.2015.11.014-Figure8-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000175_j.mechmachtheory.2015.11.014-Figure8-1.png", + "caption": "Fig. 8. Normal force: (a) twisted tooth flank; (b) twist-free tooth flank.", + "texts": [ + " The basic data for work gear, honing cutter and coefficients, and the basic input data and load distribution factors of the meshing gear pair are given in Tables 1 and 2, respectively. The normal force distribution on gear surfaces is shown in Figs. 7\u20138. According to Fig. 7, it confirms that the edge contact is eliminated for both two cases by applying our proposed longitudinal crowning method in external gear honing process. Besides, it shows that maximum normal force on gear surfaces of Case 1 (840.0 N/mm) is higher than that of Case 2 (630.0 N/mm), as shown in Fig. 8. And maximum contact stress distribution on tooth flank in Case 1 (1600.0 N/mm2) is much smaller than that of Case 2 (2000.0 N/mm2), as shown in Figs. 9\u201310. In addition, as shown in Fig. 11, the maximum tooth root stress Case 1 (680.0 N/mm2) is also much higher than that of Case 2 (490.0 N/mm2). Therefore, the load characteristics for the meshing gear pair of Case 2 are much better that that of Case 1. So the load capacity of the meshing gear pair in Case 2 is significantly increased. In this example, we consider effects of the load characteristics to the working conditions of the meshing gear pair" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002342_j.mechmachtheory.2021.104258-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002342_j.mechmachtheory.2021.104258-Figure1-1.png", + "caption": "Fig. 1. Linkage mechanism structural representation: four-bar, six-bar (Stephenson III), and eight-bar (Jansen) mechanisms. The designated end effector for gait pattern generation is denoted by E.", + "texts": [ + " In addition to end effector kinematics, we computed the mechanical advantage of the mechanism to identify the efficiency in force transfer of each design. This work presents a novel basis for objective comparison and insight into 1 DOF gait trainer designs. These insights will help prescribe the most appropriate mechanism for economical and accessible robotic gait training aimed at moderately to severely neurologically impaired individuals. This study focuses on three different mechanisms: four-, six-, and eight-bar configurations offered in previous work [20 , 22 , 25] . The canonical topologies of these designs are illustrated in Fig. 1 . The linkage design for the 1 DOF gait trainer should have the capability of gait pattern reproduction, and it should be easily customizable to individuals. The following section details the dimensional synthesis to human gait data to obtain a best-fit customizable linkage configuration. Dimensional synthesis is an inverse problem that obtains the linkage dimensions required to achieve a prescribed output motion. To establish generality, we used multiple prescribed output motions (gait trajectories) representing the large variability of human gait [30] ", + " The mechanism\u2019s output trajectory is defined as { x E,i \u2208 R 2 : x E,i = f (\u03b8c,i , Z) , \u2200 i \u2208 { 1 , \u00b7 \u00b7 \u00b7 , N} and \u03b8c,i = 2 \u03c0 N (i \u2212 1) \u2208 [0 , 2 \u03c0 ] } , where x E is a position vector of the end effector computed by a well-known analytical kinematic approach, the vector loop method [32] . The analytical kinematics is simplified as the function of the input crank angle, \u03b8c , and a vector of linkage parameters, Z = [ l 1 , l 2 , \u00b7 \u00b7 \u00b7 , l a , \u03b1, \u2205 1 , \u00b7 \u00b7 \u00b7 , \u2205 b ] where l is a link length, \u03b1 is an angle of a ternary link if existing, \u2205 is an angle of a ground link, for the mechanisms shown in Fig. 1 . Here, \u03b8c was equally distributed for one revolution by N samples, representing constant crank velocity. The objective function, J 1 , is the root mean square (RMS) error between T re f and T E , formulated as J 1 ( Z ) = \u221a 1 N N \u2211 i =1 \u2225\u2225x re f,i \u2212 x E,i ( Z ) \u2225\u22252 . (1) The objective function is based on a prescribed human-like trajectory [20 , 25] . However, a geometrical shape in a motion could not be ideally matched by such a structural error function since it limits the search space [33] " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002190_s43452-021-00242-2-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002190_s43452-021-00242-2-Figure1-1.png", + "caption": "Fig. 1 Geometrical configuration of chassis frame", + "texts": [ + " Optimize \u03b7 combination coefficients \u03b2l through game theory to promote the distance between v and vl to reach the minimum deviation. The function can be expressed as This formula corresponds to the optimal first-order derivative linear equation can be given as Find the optimal vector (\u03b21, \u03b22,\u2026, \u03b2\u03b7)T, which is normalized to get the most satisfactory weight vector v\u2019 as According to the space layout of traditional electric vehicle chassis frame, the novel chassis frame is proposed in accordance with the overall dimensions, as shown in Fig.\u00a01a. According to the traditional chassis frame, this study proposes novel multicellular structure including frontend module and middle module structure. This form has the same mass distribution as the traditional frame, the geometrical configurations of the novel chassis frame structure are investigated in this study are shown in Fig.\u00a01b, additionally, the length, the breadth and the height of the chassis frame are 4280\u00a0mm, 1450\u00a0mm and 465\u00a0mm, respectively. The chassis frame is assembled using a modular approach, which is composed of front-end module, middle module and back-end module. The primary function of the frontend modules is to absorb energy in frontal collision; the (10)v = \u2211 l=1 lv T l , (11)min \u2016\u2016\u2016\u2016\u2016 \u2211 i=1 iv T i \u2212 vj \u2016\u2016\u2016\u2016\u20162 ;i = 1, 2,\u22ef , (12) \u23a1\u23a2\u23a2\u23a3 v1v T 1 \u22ef v1v T \u22ee \u22ef \u22ee v v T 1 \u22ef v v T \u23a4 \u23a5\u23a5\u23a6 \u23a1 \u23a2\u23a2\u23a3 1 \u22ee \u23a4 \u23a5\u23a5\u23a6 = \u23a1\u23a2\u23a2\u23a2\u23a3 v1v T 1 \u22ee v v T \u23a4\u23a5\u23a5\u23a5\u23a6 " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000750_j.mechmachtheory.2018.11.002-Figure6-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000750_j.mechmachtheory.2018.11.002-Figure6-1.png", + "caption": "Fig. 6. Circular cutter and virtual shaper represented in the coordinate system of the shaper.", + "texts": [ + " Let us now consider a fixed coordinate system S d coinciding with coordinate system S h in Fig. 4 . The section of the circular cutter defined in coordinate system S h is rotated around axis z d an angle \u03b8 to form the generating surfaces of the circular cutter in fixed coordinate system S d as follows: r (c) d (u, \u03b8 ) = M dh (\u03b8 ) r h (u ) (18) where M dh (\u03b8 ) = \u23a1 \u23a2 \u23a3 cos \u03b8 sin \u03b8 0 0 \u2212 sin \u03b8 cos \u03b8 0 0 0 0 1 0 0 0 0 1 \u23a4 \u23a5 \u23a6 (19) The transformation from coordinate system S h to coordinate system S d is illustrated in Fig. 5 . Fig. 6 shows the relative position of the left side generating circular cutter and the virtual shaper that it mimics. Only the circular cutter for the left side of the shaper is shown. There will be another circular cutter (not represented in Fig. 6 ) to mimic the generating surface of right side of the tooth surfaces of the virtual shaper. The generating surface of the circular cutter has to be represented in the coordinate system S s \u2032 fixed to the virtual shaper ( Fig. 6 ). For that, we will now consider the transformation from coordinate system S d to coordinate system S s \u2032 by considering the matrices M g \u2032 d , M f \u2032 g \u2032 , M e \u2032 f \u2032 , and M s \u2032 e \u2032 as follows r (c) s \u2032 (u, \u03b8 ) = M s \u2032 e \u2032 M e \u2032 f \u2032 M f \u2032 g \u2032 M g \u2032 d r d (u, \u03b8 ) (20) The transformation matrices needed to represent the generating surface of the circular cutter in the coordinate system S s are similar to those used to represent the cross section of the circular cutter and represented using the coordinate systems shown in Fig. 4 . Therefore, M s \u2032 e \u2032 = M \u22121 es , M e \u2032 f \u2032 = M \u22121 f e , M f \u2032 g \u2032 = M \u22121 g f , M g \u2032 d = M \u22121 hg . We recall that the shaper represented in Fig. 4 and the shaper represented in Fig. 6 have different purposes. The former is only for generating the circular cutter surface, however the latter is for describing the generating relationship between the circular cutter and the face gear. Although the cross section of the circular cutter coincides with that of the reference shaper, when the section is rotated about the axis of the circular cutter considering the cutter profile tilt angle \u03b4, the generating surface of the circular cutter might be below the surface of the reference shaper if the cutter profile tilt angle is not sufficient. Fig. 6 shows this situation that leads to the need of incrementing the cutter profile tilt angle. There is a minimum value of the cutter profile tilt angle that provides the generating surface of the circular cutter on or above the reference shaper allowing for negative deviations of the generated face gear and therefore providing good conditions of meshing and contributing to avoid interference with the pinion member of the gear set. In addition, the cross section of the circular cutter has a higher addendum as previously discussed so that it will generate a root surface with some longitudinal curvature" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000769_978-3-662-43893-0-Figure2.3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000769_978-3-662-43893-0-Figure2.3-1.png", + "caption": "Fig. 2.3 Definition of hypoid offset", + "texts": [ + "2 into: \u2013 straight bevel gears \u2013 skew bevel gears \u2013 spiral bevel gears On spiral bevel gears, it is possible to draw a further distinction in terms of the form of the tooth trace, which may be: \u2013 a circular arc, \u2013 an elongated epicycloid, \u2013 an involute or \u2013 an elongated hypocycloid Bevel gears may likewise be classified with respect to their hypoid offset. Bevel gears with no pinion offset have intersecting axes while bevel gears with pinion offset, 12 2 Fundamentals of Bevel Gears known as hypoid gears, have crossed axes. In the latter case, a further distinction may be drawn between gears with positive or negative offset (see Fig. 2.3). Positive offset: \u2013 the pinion axis is displaced in the direction of the spiral angle of the wheel, \u2013 the mean helix angle of the pinion is larger than that of the wheel, \u2013 the diameter of the pinion increases when compared to that of an equivalent gear set with no offset. Negative offset: \u2013 the pinion axis is displaced in a direction opposite that of the spiral angle of the wheel, \u2013 the mean helix angle of the pinion is smaller than that of the wheel, \u2013 the diameter of the pinion decreases when compared to that of an equivalent gear set with no offset", + "1 O v er v ie w o f th e m aj o r cu tt in g m et h o d s fo r sp ir al b ev el g ea rs M a n u fa ct u ri n g m et h o d In d ex in g m et h o d T o o th tr a ce cu rv e T o o th d ep th S lo t w id th a P ro fi le cr o w n in g L en g th w is e cr o w n in g p in io n w h ee l Z y k lo -P al lo id \u00ae / Z y k lo m et \u00ae C o n ti n u o u s E p ic y cl o id C o n st an t V ar ia b le V ar ia b le In th e to o l R ad iu s d if fe re n ce P al lo id \u00ae C o n ti n u o u s In v o lu te C o n st an t C o n st an t C o n st an t _ b In th e to o l N -m et h o d C o n ti n u o u s E p ic y cl o id C o n st an t V ar ia b le V ar ia b le In th e to o l L ea d an g le d if fe re n ce S p ir o fl ex /S p ir ac \u00ae C o n ti n u o u s E p ic y cl o id C o n st an t V ar ia b le V ar ia b le In th e to o l R ad iu s d if fe re n ce T R IA C \u00ae /P E N T A C \u00ae -F H C o n ti n u o u s E p ic y cl o id C o n st an t V ar ia b le V ar ia b le In th e to o l R ad iu s d if fe re n ce K u rv ex S in g le C ir cu la r ar c C o n st an t V ar ia b le V ar ia b le _ c R ad iu s d if fe re n ce A rc o id S in g le C ir cu la r ar c V ar ia b le V ar ia b le V ar ia b le / co n st an t M ac h in e k in em at ic s R ad iu s d if fe re n ce 5 -c u t S in g le C ir cu la r ar c V ar ia b le V ar ia b le C o n st an t M ac h in e k in em at ic s R ad iu s d if fe re n ce C o m p le ti n g S in g le C ir cu la r ar c V ar ia b le C o n st an t C o n st an t M ac h in e k in em at ic s R ad iu s d if fe re n ce W ie n er 2 -S p u r S in g le C ir cu la r ar c C o n st an t V ar ia b le V ar ia b le In th e to o l R ad iu s d if fe re n ce W ie n er 1 -S p u r S in g le C ir cu la r ar c C o n st an t V ar ia b le C o n st an t In th e to o l R ad iu s d if fe re n ce S em iC o m p le ti n g S in g le C ir cu la r ar c C o n st an t V ar ia b le V ar ia b le In th e to o l R ad iu s d if fe re n ce a A t ro o t co n e in n o rm al se ct io n b T o o l w it h ti p re li ef c T o o l w it h p ro tu b er an ce 2.1 Classification of Bevel Gears 19 This chapter deals with the macro geometry of bevel gears, leaving aside modifications to the micro geometry which affect tooth contact (see Sect. 3.3). Because of their conical nature, the macro geometry of bevel gears alters continuously along the face width. Therefore, bevel gears cannot generally be described in such a simplified way as cylindrical gears. The pitch cones of a hypoid gear pair can be obtained from different definitions (see Fig. 2.3). In the past, numerous definitions were made with different perspectives in mind. The main geometry, for example, was described partly in the centre and partly at the heel of the tooth. Since 1997 an expert group in the International Organisation for Standardization has concentrated on the basic geometry of bevel gears, creating the ISO 23509 standard \u201cBevel and Hypoid Gear Geometry\u201d, which aims to unify all the commonly-used methods for the definition of bevel and hypoid gear geometry. Amongst other things, it was agreed that the term \u201cbevel gears\u201d should be used as a generic name, embracing all sub-species like spiral bevel gears, non-offset bevel gears, Zerol\u00ae gears and hypoid gears", + " 30 2 Fundamentals of Bevel Gears profile contact ratio \u03b5\u03b1 \u00bc path of contact base pitch \u00f02:1\u00de overlap ratio \u03b5\u03b2 \u00bc included angle of the face width included angle of one axial pitch \u03c6\u03b2 \u03c4 \u00f02:2\u00de total contact ratio \u03b5\u03b3 \u00bc \u03b5\u03b1 \u00fe \u03b5\u03b2 valid for conjugate tooth flanks \u00f02:3\u00de total contact ratio \u03b5\u03b3 \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffi \u03b52\u03b1 \u00fe \u03b52\u03b2 q for an elliptical contact pattern \u00f02:4\u00de If an offset is included in a bevel gear set, the starting point is usually the wheel and the pinion axis is displaced by the desired offset in such a way that, in a common tangential plane (the crown gear plane), the pitch cones of the two gears contact each other in the middle of their face widths at the mean point. The direction of the offset results from the hand of spiral of the wheel, depending on whether a positive offset (larger pinion diameter than without offset) or a negative offset (smaller pinion diameter than without offset) is desired (cf. Sect. 2.1 and Fig. 2.3). Amongst other things, offset is a design element allowing the gear engineer to reconcile the demands of ground clearance of cars, gearbox space, load capacity and noise behavior. It is useful to define a relative offset arel for the assessment of hypoid offsets, irrespectively of the size of the wheel. arel\u00bc 2a/dm2 arel\u00bc 0 applies to bevel gears without hypoid offset arel\u00bc 1 applies, for example, to crossed axes helical gears The sign is determined by the sign rule for hypoid offsets (see Fig. 2.3). With a positive relative offset, there is a corresponding increase not only in the diameter, 2.2 Gear Geometry 31 the spiral angle and the pitch angle of the pinion, but also in the overlap ratio and the axial force. With a negative relative offset, the parameters behave in exactly the opposite way until, in the extreme case (arel\u00bc 1), the pinion becomes cylindrical. The effects on load capacity, efficiency and noise behavior are described in Chaps. 4 and 5. Hypoid gears represent the general case of bevel gears" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002205_s10544-021-00568-x-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002205_s10544-021-00568-x-Figure3-1.png", + "caption": "Fig. 3 Wiring diagram for the Arduino UNO connection with the 16 \u00d7 2 character green backlight LCD and testing capacitor", + "texts": [ + " 20\u00a0\u00b5l of PBS is injected between the parallel plates at the PVA-CuO coated interface, and readings were taken for a few minutes. 20\u00a0\u00b5l of various prepared concentrations of glucose were injected similarly for further readings. Various interference samples are also compared with the control electrode. Between two different readings, the chip was washed with PBS\u00a0as shown in Fig.\u00a02. The data was logged into the computer using a serial connection from the Arduino UNO chip. Also, the same values were displayed on the 16 \u00d7 2 green back\u00a0light LCD display. The connection for the capacitance measurement has been shown in Fig.\u00a03. The LCD is connected to Arduino UNO through wired connections, and Arduino UNO is connected to the computer display through a serial USB type B cable. Supply voltage pin (VSS) is connected to 5\u00a0V socket; Contrast adjustment pin (VDD), Read/Write (RW), and K are all connected to GND (Ground) socket; V0 is connected to GND through 1 k\u03a9 resistor; Register select (RS) is connected to 12 number socket; Enable (E) is connected to 11 number socket; Data pin (D) 4,5,6 and 7\u00a0are connected to digital pin 5, 4, 3 and 2 number sockets of Arduino UNO, respectively; A is connected to 5\u00a0V socket through 1 k\u03a9 resistor; Analog pin (A0) 0 and 2 of Arduino UNO are connected to the testing chip for capacitance measurement" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000682_s00542-018-3710-z-Figure7-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000682_s00542-018-3710-z-Figure7-1.png", + "caption": "Fig. 7 Schematic illustration of the sensor design Pd-SWCNT/ PAA (a) and Pt-SWCNT/PAA (b) electrodes (Claussen et al. 2011)", + "texts": [ + " The nanoparticles were able to store and release oxygen in its crystalline structure; it could supply O2 to GlutOx to generate H2O2 in the absence of environmental oxygen. Though the fabrication procedure was found to be complex; it can be compromised due to its good sensitivity. Claussen et al. (2011) described two hybrid nanomaterial based biosensor platforms. In this sensor, a network of single-walled carbon nanotubes (SWCNTs) enhanced with Pd nanocubes and Pt nanospheres that are grown in situ through a porous anodic alumina (PAA) template (Fig. 7). These nanocube and nanosphere SWCNT networks are converted into glutamate biosensors by immobilizing the enzyme GlutOx onto the electrode surface. However, Pt nanosphere/SWCNT sensor found to be superior compared to Pd nanocube/SWCNT sensor, with a limit of detect for the Pt nanosphere/SWCNT was 4.6 nM, with a linearity range from 50 nM to 1.6 mM. Moreover, glutamate detection sensitivity for the same found to be five times higher (27.4 mA/mM/cm2) compared to Pd nanocube/ SWCNT based electrode (5" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001156_s00170-018-2786-z-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001156_s00170-018-2786-z-Figure2-1.png", + "caption": "Fig. 2 Experimental preparation. a Experimental setup; b schematic diagram of the experimental setup", + "texts": [ + " The chemical compositions of A2 tool steel and Wallex 40 are given in Table 1. The supplier of the Wallex 40 alloys is Wall Colmonoy. These powders were characterized to analyze particle shape and size distribution. Images of powders were taken using a scanning electron microscope (Hitachi S4700) and are shown in Fig. 1a. It can be seen that most powders were spherical, and some irregular- and angular-shaped particles were also discovered. The particle size distribution is depicted in Fig. 1b, with an average diameter of 71.25 \u03bcm. Figure 2a shows the experimental setup. The LMD system consists of a laser source, a powder feeder, a motion control, and inert gas feeding systems. The continuous-wave (CW) fiber laser with a maximum of 1000 W from IPG Photonics Corporation was used as the heat source. The laser beam diameter was 1.8 mm. A commercial powder feeder (Model 1200, Bay State Surface Technologies, Inc) was utilized to deliver particles into the melting pool. Argon gas was used to carry powder from the powder feeder to a vertical ceramic powder feed nozzle right above the melting pool, with a standoff distance of 10 mm. Relative movement between the substrate and the laser beam was realized by a four-axis table including three rectilinear movements in the x-, y-, and z-axes and one rotation around the x-axis. Figure 2b illustrates the schematic diagram of the experimental setup. Three types of tool path were employed, namely helix (H), circle-line-circle (CLC), and line-arc-line (LAL), as shown in Fig. 3. The axial length of each tool path was 10 mm. For the H route, the round substrate rotates andmoves along the x-axis continuously. The combined motion results in the H tool path (Fig. 3a). For the CLC path, the substrate rotates in a complete circle, then a rapid linear movement was performed and another complete circular motionwas followed" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000951_j.triboint.2016.06.032-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000951_j.triboint.2016.06.032-Figure2-1.png", + "caption": "Fig. 2. Load on the combined planetary of model contain sliding (rolling) friction in section Oiyiz.", + "texts": [ + " The origin O2 is set as the center of the combined planetary plate, and the x axis points from O2 to P, the relative instantaneous center. The y axis is rotated 90\u00b0 counterclockwise to the x axis. The z axis is parallel to the eccentric input shaft. The spatial normal meshing force apply to the combined planetary at point A, B inside and outside of the cycloid groove by the ith steel ball in the ball system are NiA 1 , NiB 1 . The spatial force are projected to surface O2xy, are NixyA 1 and NixyB 1 , respectively. Where \u03b2=N N cosixyA iA 1 1 , \u03b2=N N cosixyB iB 1 1 , and \u03b2 is grooved angle. Fig. 2 shows the load on the combined planetary plate in the shaft section Oiyiz which contains POi, where Oi is the center of the ith steel ball. When the input shaft rotates clockwise, the ball prevents the combined planetary plate from the motion. In the first half cycle, half of steel balls are main transmission force points at A, in the second half cycle, half of steel balls are secondary transmission force points at B. The tangential sliding friction force applied to the combined planetary plate by the ball at the points A, B, is \u03bc \u03bc= = ( )F N F N, 1ixyA ixyA ixyB ixyB 1 1 1 1 1 1 where \u03bc1 is the sliding friction coefficient between steel ball and cycloid groove" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000820_s11044-014-9445-4-Figure11-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000820_s11044-014-9445-4-Figure11-1.png", + "caption": "Fig. 11 Model of underactuated rimless wheel with torso", + "texts": [ + " This is because the initial state error is very small and Eq. (68) converges to an indeterminate form after taking a few steps. This problem on numerical calculation arises in a fast convergent gait and is discussed again in the next section. This section discusses the discrete behavior of HZD of an underactuated rimless wheel (URW) with a torso in the same manner as the previous sections. The main purpose is to specify the analysis method with general formulae. 5.1 Model of underactuated rimless wheel and its linear approximate equation of motion Figure 11 shows the model of an URW with a torso [16]. This URW consists of an eightlegged symmetrically-shaped RW of Fig. 1 and a torso link. The torso link is connected to the RW at the central position and the moment of inertia about the joint is I [kg \u00b7 m2]. Let \u03b8 = [\u03b81 \u03b82]T be the generalized coordinate vector. The equation of motion then becomes [ Ml2 0 0 I ][ \u03b8\u03081 \u03b8\u03082 ] + [\u2212Mgl sin \u03b81 0 ] = [ 1 \u22121 ] u. (72) By linearizing Eq. (72) about \u03b8 = \u03b8\u0307 = 02\u00d71, we get[ Ml2 0 0 I ][ \u03b8\u03081 \u03b8\u03082 ] + [\u2212Mgl 0 0 0 ][ \u03b81 \u03b82 ] = [ 1 \u22121 ] u" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000953_1350650115586032-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000953_1350650115586032-Figure1-1.png", + "caption": "Figure 1. Principle of operation.", + "texts": [ + " In this case, the heat stress may arouse new failure mode which is the consequence of temperature increase, and that may not be acceptable in accelerated life test. In the present study, life test and mixed lubrication analysis of HD with new lubricating materials for space applications are conducted to simulate the degeneration of transmission performance and the failure mechanism. Based on the obtained results, a new method of accelerated life test is proposed and verified by experiments. The principle of HD is unique, capable of transmitting high torque through an elastically deformable component. It consists of three concentric elements (Figure 1)17: 1. The circular spline(CS) is a rigid ring with internal teeth, engaging the teeth of the flexspline across the major axis of the wave generator. 2. The flexspline(FS) is a non-rigid, thin cylindrical steel cup with external teeth on a slightly smaller pitch diameter than the circular spline. 3. The wave generator (WG) is a thin-raced ball bearing fitted onto an elliptical plug serving as a high efficiency torque converter. In this research, type HD60 with a reduction ratio of 100 is provided for the test, the material characteristics and lubrication of the HD are listed in Tables 1 and 2, and the performance parameters of WC-DLC:H is listed in Table 3" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000080_j.engfailanal.2013.03.008-Figure17-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000080_j.engfailanal.2013.03.008-Figure17-1.png", + "caption": "Fig. 17. Shape of cracked gear (5th order, a = 0.4, reference circle crack).", + "texts": [], + "surrounding_texts": [ + "Assuming a crack length of a = 0.8 mm at the pitch circle, when the other parameters are fixed, and only the magnitude of the load is changed, the relationship between load and SIF can be obtained by analysis as shown in Figs. 23 and 24. From these figures, it can be seen that the relationship between load and SIF is linear, that is to say: KI = A1w, KII = A2w, where A1 and A2 are coefficients that are relevant to the main parameters of the gear, position of load, crack length, and so on. From the figures, it can also be seen that the rate of changed for KI, KII is fastest when the load acts at point E." + ] + }, + { + "image_filename": "designv10_14_0000032_scis-isis.2014.7044693-Figure5-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000032_scis-isis.2014.7044693-Figure5-1.png", + "caption": "Fig. 5. Verification result of the holding force in the case of changing the filling factor", + "texts": [ + " This result suggests that the grip force changes with the press force. Thus, we conducted an experiment to confirm the relationship between the press force and the maximum grip force. The filling material and quantity for the UJG were chosen according the conditions in which the best result was achieved in the above-mentioned experiment, i.e., the filling material was ground coffee and the filling ratio was 50%. We varied the press force and observed the grip force. The experimental result is shown in Fig. 5. First, it is found that the press force of the UJG on a target object affects the grip force. Moreover, it is found that there is an appropriate force for maximizing the grip force of the UJG. That is, if the press force is too small or too large, the UJG cannot generete the maximum grip force. Furthermore, it is found that there is a tendency for the distribution of the fluctuation in grip force to reduce near the press force that generates the maximum grip force. V. AUTOMATIC HOLD MECHANISM Based on the results from the previous experiment, we developed a force feedback mechanism for the detection of press force" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000894_tro.2015.2489499-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000894_tro.2015.2489499-Figure1-1.png", + "caption": "Fig. 1. Projection of a cylinder in the image.", + "texts": [ + " Notice that, using this notation, the well-known (normalized) Plu\u0308cker coordinates [24], [25] are the couple (cu, cncn). The projection of such a line in the image plane, expressed in the camera frame, has the characteristic equation [23] cnT cp = 0 (2) where cp are the coordinates in the camera frame of a point P in the image plane, lying on the line. The legs of parallel robots usually have cylindrical cross sections [25]. The edges of the ith cylindrical leg are given, in the camera frame, by [14] (see Fig. 1) cn1 i = \u2212 cos \u03b8i chi \u2212 sin \u03b8i cui \u00d7 chi (3) cn2 i = + cos \u03b8i chi \u2212 sin \u03b8i cui \u00d7 chi (4) where cos \u03b8i = \u221a ch2 i \u2212 R2 i / chi , sin \u03b8i = Ri/ chi and (cui , chi , chi) are the Binormalized Plu\u0308cker coordinates of the cylinder axis, and Ri is the cylinder radius. It was also shown in [14] that the leg orientation, expressed in the camera frame, is given by cui = cn1 i \u00d7 cn2 i \u2016cn1 i \u00d7 cn2 i \u2016 . (5) 2In the following of the paper, the superscript before the vector denotes the frame in which the vector is expressed (\u201cb\u201d for the base frame, \u201cc\u201d for the camera frame, and \u201cp\u201d for the pixel frame). If there is no superscript, the vector can be written in any frame. Let us remark that each cylinder edge is a line in space, with Binormalized Plu\u0308cker expressed in the camera frame (cui , cnj i , cnj i ) (see Fig. 1). The proposed control approach was to servo the leg directions cui [12]. Some brief recalls on this type of controller are done below. 1) Interaction Matrix: Visual servoing is based on the socalled interaction matrix LT [26], which relates the instantaneous relative motion Tc = c\u03c4 c \u2212 c\u03c4 s between the camera and the scene, to the time derivative of the vector s of all the visual primitives that are used through s\u0307 = LT (s)Tc (6) where c\u03c4 c and c\u03c4 s are, respectively, the kinematic screw of the camera and the scene, both expressed in Rc , i", + " By extension of these results, it could be straightforwardly proven that all robots with three translational dof of the platform, or with Scho\u0308nflies motions (three translational dof of the platform plus one rotational dof about one fixed axis), which are composed of identical legs made of PR\u03a0R architecture, or also with PUU architecture and for at least two P joints are not parallel (e.g., the Y-STAR [45]) are fully controllable in their whole workspace using the leg direction observation. After this classification, one additional question is to know if, by adding additional information in the controller, the robots that were uncontrollable or partially controllable in their whole workspace can become fully controllable. For example, it was very recently proven in [46] that, from the projection of the cylindrical leg in the image plane (see Fig. 1), it is not only possible to estimate the leg direction, but also the Plu\u0308cker coordinates of the line passing through the axis of the cylinder, i.e., the direction and location in space of this line. Using this information leads to a modification of the virtual leg as shown in Fig. 16(a): the additional prismatic chain, instead of being passive, becomes active. This additional information can solve many issues of controllability mentioned above. For example, by estimating the Plu\u0308cker coordinates of the line passing through its legs, the PRRRP robot of Section IV-A becomes controllable as the hidden robot model becomes a PRRRP robot [see Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001833_tec.2020.3041658-Figure9-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001833_tec.2020.3041658-Figure9-1.png", + "caption": "Fig. 9. Open-circuit flux density distributions of the SPM and CPM machines. (a) SPM. (b) CPM.", + "texts": [ + " 7(a) and 8(a), which is mainly due to the relatively high magnetic Authorized licensed use limited to: Dedan Kimathi University of Technology. Downloaded on May 18,2021 at 14:38:15 UTC from IEEE Xplore. Restrictions apply. 0885-8969 (c) 2020 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. saturation occurring in stator tooth tips of both two machines, as illustrated in Fig. 9. However, the magnetic saturations of the two machines are not very severe for the majority of the iron parts. As a result, a satisfactory agreement between the analytical and FE predictions can be achieved, which confirms the effectiveness of the analytical model. Besides, it can be seen that the analytically predicted harmonic amplitudes are slightly higher than FE prediction in SPM and CPM machines, as shown in Figs. 7(b) and 8(b). This is mainly due to the fact that the localized magnetic saturation is neglected in the analytical model" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000897_j.cad.2015.12.001-Figure4-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000897_j.cad.2015.12.001-Figure4-1.png", + "caption": "Fig. 4. Five cutting surfaces of the outside blade.", + "texts": [ + " In order to do so, a series of matrix transformations are formulated as follows: M1h = cos (\u2212\u03b8) \u2212 sin (\u2212\u03b8) 0 0 sin (\u2212\u03b8) cos (\u2212\u03b8) 0 0 0 0 1 0 0 0 0 1 , (7) Mm1 = 1 0 0 H 0 1 0 V 0 0 1 \u2212BO 0 0 0 1 , (8) M2m = cos (\u2212 (\u03c0/2 \u2212 \u03b3m)) 0 sin (\u2212 (\u03c0/2 \u2212 \u03b3m)) 0 0 1 0 0 \u2212 sin (\u2212 (\u03c0/2 \u2212 \u03b3m)) 0 cos (\u2212 (\u03c0/2 \u2212 \u03b3m)) \u22121Xp 0 0 0 1 , (9) Mw2 = cos (\u2212Rb\u03b8) \u2212 sin (\u2212Rb\u03b8) 0 0 sin (\u2212Rb\u03b8) cos (\u2212Rb\u03b8) 0 0 0 0 1 0 0 0 0 1 , (10) where M1h, Mm1, M2m and Mw2 are transformation matrixes from Sh to S1, S1 to Sm, Sm to S2 and from S2 to Sw, respectively, and the equation of BO can be written in terms of \u03b8 as BO (\u03b8) = BO0 \u2212 a21 \u00b7 Ax \u03c92 H \u03b82 \u2212 a2 \u00b7 V0 \u03c9H \u03b8, (11) where\u03c9H ,Ax,V0 and BO0 are cutter head angular velocity (rpm), acceleration (mm/s2), initial velocity (mm/s) and initial offset to back (mm), respectively, and a1 = \u221a 450 \u03c0 and a2 = 30 \u03c0 . BO decreases during machining process with a specified acceleration till it reaches zero. Eventually, the cutting edge formulation in coordinate system Sw can be written as rw (s, \u03b8) = Mwhrh (s) , (12) where Mwh = Mw2M2mMm1M1h. (13) Eq. (12) is a formulation of a unique surface called the cutting surface in terms of two independent variables, s and \u03b8 . Fig. 4 shows cutting surfaces of a single outside blade for five rotations of the cutting system. As it can be imagined, taking into account the cutting surfaces of all the blades considering the process kinematic and their interactionwith theworkpiecemakes a very complicated scenario to be simulated. Detailed descriptions in this matter can be found in [8] in which face-hobbing is simulated numerically. In this paper, the geometry of the workpiece in face-milling is represented by the bevel gear geometric parameters (pitch angle, \u0393 , outer cone distance, C , face width, F , addendum, a, and dedendum, b)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001671_tasc.2019.2962687-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001671_tasc.2019.2962687-Figure1-1.png", + "caption": "Fig. 1. CMG topology. (a) Conventional model. (b) Proposed model.", + "texts": [ + " Therefore, in order to get higher torque and low torque ripple, it is necessary to research the CMG with simple structure and steady transmission. In this paper, a novel asymmetric pole CMG with unequal Halbach arrays on the inner rotor and spoke PMs on the outer rotor is presented to improve the output torque. Section II will introduce the configuration and working principles of the proposed CMG. In Section III, the performances of the proposed CMG will be computed by the finite element analysis (FEA). Finally, several conclusions are drawn in Section IV. As shown in Fig. 1(a), the conventional CMG is composed of two rotors with surfaced mounted PMs and a stationary ring. The stationary ring is uniformly distributed by magnetic conducting silicon steel and epoxy resin. Its function is to modulate the air gap magnetic field on both sides. The proposed CMG is shown in Fig. 1(b), PMs on the outer rotor are arranged in Spoke-type, on the inner rotor, PMs are arranged in Halbach arrays. However, the N-pole and S-poles are asymmetric, and each pole is divided into three small pieces of PMs. The width of each small PM is not equal. The structure of asymmetric poles and unequal small PMs is shown in Fig. 2. As shown in Fig. 2, the PMs on the inner rotor of the proposed model are asymmrteic, the N-pole and S-pole are not 1:1 uniform distribution. The PM of each pole is divided into three small pieces" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002015_j.asr.2020.09.040-Figure4-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002015_j.asr.2020.09.040-Figure4-1.png", + "caption": "Fig. 4. Infinitesimal region in polar coordinates. Relationship between R; dh; dR and radius R\u00fe dR between the inner and outer polygon.", + "texts": [ + " 3, which describes the key elements involved in the rendering of the crease geometry and the overall folding structure. By observing Fig. 3, given a regular polygon with center at origin O, the crease Fold A is tangential to the edge of the core polygon. Considering a regular polygon with N sides, the spiral folding pattern is expected to have N creases of the type Fold A. Then, let P be an arbitrary point in Fold A, it is possible to describe an infinitesimal region bounded by an inner polygon with apothem R and an outer polygon with apothem R\u00fe dR, in polar coordinates, as shown by Fig. 4, which will aid in outlining the governing equations to render the geometry of Fold A. Here, by looking at the relation among the infinitesimal distances dR and Rdh, and the angle / in Fig. 4, the following holds tan /\u00f0 \u00de \u00bc Rdh dR : \u00f01\u00de The above expression is useful to find the angle h by numerical integration of dh dR for a defined interval in R 2 R0;Rf . However, since the angle / is expected to vary with R as well, it becomes necessary to find further relationships between / and R. In the following we find such relationships. Let Fold A0 be the adjacent crease to Fold A (as shown by Fig. 3), it is possible to define an infinitesimal region between the inner polygon and the outer polygon, as shown by Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001308_icems.2019.8922504-Figure8-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001308_icems.2019.8922504-Figure8-1.png", + "caption": "Fig. 8. Highest temperature of motor stator(a)laminate (b)SMC", + "texts": [ + " 7 shows that the copper loss of SMC motor is slightly higher than that of laminated motor, but the iron loss of SMC motor is lower than that of laminated motor, and the total loss of SMC motor is lower than that of laminated motor. The temperature characteristics of each part of the motor are shown in Table \u2164. The stator iron loss of the motor is the major loss, so the temperature rise of the stator is the highest. The stator highest temperature of SMC motor and laminated motor under rated operating conditions as shown in Fig. 8. TABLE \u2164 THERMAI PARAMETERS OF COMPONENTS MATERIAL Motor part Material \u03c1 [Kg/m3 ] c[J/(kg\u00b7\u2103)] Conductors Copper 8933 386 Core Lamination 7650 450 Core SMC 7400 450 PM NdFeB 7400 440 III. VOLUME OPTIMIZATION OF SMC MOTOR The volume of SMC motor can be reduced and the power density of SMC motor can be increased, because of the low temperature of SMC motor. Fig. 8 shows the temperature of the motor decreases by 8.5% when SMC material is used as the stator of the motor. Reducing the length of the motor without changing other parameters of the motor, not only can increase the power density of the motor, but will affect the torque of motor. For BLDC motor, reducing length of motor the torque will increase, and the current will also increase. It is necessary to increase winding turns to reduce motor torque and current. The relationship between iron loss and motor length of SMC motor is shown in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000372_s11249-015-0565-7-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000372_s11249-015-0565-7-Figure2-1.png", + "caption": "Fig. 2 Illustration of the measured spots on the apparatus", + "texts": [ + " Fluorescence emission: a photon is emitted returning the fluorophore to its ground state. Because of the energy dissipation during excited-state lifetime, this photon has a lower energy and therefore the longer wavelength than the excitation photon. The difference in wavelengths, also known as Stokes shift, is absolutely fundamental since it allows a separation of emission and excitation. The evaluation of lubricant rupture in the contact outlet was based on qualitative comparison of the intensities of pixels in contact tracks on both rubbing surfaces as illustrated in Fig. 2. It was proved by Azushima [20] that in case of thin films there is a linear dependence between the intensity of emission and thickness of the lubricant layer, so that the intensity can be considered as the non-calibrated film thickness. Measured thickness is labelled as d1 and d2 in the results. Thickness was not calibrated to the specific one because only a ratio between two values is considered in this study. Examples of the fluorescence technique output are presented in Fig. 3. Darker areas can be interpreted as areas with thinner lubricant film" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001219_s11041-019-00351-z-Figure4-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001219_s11041-019-00351-z-Figure4-1.png", + "caption": "Fig. 4. Scheme of cutting (a) and drawing (b ) of specimens for mechanical tests.", + "texts": [ + " The specimens for studying the properties in directions OX and OZ were fabricated from billets of the second type in the form of one-bead-thick narrow walls with a size 60 15 100 mm. Both types of billets were grown in direction OX. Table 2 presents the used modes of deposition. The grown billets were used to fabricate cylindrical specimens for static tensile tests and laps for studying the macrostructure and determining the Vickers microhardness in accordance to the GOST 6996\u201366 and GOST R ISO 6507-1\u20132007 Standards. Figure 4 presents a scheme of fabrication and a drawing of specimens tested for static tension until failure in a Tinius Olsen H25KT facility (maximum load 2500 kg). The macrostructure of the billets of both types was studied on laps prepared using a Buehler system. The structure was analyzed using a Leica-DMI5000 optical microscope equipped with a Thixomet system for image analysis. The porosity of the deposited metal was determined from the ratio of the area of the pores to the total area of the lap by processing structure images with the help of the CorelDraw and MatLab (Matlab Region Analyser and Matlab Color Thresholder) software" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001112_j.vacuum.2018.04.058-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001112_j.vacuum.2018.04.058-Figure2-1.png", + "caption": "Fig. 2. Laser scanning strategy to obtain open/closed cell pores.", + "texts": [ + " This study aimed to examine the effect of this parameter on the foaming and deposition characteristics by focusing on the height of the deposited material as well as the number and size of the pores generated under each set of conditions. In addition, the effects of the foaming agent ratio (FR), power (P), and powder feed rate (PF) were also evaluated. These parameters affect the energy density and powder density, which determine the powder melting process during deposition. In a preliminary experiment, the following conditions were determined to be suitable for forming pores within the tracks: FR of 30%, P of 900W, and PF of 2.5 g/min. Table 2 lists the process conditions adopted for the study. As shown in Fig. 2, a cube of 20mm\u00d715mm\u00d71.5mm was deposited in a zigzag manner to analyze the effect of the track spacing. The width of each track was 1.0 mm and, as shown in Fig. 3, the deposited specimen was fabricated to have different TS values. The heights of the specimens fabricated under the different conditions were measured. Because the deposition height varied with the degree of foaming during the deposition process, it was indicative of the extent of foaming under the different process conditions. In addition, in order to observe the cross-sections of the pores formed, the specimens were polished using sandpapers with different grits (P200, P400, P800, P1200, P1500, P2000, and P4000) and a 1 \u03bcm diamond suspension" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001655_tia.2019.2956717-Figure5-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001655_tia.2019.2956717-Figure5-1.png", + "caption": "Fig. 5: Winding 1: the coils are made with non-constant wire width and constant height.", + "texts": [ + " Variables \u03b2c and \u03b3c are bounded to a maximal value of 58\u25e6 for mechanical assembly reasons. A constraint on the width of the wire of segment III has been set so that it stays above a minimal value for mechanical strength reasons. The first results presented are those of a winding with a single height of the wire at the locations of end-windings (see Fig. 4a). This winding is referred to as Winding 1. The results after optimization are zbend = 6.2 mm and \u03b2c = \u03b3c = 58\u25e6. Angles \u03b2c and \u03b3c have been pushed to their maximal values. The resulting winding can be seen in Fig. 5. Table II gives the values of the resistance and the torque constant for one coil and the global motor constant. The resistance for one coil is lower by almost 25% compared to Reference Winding. Given that the number of turns has been set to have the torque constant for one coil as close as the one of Reference Winding, the motor constant is consequently increased by 15.5%. Thereafter, the height of the wire can be doubled at the locations of end-windings (see Fig. 4b). This gives Winding 2 with zbend = 6" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000554_s12541-016-0125-6-Figure6-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000554_s12541-016-0125-6-Figure6-1.png", + "caption": "Fig. 6 Schematic view of gear tooth", + "texts": [ + " Testing for the entire range of torque for unreinforced and reinforced gears under bi-directional and unidirectional loading would be more time consuming. It is well known fact that life experienced by unreinforced gears will be less as compared to reinforced gears. Hence, in the present study, unreinforced and reinforced gears were tested at lower and higher torque levels respectively. Torque exerted on gear mesh and angular displacement of gear mesh measured at driver gear end, were continuously measured and acquired at a rate of 300 Hz. Schematic view of the gear tooth is shown in Fig. 6. Induced tooth root bending stress (\u03c3b) for different torque was computed using the Lewis Eq. (1) (1) where, Ft, m and b are tangential load acting on the gear tooth, module and face width respectively. Important gear parameters used for \u03c3b Ft / mbY( )= *specimens were not tested upto rupture (extension of 40 mm observed at 1.8 kN) computation is provided in Table 1. The Lewis form factor (Y) of 0.736 was computed using Eq. (2) (2) Where, Sq and hq are critical section thickness and height at which the applied force acts from the critical section respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001034_j.sab.2017.06.012-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001034_j.sab.2017.06.012-Figure1-1.png", + "caption": "Fig. 1. Optical scheme of the set-up for the recording of molecular absorption spectra. 1 \u2013 Xe arc continuum source; 2, 5, 8 - elliptical mirrors; 3 - atomizer; 4 - beam splitter; 6, 9 - plane mirrors; 7 - DEMON entrance slit; 10 - MOSES entrance slit.", + "texts": [ + " CS measurements will not be included here, since they have been accomplished previously by us [7,12]. Two continuum source spectrometer systems were applied for the absorption measurements. The first one was used to acquire overview absorption spectra of SiS, GeS, SnS, and PbS. It is a combination of MOSES and a commercial contrAA 700 spectrometer (Analytik Jena AG, Jena, Germany). Both units are coupled by a polka-dot mirror, which is placed in the beam path of the continuum source of the contrAA700 and transfers 50% of the light to theMOSES system. The details of the set-up are shown in Fig. 1. Owing to the linkage, the absorption signals could be recorded synchronously by two optical arrangements and detectors. (See Fig. 2a,b) \u03bcL 500mg L\u22121 S (ammonium sulfate) standard solution in the graphite furnace, recorded Detailed information about the in-house development MOSES can be found in a recent publication by the authors [20]. The spectrograph originally featured a spectral coverage from 190 nm to 390 nm, which was subdivided into four slightly overlapping segments. Each of them could be addressed by proper prism rotation and could then be recorded simultaneously" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001478_j.acme.2019.06.005-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001478_j.acme.2019.06.005-Figure3-1.png", + "caption": "Fig. 3 \u2013 Preparation and discretization of the geometric model.", + "texts": [ + " Then, the geometric model underwent simplifications, consisting of omitting elements that were not significant from the point of view of the analyzed phenomenon, such as small roundings, phases, mounting holes and inclinations. In the next stage, the model was discretized with the use of solid elements (CHEXA type, six-sided isoperimetric elements with 8 nodes), material data was defined for individual elements (Young's modulus, Poisson's number) and the surface-to-surface contact between the balls and raceways was modeled corresponding to Hertzian contact model. The model building process is presented in Fig. 3. An extremely important issue affecting the accuracy of computations in a contact model is selection of the appropriate mesh density [24,25]. Bearing in mind both the aforementioned accuracy of computations and the time needed for their implementation, it was decided to perform local mesh refinement in the vicinity of the contact areas. Preliminary tests for various finite element mesh densities were performed, and a density was chosen that made it possible to obtain best agreement of the actual stress distribution with Hertzian theory (the maximum stress in the contact area in the constructed model was 3620 MPa, compared to the theoretical value of 3660 MPa)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002401_j.mechmachtheory.2021.104371-Figure9-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002401_j.mechmachtheory.2021.104371-Figure9-1.png", + "caption": "Fig. 9. Three-dimensional dynamic model of distributed meshing of a helical gear pair.", + "texts": [], + "surrounding_texts": [ + "This dynamic model of a helical gear pair is established by the Lagrange method, as shown in Eqs. (33) - (34) : The dynamic equation of gear 1 is: m 1 \u0308x 1 + c 1 x \u0307 x1 + k 1 x x 1 = \u2212F m sin ( \u03b1 + \u03b3 ) cos (\u03b2) + F f cos ( \u03b1 + \u03b3 ) m 1 \u0308y 1 + c 1 y \u0307 y1 + k 1 y y 1 = F m cos ( \u03b1 + \u03b3 ) cos (\u03b2) + F f sin ( \u03b1 + \u03b3 ) m 1 \u0308z 1 + c 1 z \u0307 z1 + k 1 z z 1 = F m sin (\u03b2) I x 1 \u0308\u03b8x 1 + I z1 \u02d9 \u03b82 x 1 = \u2212 T m cos (\u03b1 + \u03b3 ) I y 1 \u0308\u03b8y 1 \u2212 I z1 \u02d9 \u03b82 y 1 = \u2212T m sin (\u03b1 + \u03b3 ) I z1 \u0308\u03b81 = T in \u2212 F m cos (\u03b2) r b1 \u2212 F f r a 1 (33) The dynamic equation of gear 2 is: m 2 \u0308x 2 + c 2 x \u0307 x2 + k 2 x x 2 = F m sin ( \u03b1 + \u03b3 ) cos (\u03b2) \u2212 F f cos ( \u03b1 + \u03b3 ) m 2 \u0308y 2 + c 2 y \u0307 y2 + k 2 y y 2 = \u2212 F m cos ( \u03b1 + \u03b3 ) cos (\u03b2) \u2212 F f sin ( \u03b1 + \u03b3 ) m 2 \u0308z 2 + c 2 z \u0307 z2 + k 2 z z 2 = \u2212 F m sin (\u03b2) I x 2 \u0308\u03b8x 2 + I z2 \u02d9 \u03b82 x 2 = T m cos (\u03b1 + \u03b3 ) I y 2 \u0308\u03b8y 2 \u2212 I z2 \u02d9 \u03b82 y 2 = T m sin (\u03b1 + \u03b3 ) I z2 \u0308\u03b82 = T out \u2212 F m cos (\u03b2) r b2 \u2212 F f r a 2 (34) where, T in and T out are the input and output torques of the system, respectively. k nx , k ny , and k nz and c nx , c nx , and c nx are the stiffness and damping of each gear centre bearing. I xn , I yn and I zn ( n = 1 , 2 ) are the moments of inertia of the gears. F\u0304 f is the tooth surface friction, as shown in Eq. (34) : F\u0304 f = \u03bcF\u0304 m (35)" + ] + }, + { + "image_filename": "designv10_14_0002403_j.isatra.2021.04.031-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002403_j.isatra.2021.04.031-Figure1-1.png", + "caption": "Fig. 1. The coordinate system of RLV.", + "texts": [ + " In Section 3, the proposed FTDO-NOC scheme, which is based on the novel AMGST algorithm and ADP method, is developed. Section 4 presents the comparative simulation to verify the effectiveness and superiority of the proposed control scheme. Finally, in Section 5, some conclusions are drawn. 2. Problem formulation and preliminaries 2.1. Problem formulation The 6-DOF dynamic equations of the rigid-body RLV consist of 3-DOF translational equations and 3-DOF rotational equations [27]. 3-DOF rotational equations, which describe RLV\u2019s motion around the center of mass (shown in Fig. 1), are used to design the reentry attitude control scheme and given in (1)\u2013(6). \u03b1\u0307 = \u2212p cos\u03b1 tan\u03b2 \u2212 r sin\u03b1 tan\u03b2 + sin \u03c3 cos\u03b2 [ \u03c7\u0307 cos \u03b3 \u2212 \u03c6\u0307 sin\u03c7 sin \u03b3 + q + (\u03b8\u0307 +\u2126E) \u00d7 (cos\u03c6 cos\u03c7 sin \u03b3 \u2212 sin\u03c6 cos \u03b3 ) ] \u2212 cos \u03c3 cos\u03b2 \u00d7 [ \u03b3\u0307 \u2212 \u03c6\u0307 cos\u03c7 \u2212 ( \u03b8\u0307 +\u2126 ) cos\u03c6 sin\u03c7 ] , (1) E C. Zhang, G. Zhang and Q. Dong ISA Transactions xxx (xxxx) xxx \u03b2 \u03c3 p p I t r r b ( f A m t m c i w o t r t k \u2206 a g I R R i n s c t t 2 d s N E d | m \u2308 o N a o c \u23a1\u23a2\u23a3 \u23a1\u23a2\u23a3 D x t P m p m \u02d9 = p sin\u03b1 \u2212 r cos\u03b1 + sin \u03c3 [ \u03b3\u0307 \u2212 \u03c6\u0307 cos\u03c7 + (\u03b8\u0307 +\u2126E) \u00d7 cos\u03c6 sin\u03c7 ] + cos \u03c3 [ \u03c7\u0307 cos \u03b3 \u2212 \u03c6\u0307 sin\u03c7 cos \u03b3 \u2212 ( \u03b8\u0307 +\u2126E ) (cos\u03c6 cos\u03c7 sin \u03b3 \u2212 sin\u03c6 cos \u03b3 ) ] , (2) \u02d9 = \u2212p cos\u03b1 cos\u03b2 \u2212 q sin\u03b2 \u2212 r sin\u03b1 cos\u03b2 + \u03b1\u0307 sin\u03b2 \u2212 \u03c7\u0307 sin \u03b3 \u2212 \u03c6\u0307 sin\u03c7 cos \u03b3 + ( \u03b8\u0307 +\u2126E ) \u00d7 (cos\u03c6 cos\u03c7 sin \u03b3 + sin\u03c6 cos \u03b3 ) , (3) \u02d9 = IzzMx IxxIzz \u2212 I2xz + IxzMz IxxIzz \u2212 I2xz + ( Ixx \u2212 Iyy + Izz ) IxxIzz \u2212 I2xz \u00d7 Ixzpq + ( Iyy \u2212 Izz ) Izz \u2212 I2xz IxxIzz \u2212 I2xz qr, (4) q\u0307 = My Iyy + Ixz Iyy ( r2 \u2212 p2 ) + (Izz \u2212 Ixx) Ixz Iyy pr, (5) r\u0307 = IxzMx IxxIzz \u2212 I2xz + IxxMz IxxIzz \u2212 I2xz + ( Ixx \u2212 Iyy ) Ixx IxxIzz \u2212 I2xz \u00d7 pq + ( Iyy \u2212 Ixx \u2212 Izz ) Ixz IxxIzz \u2212 I2xz qr, (6) where \u03b1, \u03b2 , \u03c3 represent angle of attack (AOA), sideslip angle and bank angle, respectively; p, q, r represent roll rate, pitch rate and yaw rate, respectively; Mx, My, and Mz denote rolling moment, itching moment, and yawing moment, which are control inputs; ij(i = x, y, z, j = x, y, z) are inertia of the vehicle" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001924_j.jmps.2020.103959-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001924_j.jmps.2020.103959-Figure1-1.png", + "caption": "Fig. 1. (a) Three-dimensional view with an intersecting symmetry plane of a shell with a large geometric imperfection. (b) Shell cross-section on the intersecting plane, with shell geometry described by radius R , radius to thickness ratio R / t , imperfection angular width \u03b8w , imperfection amplitude h / l , and meridional angle \u03b8m . Effect of varying defect size and location: (c) increase in defect width through change of \u03b8w , (d) defect center location through variation of meridional angle, \u03b8m , (e) defect amplitude for increasing h / l . (f) Special cases: h/l = 0 yields a linear profile generating a conical shell, h/l = ( \u221a 2 \u22121 ) / 2 gives a concave profile, i.e. an arc generatrix for a surface of revolution that departs from the spherical and conical geometry.", + "texts": [ + "3 ) are presented to show a buckling response over a large volume change (above 20% the initial volume), a behaviour previously unobserved in the literature of hemispherical thin shells. Finally, a sensitivity study ( Section 4 ) unveils a direct relation between shell response, defect characteristics and shell geometry, which altogether can be tuned to render three distinct snap-through modes besides bifurcation buckling, which can be exploited for the design of soft metamaterials for soft robotics, mechanism-based structures and smart actuators. We consider a hemispherical thin shell ( Fig. 1 ) with a large geometric imperfection in the form of an axisymmetric circular-arc indentation that can vary in amplitude, angular width and location. The cross-section of the shell is defined by the radius R and thickness t , defining its slenderness, R / t . The large imperfection traces a circular arc with center O 2 and extent defined by h / l , i.e. the amplitude h to width l ratio, and the angular width \u03b8w . We examine the cases where the imperfection can vary in angular width \u03b8w ( Fig. 1 c), in position through the meridional angle \u03b8m ( Fig. 1 d), defining the position of its center O 2 , and in amplitude h / l from 0 and 0.5 ( Fig. 1 e), the former describing the case of an arc collapsed to a line segment, and the latter being a defect in the form of a semicircle. The imperfection is also assumed to lie between the equator and the upper pole of the semi circumference ( Fig. 1 e), hence satisfying the constraint on the meridional angle \u03b8m and the angular width \u03b8w : \u03b8w 2 < \u03b8m < \u03c0 2 \u2212 \u03b8w 2 (1) Fig. 1 f shows two special cases for extremely large imperfections with width \u03b8w = \u03c0/ 2 and meridional angle \u03b8m = \u03c0/ 4 . The hemispherical shell degenerates into either a cone for amplitude h/l = 0 , or for h/l = ( \u221a 2 \u22121 ) / 2 into a surface of revolution obtained by rotating the generatrix, a concave arc, around the vertical axis. Fig. 2 shows the basic steps of the manufacturing process adapted from the literature ( Lee et al., 2016a , b ) to build shell samples with thin hemispherical smooth geometry", + " For the metrics here used, we adopt the normalized maximum pressure p max / p C , the snap-ratio | p 1 \u2212 p 2 | / p 1 , and the ratio between snapthrough pressure and the maximum pressure p 1 / p max , along with their counterparts in volume change V max / V 0 , | V 1 \u2212 V 2 | / V 1 , and V 1 / V max . p C is Zoelly\u2019s buckling pressure. V 0 is the volume of a perfect hemisphere. The definition of these snap ratios enables the assessment of the relative difference in pressure and volume change between the limit points of the snapthrough buckling branch ( Vieira et al., 2017 ). Although the large imperfection here presented applies also to degenerate imperfect shells reducing to a cone ( Fig. 1 f), in this section we restrict our sensitivity study to hemispherical geometries with imperfection parameters falling within given geometric ranges. Specifically, the angular width \u03b8w is assumed to range from 2.4 \u00b0 to 24 \u00b0, the amplitude h / l from 0.05 to 0.5, the meridional angle \u03b8m from 30 \u00b0 to 70 \u00b0, and the radius to thickness ratio R / t from 20 to 150. We prescribe also the radius of the shell to be R = 25 mm. To ease a systematic interpretation of the results, we divide the other four geometric parameters into two groups" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001847_tro.2020.3038687-Figure11-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001847_tro.2020.3038687-Figure11-1.png", + "caption": "Fig. 11. Design and implementation of 3M3D shoulder joint.", + "texts": [ + " The maximum torque at joint 1 (shoulder adduction/abduction) and joint 2 (shoulder extension/flexion) is increased to three times as much as the motor torque, compared with the corresponding 1-unrouted motor-joint form in Table II. When only joint 2 is driven, the three motors provide equal torque, which becomes more balanced than before adjustment. Authorized licensed use limited to: University of Gothenburg. Downloaded on December 21,2020 at 07:20:41 UTC from IEEE Xplore. Restrictions apply. The implemented 3M3D shoulder joint is shown in Fig. 11. The tendons starting from motor pulley rm1 turn around the joint pulley r21 before reaching joint pulley r31, the joint pulley r31 is fixed with the stator end of motor 3. The rotation of the stator end of motor 3 acts as the rotation of joint 3. The tendon from rm2 turns around the joint pulley r12. The rotation of motor 1 acts as the rotation of joint 1, and the rotation of the stator end of motor 2 acts as the rotation of joint 2. Similarly, the tendons starting from motor pulley rm3 turn around the joint pulley r23 of joint 2 before reaching the joint pulley r13" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000178_978-3-319-09489-2_6-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000178_978-3-319-09489-2_6-Figure3-1.png", + "caption": "Fig. 3 Pulley modelling", + "texts": [ + " Is has not been simulated but the inertia effects have been included in the drum one as described in the following: \u2022 Drum inertia vs rotation axis: 0.0285 Kg m2 \u2022 Motor rotor inertia vs rotation axis: 0.015 Kg m2 \u2022 Simulated drum inertia vs rotation axis: 0.0285 + 9 \u00d7 0.015 = 0.1635 Kg m2 The pulleys are modeled with three rigid bodies: a base to be fixed to the robot frame, a bracket linked to the base with a pivot joint and the wheel of the pulley linked to the bracket with another pivot joint (Fig. 3). The 3d mesh of the wheel is used for the collision interaction with the cable. The collision interaction for the base and the bracket are deactivated. Dry and viscous friction parameters of the pivot joints can be tuned. The position and the orientation of a pulley are arguments of the constructor of the \u201cPulley\u201d class. The mass parameters are: \u2022 Bracket mass: 3.456 Kg \u2022 Wheel mass: 1 Kg The cable interfaces with the platform are implemented in the \u201cCable_Ball_Joint\u201d class. They are made of two rigid bodies, the base and the fastening part, linked together by two pivot joints forming a U-joint (Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001739_tec.2020.2995880-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001739_tec.2020.2995880-Figure1-1.png", + "caption": "Fig. 1 The configuration of the AFPMM", + "texts": [ + " In section III and IV, the back EMF and cogging torque of AFPMMs under different mixed eccentricity conditions are calculated, respectively. Compared with the FEA, the analytical model is faster and more flexible in the analysis of AFPMMs with the mixed eccentricity with the similar accuracy. Due to the limitation of the experiment condition, a 24-pole-36-slot AFPMM is only measured in healthy condition to indirectly verify the proposed analytical model. The configuration and the main parameters of the AFPMM with fractional slot concentrated windings are shown in Fig. 1 and Tab. I, respectively. The sketch map of an AFPMM with the mixed eccentricity is shown in Fig. 2. In this figure, e denotes the maximum distance caused by the angular eccentricity and dr denotes the rotor radial deviation distance. The rotation center of the AFPMM with mixed eccentricity is located in the rotor geometric center. Firstly, the healthy quasi-3-D analytical model is given, which is taken as the foundation to establish the analytical model of AFPMMs with the mixed eccentricity. The model of the AFPMM is divided into a number of layers along the radial direction, and every layer is extended and regarded as a 2-D linear motor model" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002152_13506501211010030-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002152_13506501211010030-Figure1-1.png", + "caption": "Figure 1. Diagram of the structure of the intermediate gearbox.", + "texts": [ + " Furthermore, what would happen when considering the mesh of spiral bevel gears and the interaction between housing wall and gears? To date, there is still little knowledge about this. In this study, a CFD model of spiral bevel gears under splash lubrication was established. Then, the result of the simulation was compared with the experimental result. On this basis, the churning power loss of the intermediate gearbox under splash lubrication was calculated and the optimum parameters of rotational speed and oil immersion depth were proposed. The structure of the intermediate gearbox is shown in Figure 1. It mainly consists of a pair of spiral bevel gears, double row conical roller bearing, gearbox housing, end covers and air vent. In flat flight, the centre line of the driving gear is horizontal, whereas the driven gear is inclined. The lubricating oil is at the bottom of the gearbox housing. Gears splash the oil to the meshing place. As the bearings generate substantial heat and the gears cannot precisely transport oil to the bearing, two oil guide pipes are set up at the left and right end covers to ensure normal bearing lubrication" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000447_20140824-6-za-1003.01586-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000447_20140824-6-za-1003.01586-Figure1-1.png", + "caption": "Fig. 1. Fully Assembled Variable Pitch Quadrotor Prototype", + "texts": [ + " (2012)). The pitch actuation of the rotor blades addresses the issue of the relatively slow motor time constant for acrobatic flight but does not alleviate the large mass moment of inertia caused by having four motors on the outside of the vehicle frame. This paper seeks to remedy the latter issues by the development of a novel variable pitch quadrotor ? Financial support of the National Research Foundation and the University of Cape Town is gratefully acknowledged. with low inertia, shown in Fig. 1, designed for performing high angle of attack manoeuvres. The design uses two centrally located motors, spinning at the same speed in opposite directions, to drive each rotor pair at matched speeds via torque shafts. The thrust of each rotor blade is controlled by changing the blade pitch and this allows for high bandwidth thrust vectoring. The quadrotor is also designed to minimise the mass and rotational inertia of the vehicle and to remove attitude coupling from centrifugal and gyroscopic effects" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002152_13506501211010030-Figure6-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002152_13506501211010030-Figure6-1.png", + "caption": "Figure 6. Gear transmission test rig.", + "texts": [ + " The initial oil distribution of the intermediate gearbox is shown in Figure 5. The simulation was calculated up to 20 rotations of the driving bevel gear. The central processing unit (CPU) time was about 6.5 days with 48 CPUs. The computational effort is as follows: 10% for the three-dimensional model establishment; 20% for pre-processing of numerical calculation, including model simplification, calculation domain model establishment and meshing; and 70% for numerical solution. A gear transmission test rig is shown in Figure 6. It consists of an electric cabinet, a motor, plum couplings, a torque\u2013speed sensor and a test gearbox. The spiral bevel gears are lubricated by splashing oil and the angular contact ball bearings are lubricated by injection oil. The end of the output shaft of the gearbox is set to be free (zero-load condition) to minimise the friction loss caused by the contacts of gears and bearings. The fundamental parameters of the spiral bevel gears in the test gearbox are presented in Table 3. The density of the lubricating oil is 875" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000893_j.ymssp.2015.10.028-Figure13-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000893_j.ymssp.2015.10.028-Figure13-1.png", + "caption": "Fig. 13. Visual representation of the non-deformed (gray) and the deformed (the range of colors illustrates the strength of the tangential strain) areas of the ring gear at nominal load (the deformation is scaled by factor 1000). (For interpretation of the references to color in this figure legend, the reader is referred to the web version of this article.)", + "texts": [ + " In a direct comparison between the results of both partial models it can be found that the graph of the tangential strain in the area around the meshing becomes more discontinuous due to the additional drill holes. Notably, the drill hole above the middle tooth causes an increase of the strain peak, which could either be caused by a displacement of the strain peak along the circumference or by the attenuation of the cross section of the ring gear. Besides the global peak, there are local peaks to its left and to its right side. Similar characteristics arise around the other two drill holes. Finally, the simulation results of the deformed and the non-deformed ring gear with holes are compared in Fig. 13. In the following, the acquired measurement data is analyzed regarding the defined objectives of the measurement campaign. The following measurement plot in Fig. 14 shows an example of the strain signal recorded in AC coupling at the 12 o'clock position of the ring gear plotted against the rotation angle of the main shaft. The rotating direction is clockwise when seen from the front of the WEC. The rotation angle is recorded by an inductive strap transmitter. The offset is due to the mounting of the strap transmitter" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000991_978-3-319-51532-8_2-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000991_978-3-319-51532-8_2-Figure1-1.png", + "caption": "Fig. 1 Illustration of a flying wing tailsitter vehicle. The vehicle is actuated by two propellers that produce forces along the body z-axis, and two flaps that produce aerodynamic forces by deflecting the airflow over the wings", + "texts": [ + " Typically, the complex aerodynamic properties are obtained either by CFD methods [17, 20], by measurement series covering all relevant operating points [21], or by first-principles models combined with heuristics that capture some of the unmodeled effects [22]. Since herein we focus on control design, we follow a similar approach as proposed in [22] and derive a first-principles model of the considered tailsitter vehicle. The vehicle is actuated by two propellers, one in front of each wing, and two flaps located at the wings\u2019 trailing edges. An illustration of the tailsitter is shown in Fig. 1. The control inputs are the propeller forces f prop,l and f prop,r , and the flap angles \u03b4 f lap,l and \u03b4 f lap,r . All four control inputs are subject to saturations: f prop,min \u2264 f prop,l , f prop,r \u2264 f prop,max , \u03b4 f lap,min \u2264 \u03b4 f lap,l , \u03b4 f lap,r \u2264 \u03b4 f lap,max . (1) We assume that the vehicle\u2019s airspeed is small, such that the range of attainable propeller forces [ f prop,min, f prop,max ] can be considered to be constant. The propellers are of fixed-pitch type and the motors are not able to reverse direction mid-flight, meaning that the propellers cannot produce negative thrust, i.e. f prop,min > 0. We introduce a body-fixed coordinate frame B with origin at the vehicle\u2019s center of mass, as shown in Fig. 1. The z-axis of the body frame B is aligned with the thrust direction, the x-axis points along the left wing, and the y-axis completes the right-handed coordinate system. We denote unit vectors along the axes of the body frame as eBx , eBy , and eBz , respectively. The position of the vehicle\u2019s center of mass relative to an inertial coordinate frame I , expressed in this inertial frame I , is denoted as I p = (px , py, pz). (In order to simplify notation, vectors may be expressed as ntuples x = (x1, x2, " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000813_j.phpro.2014.08.144-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000813_j.phpro.2014.08.144-Figure2-1.png", + "caption": "Fig. 2. Schemetic representation of 3D molten pool shape with dynamic local coordinate system.", + "texts": [ + " A laser beam with a Gaussian profile is used as the heat source, which is focused on the substrate surface with a spot size of 0.6 mm at a distance of 9 mm below the nozzle tip. In the above coaxial LPD processing, the transient evolution the molten pool shape corresponds to the temperature field and is updated at every time step in the simulation. Based on the temperature field and dynamic molten pool shape calculated from the above 3D mathematical model, the crystal growth and microstructure formation in the molten pool can be predicted by an established crystal growth model. Figure 2 shows the geometric relationship of variables used in the following derivation of crystal growth model. The crystal growth and microstructure formation in molten pool responds to the solidification conditions ahead of the advancing solidification interface[10]. The criterion which states the microstructure is columnar dendrite can be derived when the following condition is satisfied everywhere in the molten pool[4]: n c CET n nIn NaK V G 1 1 13 4 3 0 (1) where G is the temperature gradient, V is the solidification velocity of the columnar dendrite tip and KCET is the critical value of CET", + " During the LPD processing conditions, the normal temperature gradient Gn ahead of solidification front can be expressed as 222 zyxn GGGG (2) where Gx, Gy and Gz are temperature gradient along X, Y and Z axis direction, and can be calculated from the temperature field of mathematical model. The normal solidification velocity Vn ahead of solidification front is linked geometrically to the heat source travel speed Vb by the angle between the normal direction of the solidification interface and the travel direction of the heat source, as given by the following equation, as shown in Fig. 2: cosbn VV (3) Since the normal temperature gradient Gn and normal solidification velocity Vn are parallel with same direction, and laser scans along positive X axis direction which also is parallel to Vb, the angle can also be described as n x G G cos (4) In the present analysis, it is assumed that the growth of dendrites is epitaxial from the substrate and the SX nature of the substrate is thus maintained after the LPD process. For the face-centered cubic (FCC) nickel-based superalloys, the six direction ((001)/[100], (00-1)/[100], (010)/[100], (0-10)/[100], (100)/[100], (-100)/[100]) are the preferred growth directions" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001925_978-3-030-46817-0_30-Figure27-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001925_978-3-030-46817-0_30-Figure27-1.png", + "caption": "Fig. 27. Stress validation of FEA model of ball joint.", + "texts": [], + "surrounding_texts": [ + "Figs. 23, 24, 25 and 26. 142 I. Alagi\u0107 In case of ball joint, the results of stress distribution and allowed displacement are presented in Figs. 27, 28, 29, 30, 31 and 32. Finite Element Analysis (FEA) of Automotive Parts Design as Important Issue 143 The displacement achieved as a result of finite element analysis doesn\u2019t much differ from results of the laboratory-test performed by control device MR 96. The maximum displacement appeared into Z direction 0,145 [mm]. On the basis of conducted simulations were possible to affirm that the magnitude of deformations depends on model geometry. The largest concentration of stresses appeared in places near of cover of ball joint. Through change of model geometry it was possible to influence on expansion of stresses and displacement distribution. To analysis of maximum warp angle of ball joint using FEA solver was very low compared to its allowed value 58 \u00b1 6\u00b0." + ] + }, + { + "image_filename": "designv10_14_0001352_j.mechmachtheory.2016.02.015-Figure5-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001352_j.mechmachtheory.2016.02.015-Figure5-1.png", + "caption": "Fig. 5. Coordinate system applied for generating gear.", + "texts": [ + " Theoretically, the curvature parameters and the unit normal vectors of the pinion tooth surfaces determined by the two above-mentioned methods should be identical, which is represented by a system of equations as A 0 1d \u00bc A1d A 0 1c \u00bc A1c B 0 1d \u00bc B1d B 0 1c \u00bc B1c C 0 1d \u00bc C1d C 0 1c \u00bc C1c n 0 1d \u00bc n1d n 0 1c \u00bc n1c 8>>>>>>>< >>>>>>>: \u00f013\u00de The basic machine settings for the pinion can be obtained by solving Eq. (13). Details of application of the mentioned five steps of the basic machine settings for formate spiral bevel and hypoid gears generated by the duplex helical are given as follows. Coordinate system Sm{Xm, Ym, Zm} is rigidly connected to the cutting machine (Fig. 5). The top, bottom (A\u2013A) and right (B\u2013B) of Fig. 5 are the machine front view, the machine bottom view and the side view (the projection of head cutter), respectively. The cradle rotates about the Ym-axis; the p-axis and g-axis are projections of gear and pinion axes in the XmOmYm-plane. The points M, M1 and O0 are the reference points of the tooth surface and the projection of M on the cutting edge and the center of the head cutter, respectively. O2 is the cross point of the gear, and Om is the machine center. Some of the basic machine-tool settings for generating gear are: the machine root angle of gear \u03b3m2, the machine center to back of gear XG, the horizontal and vertical setting (H2 and V2) of the gear head-cutter, the cradle angle of gear q2, the radial distance of gear Sr2", + " The unit vectors of the Ym-axis, g-axis, and p-axis can be represented in coordinate system Sm by the following equations: e \u00bc 0 1 0\u00bd \u00f014\u00de g \u00bc \u2212 cos \u03b3m2 0 sin \u03b3m2\u00bd \u00f015\u00de p \u00bc \u2212 cos \u03a3\u2212\u03b3m2\u00f0 \u00de 0 \u2212 sin \u03a3\u2212\u03b3m2\u00f0 \u00de \u00f016\u00de The unit normal vector n0, the unit vector t0 and the position vector a0 of any point on the outside blade edge of the gear to the head-cutter generating surface (drive or convex side) can be defined in the coordinate system Sm by the following equations: n0 \u00bc \u2212 cos \u03b121 sin \u03b202 cos\u03b121 cos \u03b202 sin \u03b121\u00bd \u00f017\u00de t0 \u00bc \u2212 sin \u03b121 sin \u03b202 \u2212 sin \u03b121 cos \u03b202 cos\u03b121\u00bd \u00f018\u00de a0 u\u00f0 \u00de \u00bc ut0 \u00fe H\u2212XG cos \u03b3m2 \u00fe rcG1 sin\u03b202 rcG1 cos \u03b202\u2212V XG sin \u03b3m2 2 4 3 5 T \u00f019\u00de Here, u is a profile parameter. The position vector of point M1 on the blade edge can be represented as a0(sG1) when u is equal to sG1 (Fig. 5). According to step 1, the first principal curvature Ag of the generating tooth surface for the drive side is represented as Agd \u00bc 1 R2d \u00f020\u00de Here, R2d denotes the curvature radius of the inside blade of the gear head-cutter. The first principal torsion Bgd and the second principal curvature Cgd of the gear tooth surface are equal to 0. The formate-cut gear tooth surface is copy of the surface of the head-cutter, which is a surface of revolution, the vector a0(sG1), n0, t0 and the curvature parameters of the gear surfaces (A2d = Agd, B2d = 0, C2d = 0) are the same as those of gear tooth surface", + " The position vector r2 and the unit normal n2 of the coast side of the pinion tooth surface at the reference point M2 \u2032 , can be represented as r2 \u00bc r1c R p; \u03b82\u00bd \u00f024\u00de n2 \u00bc n1c R p; \u03b82\u00bd \u00f025\u00de Here, R[p, \u03d12] is a transformation matrix that denotes the rotation angle \u03d12 about the vector p. According to step 4, the curvature parameters of the pinion tooth surface based on the tangency of the generating and pinion tooth surfaces at the reference points are determined in this section. The configuration in Fig. 7 is the same as that in Fig. 5. Coordinate systems Sm{Xm, Ym, Zm} are rigidly connected to the cutting machine. The top (A\u2013A), bottom and middle of Fig. 7 are the machine's front view, top view and side view (projection of the head cutter), respectively. The cradle rotates about the G-axis. The p-axis is the unit vector of the pinion spindle. O1 is the cross point of the pinion. The manufacturing coordinate systems of the gears and pinions and the installation coordinate system of the hypoid gear sets are represented in the coordinate system Sm. Some of the basic machine-tool settings for generating pinion are: the machine root angle \u03b3m1, the machine center to back Xp, the cradle angle q1, the radial distance Sr1, the blank offset Em1, the sliding base Xb1, the swivel angle J1, the tilt angle I1, the horizontal and vertical setting (Hp and Vp) of pinion head-cutter. The pinion cradle spindle G does not always coincide with the Ym-axis (Fig. 5), and it can be represented in the coordinate system Sm.as G \u00bc G \u03b1G1;\u03b3m1\u00f0 \u00de \u00f026\u00de Here, \u03b1G1 is the pressure angle of the drive side of the pinion generating surface. The unit vector of the blank offset direction of the generating gear and the pinion is represented as ep \u00bc ep \u03b1G1;\u03b3m1\u00f0 \u00de \u00f027\u00de The position vector of the reference point M1 is represented in the coordinate system S{ep \u00d7 G, G, ep} as ap1 Xp; Em1 \u00bc r1d \u00fe Xpp\u00fe Em1ep \u00f028\u00de The meshing equation for the generating gear and pinion tooth surfaces at M1 may be represented as f p1 \u00bc f p1 \u03b1G1;\u03b3m1;Xp; Em1;Hl;Ra1 \u00bc vp1 n1d \u00bc 0 \u00f029\u00de Here, the relative velocity vp1 of the generating gear and pinion tooth surfaces at M1, Hl is the helical motion velocity coefficient, it represents a displacement of the pinion blank along the axis of the cradle for a rotational angle of 1 rad of the cradle, Ra1 is the ratio-of-roll of pinion and generating gear" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000769_978-3-662-43893-0-Figure3.21-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000769_978-3-662-43893-0-Figure3.21-1.png", + "caption": "Fig. 3.21 Bevel gear set with positive offset in \u201ccoast\u201d mode", + "texts": [ + " only the components which lie in the drawing plane are shown in Fig. 3.20. Force F1 is split into two components FH and FV. In this configuration, the tooth forces reduce the hypoid offset and shift the pinion axially, thereby increasing its mounting distance tB. A tapered roller bearing in an O-configuration is usually mounted behind the pinion such that the tooth forces press the pinion against the bearing. This mode of operation is termed \u201cdrive\u201d, its opposite being \u201ccoast\u201d (see Sect. 2.2.4.5). 3.4 Displacement Behavior 81 If Fig. 3.21, the bevel gear set from Fig. 3.20 is shown in coast mode. In this case, the concave tooth flank of the wheel drives the convex tooth flank of the pinion. In the opposite direction of rotation, the convex pinion tooth flank drives the concave wheel tooth flank. The pinion exerts force F2 on the tooth of the wheel, which reacts with force F1 on the pinion. Force F1 is split into components FH and FV. In coast mode, the tooth forces cause an increase in the hypoid offset and a reduction in the mounting distance, i" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000525_s12541-016-0036-6-Figure7-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000525_s12541-016-0036-6-Figure7-1.png", + "caption": "Fig. 7 Definition the gear and pinion coordinate systems", + "texts": [ + " Simultaneously, each contact line should be segmented and numbered from root to tip insequence, li(j) j = 1, \u2026, m, as shown in Fig. 6. For describing deviationsof real tooth surfaces, it is required to measure the actual tooth surface by CMM. Unlike the measurement performed by scanning the line in both the profile direction and the tooth trace direction on the tooth surface, the measuring lines are set in contact lines direction, as shown in Fig. 6. Given the possibility to numerical analysis of the results from 3D measurement with data obtained by CMM, we may define the gear and pinion coordinate systems, as shown in Fig. 7. The coordinates and unit normal vector of measurement point on the gear surface are (xg, yg, zg) and (nxg, nyg, nzg), respectively. The coordinates and unit normal vector of measurement point on the pinion surface are (xp, yp, zp) and (nxp, nyp, nzp), respectively. The deviations between the actual tooth surface and the reference tooth surface can be represented as (18) (19) Where, \u0394\u03b8g and \u0394\u03b8p denote the angular deviations of the gear and \u2202f \u2202ug ------- \u239d \u23a0 \u239b \u239e i 1+ f ugi u\u0394+ \u03b8gi,( ) f ugi \u03b8gi,( )\u2013 u\u0394 ----------------------------------------------------------= \u2202g \u2202ug ------- \u239d \u23a0 \u239b \u239e i 1+ g ugi u\u0394+ \u03b8gi,( ) g ugi \u03b8gi,( )\u2013 u\u0394 -------------------------------------------------------------= \u2202f \u2202\u03b8g ------- \u239d \u23a0 \u239b \u239e i 1+ f ugi \u03b8gi \u03b8\u0394+,( ) f ugi \u03b8gi,( )\u2013 \u03b8\u0394 ----------------------------------------------------------= \u2202g \u2202\u03b8g ------- \u239d \u23a0 \u239b \u239e i 1+ g ugi \u03b8gi \u03b8\u0394+,( ) g ugi \u03b8gi,( )\u2013 \u03b8\u0394 -------------------------------------------------------------= f ugi \u03b8gi,( ) 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000469_s11668-014-9893-4-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000469_s11668-014-9893-4-Figure2-1.png", + "caption": "Fig. 2 A schematic line diagram of the MFS Fig. 3 Machine fault simulator", + "texts": [ + " Hsu and Lin [6] showed a comparison of different multiclass SVM methods and concluded that \u2018one-against-one\u2019 is a competitive method for the multiclass classification. Here LIBSVM [3] software has been used for multiclass classification bearing with OAO approach. Now in the next section, experimental setup and data acquisition are described. Experiments were performed on Machinery Fault Simulator (MFSTM), which was capable of simulating a range of machine faults such as in the gear box, rolling bearings, motors, shaft resonances, and shaft misalignments. The schematic diagram of MFS is shown in Fig. 2. In the MFS experimental setup, a three-phase induction motor was connected to the shaft through a flexible coupling with a healthy bearing near the motor end and a faulty (test) bearing at the other end of the shaft as shown in Fig. 3. This setup allowed the study of bearing defects by introducing faulty bearings in the machine and then studying its vibrational signature. In the study of faults in bearings, five different types of bearing fault condition were considered as shown in Fig. 4 (i) no defect bearing (ND), (ii) inner race fault bearing (IRF), (iii) outer race fault bearing (ORF), (iv) bearing element fault (BEF), and (v) combination fault bearing (CFB)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000037_751476-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000037_751476-Figure1-1.png", + "caption": "Figure 1: Model of a tooth.", + "texts": [ + " Equations (1)\u2013(7) show the computing equations consisting of tooth bending stiffness, shear stiffness, axial compressive stiffness, foundation stiffness, and Hertzian contact stiffness. Bending stiffness \ud835\udc58\ud835\udc4f can be obtained as 1 \ud835\udc58\ud835\udc4f = \u222b \ud835\udf11 2 \u2212\ud835\udf11 1 ( (3[1 + (\ud835\udf112 \u2212 \ud835\udf11) sin\ud835\udf11 cos\ud835\udf111 \u2212 cos\ud835\udf11 cos\ud835\udf111] 2 \u00d7 (\ud835\udf112 \u2212 \ud835\udf11) cos\ud835\udf11) \u00d7 (2\ud835\udc38\u0394\ud835\udc59[(\ud835\udf112 \u2212 \ud835\udf11) cos\ud835\udf11 + sin\ud835\udf11]3) \u22121 ) d\ud835\udf11. (1) Shear stiffness \ud835\udc58\ud835\udc60 is calculated by 1 \ud835\udc58\ud835\udc60 = \u222b \ud835\udf11 2 \u2212\ud835\udf11 1 1.2 (1 + ]) (\ud835\udf112 \u2212 \ud835\udf11) cos\ud835\udf11cos 2 \ud835\udf111 \ud835\udc38\u0394\ud835\udc59 [(\ud835\udf112 \u2212 \ud835\udf11) cos\ud835\udf11 + sin\ud835\udf11] d\ud835\udf11. (2) Axial compressive stiffness \ud835\udc58\ud835\udc4e is 1 \ud835\udc58\ud835\udc4e = \u222b \ud835\udf11 2 \u2212\ud835\udf11 1 (\ud835\udf112 \u2212 \ud835\udf11) cos\ud835\udf11sin 2 \ud835\udf111 2\ud835\udc38\u0394\ud835\udc59 [(\ud835\udf112 \u2212 \ud835\udf11) cos\ud835\udf11 + sin\ud835\udf11] d\ud835\udf11, (3) where \u210e, \ud835\udc65, \ud835\udc51, \ud835\udf111, \ud835\udf112, and \ud835\udf11 are shown in Figure 1. \ud835\udc38, \ud835\udc3a, and V represent Young\u2019s modulus, shear modulus, and Poisson\u2019s ratio, respectively. The stiffness of tooth foundation can be obtained by 1 \ud835\udc58\ud835\udc53 = cos2\ud835\udf111 \ud835\udc38\u0394\ud835\udc59 \u00d7 {\ud835\udc3f \u2217 ( \ud835\udc62\ud835\udc53 \ud835\udc46\ud835\udc53 ) 2 +\ud835\udc40 \u2217 ( \ud835\udc62\ud835\udc53 \ud835\udc46\ud835\udc53 ) + \ud835\udc43 \u2217 (1 + \ud835\udc44 \u2217tan2\ud835\udf111)} , (4) where \ud835\udc62\ud835\udc53 and \ud835\udc46\ud835\udc53 are given in Figure 2. The coefficients \ud835\udc3f\u2217, \ud835\udc40 \u2217, \ud835\udc43\u2217, and \ud835\udc44\u2217 can be calculated by polynomial functions [26] as follows: \ud835\udc4b \u2217 (\u210e\ud835\udc53, \ud835\udf03\ud835\udc53) = \ud835\udc34 \ud835\udc56 \ud835\udf03 2 \ud835\udc53 + \ud835\udc35\ud835\udc56\u210e 2 \ud835\udc53 + \ud835\udc36\ud835\udc56\u210e\ud835\udc53 \ud835\udf03\ud835\udc53 + \ud835\udc37\ud835\udc56 \ud835\udf03\ud835\udc53 + \ud835\udc38\ud835\udc56\u210e\ud835\udc53 + \ud835\udc39\ud835\udc56, (5) where \u210e\ud835\udc53 = \ud835\udc5f\ud835\udc53/\ud835\udc5fint. \ud835\udc5f\ud835\udc53, \ud835\udc5fint and \ud835\udf03\ud835\udc53 are defined in Figure 2 and the constants \ud835\udc34 \ud835\udc56, \ud835\udc35\ud835\udc56, \ud835\udc36\ud835\udc56,\ud835\udc37\ud835\udc56, \ud835\udc38\ud835\udc56, and \ud835\udc39\ud835\udc56 are given in Table 1" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002015_j.asr.2020.09.040-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002015_j.asr.2020.09.040-Figure1-1.png", + "caption": "Fig. 1. Basic idea of the folding pattern of a flat membrane with hexagonal boundaries.", + "texts": [ + " the set of experiments using the tensile deployment based on force applied in the radial and circumferential directions on paper-based membranes have confirmed that (1) the multi-spiral with prismatic folding lines offered the improved deployment performance when compared to other spiral approaches, and (2) the deployment in curved surfaces progressed rapidly within a finite load domain. In the next sections, we describe our proposed approach, our experiments and obtained insights. In this section, we describe the governing equations of the multi-spiral folding pattern on a flat membrane. The spiral folding pattern wraps a flat membrane into a cylinder as basically portrayed by Fig. 1. Here we describe the basic idea of the spiral folding pattern and its governing equations. Generally speaking, given a flat membrane and a regular polygon positioned in its core (or hub), creases on the flat membrane are tangential to the core polygon, and the fold behaviour transforms the flat membrane into a hollow cylinder having a regular polygon in its core. Along the transition process from flat to cylinder configuration, zigzag patterns are generated in the cylindrical shape which is due to the shape and configuration of the mountain and valley folds of the membrane, as show in Fig. 1. An example of the layout structure of the spiral fold pattern is shown by Fig. 2(a), and an example of a wrapped configuration using paper is shown by Fig. 2(b). In order to portray the governing equations of the spiral folding pattern, we first refer to Fig. 3, which describes the key elements involved in the rendering of the crease geometry and the overall folding structure. By observing Fig. 3, given a regular polygon with center at origin O, the crease Fold A is tangential to the edge of the core polygon" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002384_j.measurement.2021.109394-Figure9-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002384_j.measurement.2021.109394-Figure9-1.png", + "caption": "Fig. 9. Variation of contact force in load zone and its corresponding angle representing the line of action.", + "texts": [ + " The strategy of estimating defect size, based on methodology combining mechanics and signal processing-based approaches, requires estimation of force on a rolling element when it interacts with the leading edge of the defect[28]. An empirical approach for obtaining this force by correlating it with peak acceleration was proposed and verified experimentally in Ref. [28]. For the experimental verification of the approach, the peak acceleration with maximum amplitude is assumed to have been generated when the defect and rolling element interact when the defect is the most heavily loaded point in the load zone. This instance is represented by \u03c6 = 0 in Fig. 9(a), where the force on the rolling element is maximum(Fmax) as shown in Fig. 9(b). Based on the magnitude of other peaks in the acceleration signal, they are empirically correlated to angle \u03c6 from which force on the rolling element is estimated. For this, a series of steps are involved. In the first step, the maximum mean peak acceleration is assigned the load zone location of maximum load \u03c6 = 0. Then, lower threshold corresponding to 20% of maximum acceleration value is computed and assumed to be generated when the defect is located at 20% extent (Fig. 2(b)) of the load zone (\u03c6 = 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002144_s12206-021-0406-6-Figure11-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002144_s12206-021-0406-6-Figure11-1.png", + "caption": "Fig. 11. (a) Planetary gearbox test rig; (b) ring gear; (c) fault size.", + "texts": [], + "surrounding_texts": [ + "In the previous section, we have compared and verified the correctness of the improved phenomenological model in the time domain and frequency domain, respectively through noiseless simulation signals. However, the background noise is inevitable for planetary gearbox operating under complex working conditions. Therefore, it is necessary to denoise the vibration signal when verifying the correctness of the improved model with experimental signals. The following is a brief introduction to the noise reduction algorithm used in this paper. The maximum correlation kurtosis deconvolution (MCKD) was proposed by Mc Donald et al. [32] and applied to the compound faults of gears and bearings of fixed-axis gearboxes. This method utilizes the periodic characteristics of the fault and the pulse-like vibration characteristics related to the fault. Since this method can extract the weak periodic impact component of the signal when the signal-to-noise ratio is very low, it has been widely used in the fault diagnosis of various rotating machinery [33, 34]. Also, some scholars [35, 36] proposed corresponding improved methods for the parameters of MCKD to satisfy the parameter adaptive, such as shift number M, deconvolution period T, etc. In this paper, the algorithm of MCKD is used to reduce the noise of the vibration signal of the planetary gearbox and highlight its impact characteristics. The flow chart of the paper in this part is shown in Fig. 10. In order to obtain sufficient vibration and impact information, a high sampling rate fs is used in this paper. This will result in low efficiency when directly performing noise reduction processing on the original signal x(n). Besides, the research goal of the paper is to calculate the phase between the fault impact and its adjacent meshing impact, so the signal segment containing these two impacts in the vibration signal is worthy of attention. Thus, the paper proposes to divide the original signal into pieces and extract the segments xk(n0) containing the two kinds of impacts, and then combine them into a new signal y(m). Where, the signal segment xk(n0) is centered on the position l(k) corresponding to the peak value of the k-th fault impact, and its length is 2NTm. NTm represents the number of sample points included in a meshing period and NTm = fs\u00b7Tm. Assuming that K segments of signals are extracted from the original signal, the length of the reconstructed signal y(m) is Lm = 2K\u00b7NTm; and the length of the original signal x(n) is Ln \u2265 (Zr /N)\u00b7K\u00b7NTm. Taking the gearbox in the paper as an example (Zr = 62, N = 3), the ratio of the length of the two signals is Ln /Lm \u2265 10.33, which greatly improves the efficiency of the noise reduction algorithm. In addition, the number of signal segments K can be adjusted by setting the threshold value AT. Setting a large threshold will reduce K and improve subsequent computing efficiency, but it may cause misjudgment. Therefore, it is necessary to set an appropriate threshold value AT according to the analyzed signal. It is obvious that the reconstructed signal y(m) includes both the fault impacts and meshing impacts, and it can be regarded as a compound fault condition. Although MCKD can be used for compound fault separation of rotary machinery [32], it is not suitable for the signal studied in this paper. The main reasons are as follows: 1) In the condition of broken tooth fault of the ring gear, the amplitudes of the fault impact and the meshing impact are very different; 2) The periods of these two kinds of impacts in the reconstructed signal y(m) are the same. For these reasons, the paper proposes to extract these two impacts separately. Let x'k(n0) = xk(n0), and replace the values of the sample points corresponding to the k-th fault impact with noise. By this method, the reconstructed signal y'(m) without fault impacts is obtained. At this time, the MCKD with the same parameters are used to enhance the impacts of the two signals y(m) and y'(m). Then the output signals yout(m) and y'out(m) will be mainly the fault impacts and meshing impacts, respectively. Extract the time when these two types of impacts occur from yout(m) and y'out(m), and calculate the phase difference \u2206T between these two impacts. Finally, according to this phase and the assisted phase proposed above, the vibration mechanism of the gear pair and the correctness of the improved phenomenological model are verified." + ] + }, + { + "image_filename": "designv10_14_0001741_j.autcon.2020.103264-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001741_j.autcon.2020.103264-Figure3-1.png", + "caption": "Fig. 3. Concept of VSD hypothesis for controlling a position of the blade cutting edge.", + "texts": [ + " The distance between the end effector of manipulator and the target position is modeled as a virtual spring and damper, which moves the end effector to the target position through control method. This is generally applied to systems such as multi-DOF manipulators with a DOF of 6 or higher to control manipulators without a complex method of inverse kinematics analysis [21,22]. This study applied the Cartesian space control algorithm based on the VSD-hypothesis to a bulldozer to control the position of the blade cutting edge. When a bulldozer is on the current landform that requires leveling as shown in Fig. 3, it must move forward and control the position of the blade cutting edge to make that into the designed surface. At this time, two points for blade control are defined at the equilibrium point of a bulldozer. One of these points represents the desired position where the blade cutting edge should ultimately be located. This point is located on the design surface and it is presented at arbitrary distance spacing when the bulldozer moves. The other point represents the position of the current blade cutting edge" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001319_1687814020940072-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001319_1687814020940072-Figure3-1.png", + "caption": "Figure 3. The 6-UPU Stewart platform.", + "texts": [ + " Therefore, while adjusting the pose of the subreflector, the Stewart platform follows the primary reflector in elevating motion. The elevation range of primary reflector during observations is 58 888, and the elevation rotatable range is 08 908. As shown in Figure 2, the Stewart platform for the NSRT consists of a base platform, a moving platform, and six extendable actuators. Throughout this article, we assume that the base platform and the moving platform are both rigid and all six legs are identical. As shown in Figure 3, the piston of ith (i=1,2,3,4,5,6) leg is attached to the moving platform with an universal joint at point pi, and the cylinder is attached to the base platform with an universal joint at point bi. For the convenience of expressions, the Cartesian coordinate frame xPyPzP is fixed on the moving platform at origin P, and the Cartesian coordinate frame xByBzB is fixed on the base platform at origin B. The pose of the moving platform is defined by a six-dimensional (6D) vector \u00bdx, y, z,a,b, g , where the first three components are the coordinates of frame P in the reference frame B and the last three components are parameters describing the orientation of the platform" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000862_303-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000862_303-Figure2-1.png", + "caption": "Figure 2. Optical system for measuring transmission error.", + "texts": [ + " This was not practicable, however, because of the difficulty of measuring angular position continuously with the necessary accuracy (this difficulty has subsequently been overcome at the National Engineering Laboratory and is reported by Timms (1960)). An optical method of measuring the transmission error directly is described in the next section. This method depends essentially on the fact that the rotation of one gear is the same, apart from the small difference which has to be observed, as that of the other when viewed in a mirror. 3. The optical system The optical arrangements are shown diagrammatically in figure 2. The two shafts carrying the test gears are extended to the outside of the gearbox and a Perspex disk is fixed to the end of each. The disks are engraved with 500 radial lines, 1 in. long and 0,020 in. wide, filled in so that they are opaque. The lines are at 5.75 in. mean diameter and the mean pitch is therefors 0.036 in. Forty-eight 6v, 0 . 1 ~ filament bulbs placed between one disk and the gearbox form an annular source of light illuminating the disk. An image of the illuminated gaps between the engraved lines on this disk is focused concentrically on the second disk using a lens and two plane mirrors as shown in the figure" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001772_s00500-020-05202-1-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001772_s00500-020-05202-1-Figure1-1.png", + "caption": "Fig. 1 Line-loading mode of PLI robot", + "texts": [ + " However, it can be concluded from the simulation results that GT2FSMC remains to be the most useful method because general type-2 fuzzy logic control can handle uncertainties more effectively. This paper is organized as follows. Problem formulation is presented in Sect. 2. In Sect. 3, general type-2 fuzzy logic system is briefly introduced. Section 4 describes the design procedure ofGT2FSMC. In Sect. 5, the simulation results are presented to show the effectiveness of the proposed method. Finally, conclusions and future work are given in Sect. 6. In this section, the dynamicmodel of PLI robotmanipulator is presented. The line-loading model of the PLI robot is shown in Fig. 1. In Fig. 1, the PLI robot is working on the label; therefore, it will be really sensitive to winds, shaking of label and friction. In this paper, the various environment factors are regarded as the external disturbances, and the final control goal is to control the PLI robot keep balanced operating on the label. The balance adjustment parameters of the PLI robot are shown in Fig. 2. The physical meaning of each parameter is shown in Table 2. As can be more clearly seen in Fig. 2, the control target is to adjust \u03b81 in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001762_j.triboint.2020.106536-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001762_j.triboint.2020.106536-Figure1-1.png", + "caption": "Fig. 1. Churning parameters for the spiral bevel gear with partial immersion in lubricating oil.", + "texts": [ + " Moreover, the typical theoretical results obtained from the proposed quasi-analytical model are compared and validated with the experimental data from the existing literature. A spinning spiral bevel gear under splash lubrication is associated with the interactions of the surrounding oil and air, which produce the no-load power losses expressed as the sum of churning losses and windage losses, i.e. Pall\u00bcPc \u00fe Pw (1a) where Pall refers to overall no-load losses, Pc denotes the churning loss due to oil drag, Pw represents the windage losses due to air drag. The physic-based churning loss model of a partial-immersion spiral bevel gear, as shown in Fig. 1, decomposes the losses into three categories expressed as Pc\u00bcPcp \u00fe Pcf \u00fe Pcrf (1b) where Pcp denotes the power loss caused by the oil acting on pitch cone\u2019s flank of the spiral bevel gear; Pcf refers to the power loss in the toe/heel; Pcrf represents the oil drag power loss induced within the tooth space/ cavity. With each loss determined, the total churning power losses Pc could be then obtained. Following the periphery oil churning loss formulation proposed in Seetharaman et al. [16,17], the drag torque is the fluid mechanical force exerted by the movement of lubricating oil surrounding the peripheral of any element cylindrical gear, and then multiplied by the corresponding oil speed to acquire the churning loss on the periphery", + " Tribology International 151 (2020) 106536 \u03c4\u00f0r\u03b8\u00de dc \u00bc \u03bcoil \ufffd r \u2202 \u2202r \ufffdv\u00f0\u03b8\u00dedc r \ufffd\ufffd \u00bc 2\u03bcoil\u03c9r2 o r2 (4) Then, the resistance force of lubricating oil at any element cylinder gear periphery is the product of shear stress \u03c4\u00f0r\u03b8\u00dedc and wetted surface area dScp when the radius approaches the maximum (r \u00bc ro) (Seetharaman et al. [16,17]), namely: Fp \u00bc \u03c4\u00f0r\u03b8\u00de dc dScp (5a) and dScp represents the wetted surface element, that is as dScp \u00bc 2\u03c6rodb, where db is the gear width of the element gear and \u03c6 \u00bc cos 1\u00bd1 h\ufffd The dimensionless immersion depth is expressed (see Fig. 1) as h \u00bc h= ro, in which h represents the immersion depth of the bevel gear in the heel. Obviously, h \ufffd 0 indicates that the bevel gear is not immersed in oil at all, and h \ufffd 2 represents to the case where the gear is fully immersed. Here, the spiral bevel gear is partial immersed (see Fig. 1), and h is between 0 and 2. Then, the drag force can also be expressed as Fp \u00bc 2\u03c6rodb\u03c4\u00f0r\u03b8\u00de dc \u00bc 4\u03bcoil\u03c9rocos 1\u00bd1 h\ufffddb (5b) It must be pointed that the oil levels are regarded as static, in other words, the immersion depth is fixed. However, the rotational movement of the gear will be dynamically changing the oil levels (i.e. the immersion depth is dynamically changing). The higher speed the gear is, the more prominent the phenomenon is. This paper will pay attention to the dynamic immersion depth in Section 3", + " D3 and D4 are calculated based on the boundary conditions \u03c8 \u00bc 0 atr \u00bc rrj\u03b8j \ufffd \u03b8c orr \u00bc rij\u03b8j \ufffd \u03b8c (18) and d\u03c8 dr \u00bc \ufffd \u03c9rr at r \u00bc rr j\u03b8j \ufffd \u03b8c \u03c9ri \u03c9ri at r \u00bc ri j\u03b8j \ufffd \u03b8c (19) to the biharmonic equation for radial flow: \u03c8 \u00bcD1 \u00fe D2r2 \u00fe D3 log r \u00fe D4r2 log r (20) For a splash lubricated spiral bevel gear partially immersed in oil and partially surrounding by air, the windage power losses similar to the churning losses are composed of three individual parts Pw \u00bcPwp \u00fe Pwf \u00fe Pwts (21) where Pwp denotes the power loss caused by the air acting on pitch cone\u2019s flank of the spiral bevel gear; Pwf refers to the power loss in the toe/heel; Pwts represents the air drag power loss induced within the tooth space/cavity. With each loss determined, the total windage power losses Pw could be then obtained. Windage loss exerted by air on the periphery of a spinning spiral bevel gear partially surrounding by air, seen in Fig. 1, is determined as the product of drag force and tangential velocity. The final form of the windage loss on the periphery is found as Pp\u00bc Z B 0 4\u03bcair\u03c92\u00f0Ro b tan \u03b4a\u00de 2cos 1 \ufffd Ro h0 Ro b tan \u03b4a \ufffd db (22) Here, h0 \u00bc 2Ro h denotes the immersion depth in air. Under the laminar flow regime Re \u00bc 2\u03c1air\u03c9r2 o=\u03bcair < \u00f010\u00de5 and turbulent flow regime Re \u00bc 2\u03c1air\u03c9r2 o=\u03bcair > \u00f010\u00de6 s (see Fig. 2), combining with the boundary layer theory, the drag force and the corresponding windage loss exerted by air on the face of the bevel gear is defined as Fwf \u00bc 1 2 \u03c1airU 2C\u00f0L\u00der2 o h\u03c0 2 sin 1\u00f01 h 0 \u00de \u00f01 h 0 \u00de ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi h 0 \u00f02 h 0 \u00de q i (23) then: Pwf \u00bc 0:205\u03c1air\u03bc0:5 air \u03c92:5r4 o h \u03c0 2 sin 1\u00f01 h 0 \u00de \u00f01 h 0 \u00de ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi h 0 \u00f02 h 0 \u00de q i ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi sinfcos 1\u00bd1 h 0 \ufffdg q (24) and F\u00f0T\u00dewf \u00bc 1 2 \u03c1airU 2C\u00f0T\u00deScf \u00bc 1 2 \u03c1airU 2C\u00f0T\u00der2 o h\u03c0 2 sin 1\u00f01 h 0 \u00de \u00f01 h 0 \u00de ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi h 0 \u00f02 h 0 \u00de q i (25) then, P\u00f0T\u00dewf \u00bc 0:0125\u03c1air\u03bc0:14 air \u03c92:86r4:72 o h \u03c0 2 sin 1\u00f01 h 0 \u00de \u00f01 h 0 \u00de ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi h 0 \u00f02 h 0 \u00de q i fsin\u00bdcos 1\u00f01 h 0 \u00de\ufffdg 0:14 (26) Here, h 0 \u00bc 2 h represents the dimensionless air immersion parameter, h denotes the static oil immersion depth, seen in Fig. 1. Distinguishing from the churning loss, the windage loss occurring in the cavities is mainly caused by the vortices close to the tooth addendum rather than the eddies occurring near the tooth root. To assess this windage loss, an assumption that the air flow on the tooth surface Y. Dai et al. Tribology International 151 (2020) 106536 sidewalls is affected by the prior tooth in the rotation direction, had been raised by Akin et al. and Ville et al. [46,47] (see Fig. 4). This windage behavior was also found in Refs" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000769_978-3-662-43893-0-Figure7.9-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000769_978-3-662-43893-0-Figure7.9-1.png", + "caption": "Fig. 7.9 Principle of the single-flank test", + "texts": [ + " However, contact pattern assessment is left to the subjective impressions and experience of the particular tester. It is a disadvantage that a contact pattern can be documented only if a copy is made with a strip of adhesive foil or if the contact pattern is photographed. Results and interpretation may change if the contact pattern test is conducted with an infra-red camera (see Sect. 4.4.2.1). In this method, the pinion and wheel are meshed in their installed positions in the gearbox, only one tooth flank being in contact at any time. The measuring principle of the single flank test is shown in Fig. 7.9. The base of all analyses is the measured fluctuation in the transmission ratio between the pinion and the wheel, referred to as the transmission error or single flank composite deviation. High-resolution incremental angular encoders are connected to the pinion and wheel spindles. The pinion spindle drives the gear spindle to which a braking torque is applied. A control unit transforms the sinusoidal signals from the incremental angular 304 7 Quality Assurance encoders such that they can be analyzed by computer software" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000141_j.renene.2013.10.024-Figure11-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000141_j.renene.2013.10.024-Figure11-1.png", + "caption": "Fig. 11. a) Nacelle aligned with wave direction; b) Nacelle perpendicular to wave direction.", + "texts": [ + " The generator was constrained in all degrees of freedom at the centre where it was imagined to be attached to the shaft. For the first design analysis, bearing supports were assumed to be of very high stiffness. To understand the influence of bearing stiffness and shaft displacement, the basic theory and some preliminary results for a multibody model are presented. In all of the cases described in detail below, the nacelle axis was assumed to be parallel to the wave direction (i.e. along z-axis), refer to Fig. 11. 5.1. Magnetic field only The normal magnetic stress calculated using equation (7) was 0.244MPa. This force was assumed to act as a uniformly distributed a) Fig. 6. a) Flux density contour plots from FEMM; b) R pressure load over each of the 25 segments of the exterior and interior surfaces of the rotor and stator. This load is relatively constant subjecting the rotor and stator structures to constant tension and compression respectively. The self weight of the structure was modelled by defining the material density and acceleration due to gravity, 9", + " 10 shows the corresponding values for the nacelle motions that were applied. It may be observed that the wave induced motions are not large enough to have significant impact on the structure. This confirmed the stiffness of the structure against external load. It is however important to estimate the impact with different wave heights and wave direction as they can result in higher nacelle accelerations. For this reason, motion loads due to 9 m waves were considered to act along the x and z-axes (refer to Fig. 11). For the case where wave loads were assumed to act along the x-axis, any axial variation of eccentricity (non-uniform rotor eccentricity along the axial length of the rotor) was also examined. The deformations were extracted for every 100 mm of the axial length of the rotor and stator structures (refer to Fig. 12). It may be noted that that negligible variation occurs laterally and this confirmed the initial assumption of no axial variation in air-gap. Also, no major difference was noted between the deformations computed for 3m and 9mwave heights" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000769_978-3-662-43893-0-Figure3.14-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000769_978-3-662-43893-0-Figure3.14-1.png", + "caption": "Fig. 3.14 Contact pattern with contact lines, path of contact and transfer points", + "texts": [ + " A profile line is formed by the points with the smallest contact distance which can be determined for each contact position (see Fig. 4.33). Points on a potential contact line which fall below a certain contact distance (usually 3 to 6 \u03bcm, depending on the thickness of the marking compound) at one meshing position at least, constitute the effective contact line and are part of the contact pattern. Those points with the smallest contact distance on each contact line form what is termed the path of contact. Figure 3.14 shows a contact pattern in which the contact lines\u2014the major axes of the contact pattern\u2014are represented. In Fig. 3.14, the short lines perpendicular to the path of contact indicate the transfer points, or the points at which contact is transferred between neighboring tooth pairs. Parameters of the ease-off topography The ease-off allows us to describe the contact geometry very clearly, and also to determine gear set relevant variables such as the position and size of the contact pattern and the transmission error. For this reason, the ease-off topography offers specialists complete information on meshing conditions" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001478_j.acme.2019.06.005-Figure5-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001478_j.acme.2019.06.005-Figure5-1.png", + "caption": "Fig. 5 \u2013 Geometry of contacting bodies and the contact ellipse.", + "texts": [ + " It should also be noted that an increase in preload is accompanied by a decrease in the argument of the stiffness function (load value) for which the inflection point occurs [22,27], thereby increasing the range in which stiffness characteristic can be described by a linear model. In technological machines, linear guides are usually selected in such a way that they work in the range of working loads above the inflection point of the stiffness characteristic. The non-linearity of the stiffness characteristic of the model is caused by the model of contact between the raceways and the ball. This model can be considered as the contact between a deformable ball and cylinder, which is schematically shown in Fig. 5. This issue is described by Hertzian equations, allowing determination of the state of deformation and the state of stress in the areas of contact between two bodies. The contact surface area changes depending on the value of the pressing forces, which results in a change in the values and distributions of stresses as well as deformations and displacements in the contact area. After the deformation of the bodies caused by their mutual pressure (Fig. 5), in a general case the contact area formed is an ellipse with semi-axes a and b (a > b), which can be determined from the equations: a \u00bc a3 ffiffiffiffiffiffiffiffi P m n r ; b \u00bc b3 ffiffiffiffiffiffiffiffi P m n r (1) m \u00bc 4 \u00f01=r1\u00de \u00fe \u00f01=r01\u00de \u00fe \u00f01=r2\u00de \u00fe \u00f01=r02\u00de ; n \u00bc 8 3 E1E2 E2\u00f01 v1\u00de \u00fe E1\u00f01 v2\u00de (2) where a and b \u2013 coefficients depending on the value of the ratio B/A where: A \u00bc 2 m ; B \u00bc 1 2 ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 1 r1 1 r01 2 \u00fe 1 r2 1 r02 2 \u00fe 2 1 r1 1 r01 1 r2 1 r02 cos2\u2019 s (3) Based on the above formulas, assuming material data analogous to the model developed in Section 2 (E1 = E2 = 210,000 N/mm2, v1 \u00bc v2 \u00bc 0:3, r1 \u00bc r01 \u00bc 2:375 mm, r2 \u00bc 2:5 mm; r02 \u00bc 1) and the load P = 10 kN, the dimensions of the contact ellipse were: a = 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000510_robio.2015.7418948-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000510_robio.2015.7418948-Figure1-1.png", + "caption": "Figure 1. Scheme of quadrotor", + "texts": [ + " DYNAMIC OF THE QUADROTOR SYSTEM The quadrotor is an under-actuated rotorcraft, compose of two diagonal rotors turns clockwise whereas the second pair turn counter clockwise. The control of the motion is achieved by altering the rotation rate of one or more rotor. The dynamic model of the quadrotor includes the gyroscopic effects resulting from both the rigid body rotation in space 978-1-4673-9675-2/15/$31.00 \u00a9 2015 IEEE 1285 and the four propeller\u2019s rotation which generates the propeller forces and to move in space. We consider two frames, earth fixed frame or the inertial frame and a body fixed frame as shown in Figure 1.The center of mass and the body fixed frame origin are assumed to coincide. The matrix and ensure the transformation of the linear and the angular velocity from to . [ ] [ ] A typical dynamic model of the quadrotor is given by the following equations [13]: { \u0308 \u0307 \u0307 \u0307 \u0308 \u0307 \u0307 \u0307 \u0308 \u0307 \u0307 \u0308 \u0308 \u0308 and are the controls and is a disturbance { The thrust forces coming from the rotors, are , where is thrust factor and is the rotor speed. The rotors are driven by DC motors, where their dynamics is given by the equations below: { \u0307 { } The quadrotor mathematical model (3) can be expressed in a state space form as \u0307 where an expansion of the states has been taken into consideration, defined as [ \u0307 \u0307 \u0307 \u0307 \u0307 \u0307] " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000769_978-3-662-43893-0-Figure3.8-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000769_978-3-662-43893-0-Figure3.8-1.png", + "caption": "Fig. 3.8 Schematic view of the virtual machine when viewed in a direction perpendicular to the bevel gear and cradle axes. \u03b1 cradle roll axis, \u03b2 work piece axis, \u03b3 machine root angle, \u03b5 horizontal distance, \u03c7 sliding base, tB mounting distance, X crossing point of axes in", + "texts": [], + "surrounding_texts": [ + "A bevel gear machine is designed such that it imposes to the tool and bevel gear blank the relative motions of the generating gear meshing with the bevel gear. For this purpose one tooth of the generating gear is replaced by the straight flank profile of the tool. Figure 3.5 shows a tool used in the single indexing method. In order to roll the tool with the blank, i.e. to machine one tooth slot, and return to the starting position, the generating gear shown in the figure must rotate by a relatively small angle. On many mechanical machine tools, the cutter moves to and fro like a cradle, from the start of roll angle to the stop roll angle and back again. Thus, conceptually, a virtual gear-cutting machine closely resembles the former mechanical machine, but with the significant advantage that the virtual machine is not subject to any limitation in terms of penetration, stability, damping, assembly, accessibility etc. Modern 6-axis CNC machines may be fundamentally different from earlier pure mechanical ones or NC models, but their relative generating motion is identical. The tool is guided using three translations and three rotations, thus a total of six axes in relation to the work piece. machine root angle Figure 3.7 shows the virtual machine as seen along the axis of rotation of the cradle. For the historical reasons noted above, the term \u201ccradle axis\u201d is used in preference to virtual crown gear axis. The figure shows only one tooth of the generating gear, where a cup-type tool has been substituted. the gearbox" + ] + }, + { + "image_filename": "designv10_14_0000769_978-3-662-43893-0-Figure1.1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000769_978-3-662-43893-0-Figure1.1-1.png", + "caption": "Fig. 1.1 Principle of a rear axle drive", + "texts": [ + " Paul Bo\u0308ttcher improved this concept and, in 1910, presented a face mill cutter system which produced spiralshaped teeth on bevel gears [KRUM50]. \u00a9 Springer-Verlag Berlin Heidelberg 2016 J. Klingelnberg (ed.), Bevel Gear, DOI 10.1007/978-3-662-43893-0_1 1 Bevel gears first acquired substantial significance with the developing car industry at the beginning of the twentieth century. At that time, rear axle transmissions, including differential gears, was the normal vehicle power train concept in motor vehicles. Figure 1.1 shows a typical rear axle drive, with the gear train shown in diagrammatic form. The smaller bevel gear, driven by the Cardan shaft, is called the \u201cpinion\u201d while the larger bevel gear driven by the pinion is called the \u201cwheel\u201d. The wheel driving the axle is linked to a carrier with four differential gears. The two gears with horizontal axes each drive a tire. When one of the tires slows down, the other tire rotates faster and there is no slippage with the surface of the road. The bevel gears in the differential are required to compensate the speed difference between the two tires, whereas the pinion and the wheel are responsible for the transmission of engine power to the tires" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001953_s12206-020-2206-9-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001953_s12206-020-2206-9-Figure2-1.png", + "caption": "Fig. 2. Temperature field distribution of the bearing.", + "texts": [], + "surrounding_texts": [ + "Xu [33] conducted life tests on 44 sets of SKF Explorer 6205 deep groove ball bearings under the condition of radial load of 4500 N and rotational speed of 6000 rpm. In this paper, the experimental data in Ref. [33] is used for research. The dimensional parameters and material parameters of bearing are given in Table 1, the thermophysical properties of lubricating oil that been applied in the bearing are given in Table 2, and the distri- bution information of impact parameters are given in Table 3. The heat and convection heat transfer coefficients of the bearing can be calculated by Eqs. (7)-(11). The results are applied to the bearings as thermal loads and thermal boundary conditions. Through the thermal analysis of Workbench, the temperature distribution of the bearing can be obtained, as shown in Figs. 2 and 3. It can be seen that the highest temperature region of the bearing appears at the contact point between the rolling element and the inner ring raceway, followed by the second point which located between the rolling element and the outer ring raceway. Sampling the impact parameters and then using Monte Carlo method and distribution fitting to obtain the distribution of Iiw, \u2206Gi, \u2206GT, \u2206Gci and \u2206Gw. The results show the effective inter- ference follows iwl N\u223c (24.21, 12), the variation of clearance caused by effective interference follows iG N\u0394 \u223c (15.43, 0.912), the variation of clearance caused by temperature follows TG\u0394 \u223c N(4.03, 1.062), the variation of clearance caused by centrifugal force follows ciG\u0394 \u223c N(0.37, 0.052), and the variation of working clearance follows wG\u0394 \u223c N(19.84, 1.42). The simulation results are shown in Figs. 4-8, respectively. The load distribution of rolling elements under different working clearances is shown in Fig. 9. Fig. 10 shows the variation of maximum element load under different working clearances at the radial load of 4500 N, and the variation of life factor under different working clearances as shown in Fig. 11. Obviously, the working clearance has significant influence on the life of the bearing. Specifically, when Gw < -30.5 \u03bcm and Gw > 0 \u03bcm, LF is less than 1, and when Gw is in the interval [-30.5, 0] \u03bcm, LF is greater than 1. Therefore, the optimal working clearance interval is selected as [-30.5, 0] \u03bcm. 19 \u03bcm in group N is selected as the original clearance (G0), then, the distribution of the working clearance is wG \u223c N(-0.84, 1.42). Based on reliability analysis model of rolling bearing, the reliability is computed to be R = 0.7257. Under the current conditions, the rated life of the bearing can be calculated by the L-P model. By analyzing the life test data [33], it can be known that 11 of the 44 sets of bearings have not reached the rated life. Then, the experimental reliability is 0.75, and the error of the proposed method and experiment is 3.24 %, which verifies the correctness of the proposed method. Fig. 12 shows the reliability under different original clearances. It can be clearly found that when G0 is selected in group 2 and the first half of group N, the reliability R > 0.99, and G0 in this interval, the normal operation of the rolling bearings can be guaranteed." + ] + }, + { + "image_filename": "designv10_14_0001111_j.jfranklin.2018.04.035-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001111_j.jfranklin.2018.04.035-Figure3-1.png", + "caption": "Fig. 3. Illustration of the HSEE-AHV.", + "texts": [ + ", Control of a time-varying hypersonic vehicle model subject to inlet un-start condition, Journal of the Franklin Institute (2018), https://doi.org/10.1016/j.jfranklin.2018.04.035 24 H. An et al. / Journal of the Franklin Institute 000 (2018) 1\u201334 To verify our proposed controller, simulations are implemented in the High-fidelity Simulation and Evaluation Environment for Air-breathing Hypersonic Vehicles (HSEE-AHV) developed by Space Control and Inertial Technology Research Center, Harbin Institute of Technology. As shown in Fig. 3 , this simulation system is mainly composed of a flight dynamic computer, a GNC computer, and a 3-DOF Stewart attitude simulator, which communicate with each other via fibre-optical reflective memory. The high-fidelity model is established based on the wind-tunnel data and computational fluid dynamics simulations for a scaled AHV model at different working points, while the atmospheric environment is calculated according to the Earth Global Reference Atmospheric Model. Therefore, parameters in the simulated AHV dynamics are uncertain and time-varying" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001298_s11012-019-01088-y-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001298_s11012-019-01088-y-Figure1-1.png", + "caption": "Fig. 1 Schematic diagram of lumped parameter dynamic model", + "texts": [ + "2, friction excitation with time-varying coefficient is presented. The calculation of TVMS including effect of friction is briefly introduced in Sect. 2.3. The simulation flowchart of this work is given in Sect. 2.4. The results and discussions are given in Sect. 3. Finally, conclusions are arrived. 2.1 Governing equations of helical gear pair Taking flexibility of both ends of shaft supported by bearings into account, an 8-DOF model is established and the lumped parameter dynamics model is shown in Fig. 1. Two local coordinates of system O1-XYZ in pinion, O2-XYZ in gear are established, where origins are located at mid-point of tooth width, X-axis parallel to the plane of action (PoA) and Y-axis normal to PoA. Ff in the figure indicates friction force. bb represents helix angle of basis circle. The driving torque T1 is applied to pinion and load torqueT2 on driven gear.Tfi (i = 1, 2) represents friction torque applied to both gears. Angular displacement is denoted by hi (i = 1, 2) and the linear displacements along three axis by x, y and z, respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001161_tia.2018.2874350-Figure5-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001161_tia.2018.2874350-Figure5-1.png", + "caption": "Fig. 5. Cross sections of machine without and with auxiliary slots. (a) Without auxiliary slots. (b) With auxiliary slots.", + "texts": [ + " The rated working condition indicates the amplitude of current equals to its rated value, and the current angle is zero, which means the fundamental component of the PM flux linkage is aligned with the d-axis. The optimized parameters are height which may be known as depth, width as well as position of auxiliary slots. The shift angle is defined as the mechanical angle between the middle of the auxiliary slots as well as the middle of corresponding tooth in clockwise direction. The comparison of topologies of machines with and without auxiliary slots are shown in Fig.5, and the detailed parameters of the optimal auxiliary slots are listed in Table II TABLE II PARAMETERS OF AUXILIARY SLOT Height (mm) 0.549 Width (mm) 9.540 Shift angle (Mech. Deg.) 16.384 0093-9994 (c) 2018 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. > REPLACE THIS LINE WITH YOUR PAPER IDENTIFICATION NUMBER (DOUBLE-CLICK HERE TO EDIT) < 4 The magnet eddy current loss is compared in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001532_j.wear.2019.01.053-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001532_j.wear.2019.01.053-Figure3-1.png", + "caption": "Fig. 3. The influence of the applied tilt angle for pinion manufacture on the pressure distributions on the two tooth pairs instantaneously in contact.", + "texts": [ + " The number of tooth pairs instantaneously in contact depends on the instantaneous rotational position of the gear pair and on the transmitted torque (load). The geometrical separation of tooth surfaces depends on the manufactured tooth surface geometry and on instantaneous rotational position of the gear pair. The influence of the tilt angle ( ) in the manufacture procedure on the geometrical separation of contacting tooth surfaces is shown in Fig. 2 for the rotational position of the gear pair when the instantaneous contact point is in the middle point of the gear tooth. Fig. 3 shows the pressure distributions for the two tooth pairs instantaneously in contact when the pinion teeth are manufactured by different values of the tilt angle. It can be seen that the tilt angle has a considerable influence on the geometrical separations of the instantaneously contacting tooth surfaces and on the pressure distributions, too. The influence of the machine tool settings for pinion manufacture on the maximum pressure, temperature and average temperature in the contact region and on the efficiency of the gear pair is shown in Figs" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001925_978-3-030-46817-0_30-Figure21-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001925_978-3-030-46817-0_30-Figure21-1.png", + "caption": "Fig. 21. FEM model - strain distribution. Fig. 22. FEM model - warp angle.", + "texts": [ + " The experimental tensile test of tie bar from housing of tie rod assembly is performed by control device MR 96 (see Table 8). The deformation achieved as result of finite element analysis is similar to the results of the tensile test performed by control device MR 96. The maximum appeared displacement (0,455 < allowed value 0, 5 mm, see Fig. 4) and maximum stress value is lower than ultimate stress allowed by documentation request (437,8 < 1100 MPa, see Fig. 5). Also, analysis of strain distribution (see Fig. 21) and warp angle (41,4 < allowed value 44\u00b0, see Fig. 22) through the whole tensile test showed to us lower angle than allowed value required by design documentation. Figs. 23, 24, 25 and 26. 142 I. Alagi\u0107 In case of ball joint, the results of stress distribution and allowed displacement are presented in Figs. 27, 28, 29, 30, 31 and 32. Finite Element Analysis (FEA) of Automotive Parts Design as Important Issue 143 The displacement achieved as a result of finite element analysis doesn\u2019t much differ from results of the laboratory-test performed by control device MR 96" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000769_978-3-662-43893-0-Figure4.30-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000769_978-3-662-43893-0-Figure4.30-1.png", + "caption": "Fig. 4.30 Tooth contact analysis using FEM on a hypoid gear set (left: pinion; right: wheel) in a rear axle gear unit [ZFAG]", + "texts": [ + " For a complete root and contact stress analysis using FEM, the FE model must be extended to include the entire surroundings of the gear, and a very fine mesh must be generated in the tooth zone, at least at points where high notch stresses are expected (tooth root zone). With regard to the applied load, all meshing positions must be taken into account. Despite high computation speeds and the ability to process large data volumes, calculations of this sort are currently performed only for specialized individual tasks and for comparative tests, owing to the high processing effort involved. As an example, Fig. 4.30 shows the results of a 3D FE simulation for a hypoid gear set in a rear axle gear unit. The aim of this simulation is to analyze the stress distribution in the tooth root during the contact. In Fig. 4.30, the resulting lines of contact are clearly visible. The available mesh fineness does not allow evaluation in terms of contact stress on the flank. The calculation was performed for 10 meshing positions (increments of 0.1 pe); results are shown for one contact position on the pinion and the wheel. Calculation method based on influence numbers A stress analysis based on influence numbers consists of two main parts: calculation of the load distribution (solving the contact problem) and calculation of the stresses resulting from this load distribution" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001646_s11771-019-4207-3-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001646_s11771-019-4207-3-Figure1-1.png", + "caption": "Figure 1 Spherical hob forming principle", + "texts": [ + " According to the assembly error of the spherical hob, the influences of each error on the face gear profile are analyzed. Finally, an assembly spherical hob is manufactured and the gear hobbing process experiment is completed. The feasibility of the assemble spherical hob machining face gear is verified. J. Cent. South Univ. (2019) 26: 2704\u22122716 2706 It is similar to the evolution of gear hob from rack to ordinary Archimedes gear hob, the evolution of cylindrical gear to spherical hob is analyzed, which is shown in Figure 1. It is envisaged that the cross section of the basic worm of the spherical hob can always correspond to the position of the cylindrical gear teeth when the cylindrical gears rotate around its own axis, so that the spherical hob rotates on the cross-section equivalent to the rotation of the cylindrical gear. The rotational ratio of the cylindrical gear to the spherical hob basic worm is Nw/N1, where Nw is the head count of the spherical hob basic worm, generally take Nw=1. It means that the spherical hob basic worm turns a circle, the cylindrical gear just turns a tooth" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002352_lra.2021.3056066-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002352_lra.2021.3056066-Figure2-1.png", + "caption": "Fig. 2. (a) Overview of experimental setup. The robotic arm holds the SPA and moves it randomly along an axis (red arrows) to introduce external contact at different location along the body. (b) Snapshots of the SPA during data collection. A high number of reflective markers are used to accurately describe the configuration, because the full-body shape under contact can be very complex, which cannot be sufficiently captured by either using only a tip marker or assuming uniform bending.", + "texts": [ + " To investigate the full-body dynamics of the SPA, we attach eight reflective markers (plus one at the very top as reference) along its constraint layer which are continuously tracked by our motion capture system. We intend to predict (i) the relative x- and y-coordinates of all eight markers and (ii) the magnitude and location of external contact force, using only limited sensory feedback. We measure the contact force magnitude with a compression load cell placed in front of the SPA. Using a robotic arm, we move the SPA randomly along the y-axis, while keeping the position of the load cell (ycell) fixed, and thus the force location can be easily obtained as ycell \u2212 yref. Fig. 2 shows the experimental setup of the robot for data collection. In addition to flex sensor feedback, we also measure the pressure supplied to the SPA and use this information as additional input to our model. The pressure sensor used for measurement is external to the SPA and does not affect its dynamics. We actuate the SPA with a series of randomly determined pressure levels within a predetermined pressure range. Between every change in the pressure level, we introduce a random delay of 1 to 5 seconds to capture both the transient and steady state response of the SPA" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002147_tcst.2021.3069106-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002147_tcst.2021.3069106-Figure1-1.png", + "caption": "Fig. 1. Model diagram of the helicopter.", + "texts": [ + " This brief is organized as follows. In Section II, a helicopter model and corresponding state-space equations are presented. In Section III, we design the helicopter controller in continuous time, and the stability analysis is provided. In Section IV, the discretization of the proposed control design, a hardware setup, and experimental results are shown. Finally, concluding remarks are discussed in Section V. The standard notation will be used in this brief. The coordinate system used throughout this brief is shown in Fig. 1\u2014north-east-down (NED) coordinates. Following the generalized control design process, we utilize the helicopter model in [17], which has rotational and translational dynamics. The helicopter model for the rotational dynamics is \u0307 = A + B FR + B\u03c3 , Y = C (1) where = [\u03c61, \u03c62, \u03b81, \u03b82, \u03c81, \u03c82]T = [\u03c6, \u03c6\u0307, \u03b8, \u03b8\u0307 , \u03c8, \u03c8\u0307 ]T \u2208 R 6, where the angles \u03c6, \u03b8, and \u03c8 are along xb, yb, and zb, respectively, and FR = [ f\u03c6, f\u03b8 , f\u03c8 ]T = FR( , ur , ut ). The uncertainty is \u03c3 = [\u03c3\u03c6, \u03c3\u03b8 , \u03c3\u03c8 ]T \u2208 R 3 in the form of angular acceleration" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000715_j.oceaneng.2018.06.068-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000715_j.oceaneng.2018.06.068-Figure1-1.png", + "caption": "Fig. 1. Coordinate system of AUV.", + "texts": [ + " In Section 5, a set of experimental comparisons are carried out in terms of accuracy and speed by simulations and experiments, while the results are also being discussed. Finally, Section 6 summarizes the key conclusions of this work. The notation, term and coordinate system are defined by SNAME (Society of Naval Architects and Marine Engineers, 1950). The worldfixed coordinate system (E-\u03be\u03b7\u03b6) has its origin E fixed to the earth. The body-fixed coordinate system (O-xyz) with origin O is moving reference frame that is fixed to AUV (see Fig. 1). For AUV, the six different motion components are conveniently defined as surge (u), sway (v), heave (w), roll (p), pitch (q) and roll (r). v [u, v, w, p, q, r]T= is the linear angular velocity with respect to the body-fixed reference frame, and \u03c4 [X, Y, Z, K, M, N]T= is the total forces and moments acting on AUV. The Euler angles are heading angle \u03c8, pitch angle \u03b8 and roll angle \u03c6. The notations are summarized in Table 1. In general, the force on AUV can be divided into two categories: the one is the hydrodynamic force Fvis (or the resistance force) due to a vehicle moving in fluid, the other is the external force Felse, such as the rudder force, propulsion force, gravity and buoyancy, etc" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000808_j.autcon.2014.08.002-Figure4-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000808_j.autcon.2014.08.002-Figure4-1.png", + "caption": "Fig. 4. Components and main two motions of an excavator.", + "texts": [ + " \u04e8T is the directive angle between the manipulator of the excavator and the dump truck, and it may influence the loading time. However, this study focused on the Hypotheses 1 and 2, and the analysis on \u04e8T is recommended for future study. Based on the hypotheses, analysis of the excavation and loadingwas performed for proper data collection. First excavation is performedwith movement of four joints. It rotates around the center joint horizontally, and the boom, arm, and bucket dig or throw the soil vertically (Fig. 4). Both rotation and vertical movement of the boom, arm, and bucket can be performed simultaneously. Asmentioned above, loading basically consists of dumping, rotating for scooping, scooping, and rotating for dumping. However, since the operator rotates the center joint horizontally while moving the boom, arm, and bucket vertically, the rotating part was again divided into horizontal rotation and vertical movement. Table 2 breaks down the loading operation into detailed motion in this thesis. Also, the operator performs arbitrary gathering during loading to facilitate scooping, and performswork location movement for efficient loading" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000769_978-3-662-43893-0-Figure2.21-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000769_978-3-662-43893-0-Figure2.21-1.png", + "caption": "Fig. 2.21 Pinion face width, inner and outer diameter [ISO23509]", + "texts": [ + "79) Modified slot width \u03a3\u03b8fM \u00bc \u03a3\u03b8f C or \u03a3\u03b8fM \u00bc 1, 3\u03a3\u03b8f S whichever is smallest Table 2.13 Addendum angle, \u03b8 a2 and dedendum angle, \u03b8 f2, for the wheel Depth wise taper Angle ( ) No. Standard depth \u03b8a2 \u00bc arctan hfm1 Rm2 \u03b8f2 \u00bc \u03a3\u03b8fS \u03b8a2 (2.80) (2.81) Uniform depth \u03b8a2 \u00bc \u03b8f2 \u00bc 0 (2.82) Constant slot width (Duplex) \u03b8a2 \u00bc \u03a3\u03b8fC ham2 hmw \u03b8f2 \u00bc \u03a3\u03b8fC \u03b8a2 (2.83) (2.84) Modified slot width \u03b8a2 \u00bc \u03a3\u03b8fM ham2 hmw \u03b8f2 \u00bc \u03a3\u03b8fM \u03b8a2 (2.85) (2.86) 42 2 Fundamentals of Bevel Gears Determination of the pinion tooth face width (see Table 2.15 and Fig. 2.21) Whereas calculation of the pinion face width is trivial for non-offset bevel gears, its value being generally equal to that of the wheel, for hypoid gears it is a function of the offset and can be calculated using three different methods (A to C) as presented below. 2.3 Bevel Gear Geometry Calculation 43 44 2 Fundamentals of Bevel Gears Determination of inner and outer spiral angles (see Table 2.16) The spiral angle changes continuously along the face width on spiral bevel gears. Accordingly, the inner and outer spiral angle values are calculated from those at the mean point" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000769_978-3-662-43893-0-Figure1.6-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000769_978-3-662-43893-0-Figure1.6-1.png", + "caption": "Fig. 1.6 Principle of a helicopter drive", + "texts": [ + " Because of the high rotation speeds and relatively low torques involved, these bevel gears are subject to quite different demands than those used for example in automotive axle gear drives. 1.3 Aircraft Engines 5 Similarly to airplanes, helicopter engines are usually gas turbines and deliver the power used to drive the main and tail rotors while the residual exhaust jet supplements propulsion during flight. Because the shaft of gas turbines is invariably mounted horizontally, an angular gear train is needed to power the main rotor. The counter torque generated by the main rotor around the vertical axis of the helicopter is compensated by means of a tail rotor (Fig. 1.6). The rotational speed of the main rotor is always chosen such that blade tip velocity at maximum forward flight speed is subsonic. Depending on the diameter of the rotor, this results in rotor speeds below 500 RPM. Typical power turbine shaft speeds being above 8,000 RPM, and thus significantly higher than that of the rotor, a reduction gearbox with a large transmission ratio is required. Planetary gears are ideal for such applications: the bevel gears are placed ahead of the input stage to the planetary gear train such that the gear designer is dealing with higher rotational speeds rather than higher torques. Other bevel gear sets are necessary to drive the tail rotor. If a continuous shaft runs from the main gear unit to the tail rotor, one bevel gear set is required such that the longitudinal shaft can drive the tail rotor mounted at right angle. When it is not possible to install a continuous shaft (Fig. 1.6), a number of shaft sections are required. Further bevel gear sets are then needed to transmit rotation to each of the sections of the tail rotor drive. Apart from rapidly rotating bevel gears in aircraft engines, another application is the actuation of wing flaps. In addition to ailerons controlling aircraft roll, the wings contain flaps to alter airfoil shape at take-off and landing. Flaps are at the trailing edge of the wing profile and as they extend aft out of the wing, their angle of attack increases" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000898_s12206-015-0417-2-Figure4-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000898_s12206-015-0417-2-Figure4-1.png", + "caption": "Fig. 4. Configuration of test rig components.", + "texts": [ + " We have experimentally evaluated the performance and fatigue life of a pitch bearing for 3.0-MW-class wind turbine. The results of this study may be used as a reference data to verify pitch bearings for multi-megawatt wind turbines. For the purpose of this study, a test rig developed for testing pitch and yaw bearings for wind turbines was employed. The test rig comprises seven main components, including a top frame main assembly and a top frame side assembly, which are the parts where loads are applied (Fig. 4). Several hydraulic cylinders were used as the loading components. From the bearing\u2019s center point, hydraulic cylinders are located at four vertical and two horizontal positions to apply six degrees of freedom (DOF) loads (Fig. 5). By using the force balancing condition, it is possible to determine the hydraulic cylinder forces\u2014A1, A2, A3, A4, A5, and A6\u2014to reproduce each external load\u2014Fx, Fy, Fz, Mx, and My\u2014as shown in Eq. (1). These forces were specified to compensate the self-weight of the test rig parts on the bearing" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001613_j.engfailanal.2019.06.084-Figure21-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001613_j.engfailanal.2019.06.084-Figure21-1.png", + "caption": "Fig. 21. Center support for inner synchronizer ring", + "texts": [ + " The new synchronizer, the highway durability vehicle failed synchronizer, and the TV test rig synchronizer have the same carbon pattern (Fig. 20b). In the dual cone synchronizer ring, the outer cone is guided by the sleeve and or hub with a nominal radial clearance of 0.5 mm. The inner cone has a radial clearance of 3 mm with the surrounding parts such as hub or gear. Due to this the inner cone can move freely in the radial direction. On the high way durability test, the wear of the inner carbon liner contributes to 85% of total synchronizer wear. The inner cone radial clearance with the hub is reduced to 0.5 mm as shown in Fig. 21. This reduces the free movement of the inner synchronizer ring and gives good guidance to the inner synchronizer ring. On the TV test rig for 65 h, the gearbox with 0.5 mm radial clearance had a wear of 0.98 mm compared with 1.4 mm wear of 3 mm radial clearance synchronizer ring (Fig. 22). Due to center support the wear of the inner and the outer carbon liners became 60:40 ratio. The reduced clearances reduced the radial movement of the synchronizer rings and reduced the intensity of the hitting of the synchronizer ring and in turn reduced the wear in all gear synchronizer positions" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002015_j.asr.2020.09.040-Figure7-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002015_j.asr.2020.09.040-Figure7-1.png", + "caption": "Fig. 7. Physical significance of the quantities S and b. Whereas the variable S denotes the length of the zigzag within the same plane of a membrane when packed in cylindrical form, the variable b denotes the height of the membrane in packed cylindrical configuration.", + "texts": [ + " (5)), thickness t, boundaries R0;Rf , and initial conditions h0, distance S0 and r0 \u00bc R0. Here, R0 denotes the circumradius of the polygon at the core of the flat membrane. Conversely, Rf denotes the circumradius of the polygon at the outer boundary of the flat membrane, thus it represents the user-defined upper bound on the value of R. From a physical point of view, the variable S denotes the length of the segment of the planar membrane when it is wrapped diagonally in a region between the two bases of the surface of the cylindrical shell, as shown by Fig. 7, and thus the variable b determines the height of the membrane in packed cylindrical configuration, as shown by Fig. 7. From the solution of the system of differential equations in Eq. (14) to Eq. (16), one can find the values of S, and b from Eq. (2). An example of the solution of variables S and b is portrayed by Fig. 8. Here, the initial values of S are large and converge towards 25.06 mm, whereas the initial value of b was 25 mm and decreased in minute proportions to 24.75 mm. The large values of S at the beginning are due to the inner layers of the membrane being wrapped in longer diagonal segments of the cylinder" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000769_978-3-662-43893-0-Figure1.4-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000769_978-3-662-43893-0-Figure1.4-1.png", + "caption": "Fig. 1.4 Principle of a four-wheel drive with a longitudinal engine", + "texts": [ + " The two axle drives are linked by means of a center differential, compensating differences in the rotation speed of the front and rear axles. A common feature in all transverse engine concepts is that they place the power unit ahead of the front axle. This brings advantages in terms of space which is partly outweighed by the greater weight on the front axle. In upmarket cars, a longitudinal engine concept with rear-wheel or four-wheel drive prevails. The superior weight distribution has a positive effect on driving dynamics and safety. Figure 1.4 shows the concept which consists of an engine mounted longitudinally to the driving direction, with front- and rear-wheel drive. The engine lies over the front axle and the front axle gear unit is beside or below the engine. This concept has advantages for bodywork design in terms of accident protection for passengers, as the heavy engine is mounted further back in the engine compartment leaving more space for an energy-absorbing design in the front of the vehicle. Although the volume of bevel gears used worldwide in the automotive sector is the largest, bevel gears also play an essential role in aircraft applications" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001308_icems.2019.8922504-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001308_icems.2019.8922504-Figure1-1.png", + "caption": "Fig. 1. Motor model TABLE \u2160", + "texts": [ + " The output torque ability of the motor is greatly improved by improving the stator tooth structure of the motor and embedding permanent magnet ring in the stator. The results show that the output capability of the motor can be improved by using the advantages of SMC materials and improving the structure of the motor. Ferrite motors can also have the same torque density as NdFeB motors[8]. In this paper, a high speed BLDC PM motor was proposed. Comparisons of iron loss and temperature characteristics of motors using SMC and lamination stators respectively. The motor model is shown as Fig. 1, and the main parameters of the motor are shown in Table \u2160. The equation of iron loss of SMC material and laminated material were obtained by fitting the loss experimental data. The iron loss of SMC motor and lamination motor were calculated by magnetic density simulation of different parts in the motor, and the results were verified by finite element analysis. Then the temperature rise of the motor is analyzed by calculating the iron loss of the motor. Finally, the volume of SMC motor is optimized" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002129_j.jallcom.2021.159609-Figure8-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002129_j.jallcom.2021.159609-Figure8-1.png", + "caption": "Fig. 8. The internal crack formation mechanism of the sample: (a) and (b) is the scanning direction; (c) is the evolution of red-hot state with time after laser step acting on the sample; (d) is the deposition direction. vs is the laser scanning speed.", + "texts": [ + " In the deposition direction, the cracks distributed in the sample top (Fig. 7a) sprouted at the shrinkage cavity and expanded to around. The cracks distributed in the sample middle and bottom were primarily longitudinal cracks, mostly sprouting from the pores and propagating along deposition direction (Fig. 7b and c). The crack surface was distributed from the sample center to the surroundings, as shown in Fig. 7(d). The formation and distribution of cracks in the sample have also been reported in the research of laser processing other ceramic materials [41]. Fig. 8 shows the cracking mechanism of the cylindrical samples. The initiation and propagation of cracks were closely related to the thermal stress generated during the deposition process, residual stress, and defects at the end of the deposition process. When the laser beam was irradiated to one side of the sample cross-section during DLD shaping the cylindrical sample (the first stage, Fig. 8a), a molten pool was formed by melting newly entered powder remelting the next deposition layer (zone 1). The temperature of zone 1 and the solidified region (zone 2) was higher than that of other solidified regions (zone 3). Therefore, the melt in zone 1 produced a pressing force on zone 2 due to thermal expansion; zone 2 was subjected to compressive stress ( 1 2). This compressive stress invariably existed during the shaping process of each deposition layer. As the laser beam left the position (the second stage, Fig. 8b), this part of the molten pool (zone 1) was solidified into zone 2. However, the temperature of zone 2 was still higher than that of zone 3, and there was a temperature difference \u2206T1. In the subsequent cooling and shrinking process, zone 2 failed to shrink freely and was limited by zone 3, forming tensile stress of zone 3 to zone 2 ( 3 2). Therefore, during the deposition process, the cylindrical sample tended to crack from center to edge (Fig. 7c). In addition, compared with Al2O3, the elastic modulus of mullite is about 223", + " According to the previous discussion, when the z-increment was 0.2 mm, more energy was entered into the melting pool, leading to higher molten pool temperature than other process conditions. The thermal conductivity of mullite (6\u20133.4 W m-1 K-1 (100\u20131200 \u2103)) is much lower than that of Al2O3 (26\u20134 W m-1 K-1 (20\u20131000 \u2103)) [42]. Therefore, the heat input to the molten pool was difficult to completely dissipate in a short time after the laser step irradiating the sample, resulting in uneven temperature distribution in the sample (Fig. 8c). Constraint stress was formed inside the sample due to displacement restrictions on the solidified layer or substrate. During the sample cooling (considering the stress distribution in the laser scanning direction), the sample bottom was subjected to tensile stress due to substrate limitations when it shrank. In the sample height direction, tensile stress decreased, and the sample top may even be transformed into compressive stress [43], as shown in Fig. 8(d). As a result, a bending moment was generated [44], which further expanded the longitudinal cracks that have already formed. With the increase of z-increment, the effect of heat accumulation decreased, the thermal stress decreased, and the cracking phenomenon slowed down. Fig. 9 shows the surface distribution of elements in the sample longitudinal-section using EPMA. The sampled contained the mullite crystal phase and the glass phase. The mullite crystals were rich in aluminum and poor in silicon" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001304_j.mechmachtheory.2019.103697-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001304_j.mechmachtheory.2019.103697-Figure1-1.png", + "caption": "Fig. 1. Schematic of the three-three orthogonal parallel mechanism.", + "texts": [ + " According to the design requirements, a set of optimal parameters are selected to manufacture prototype and verified by the finite element and experiment. 2. Configuration of six-axis force/torque sensor with three-three orthogonal parallel mechanism In the previous research of our laboratory, a new type of 6-DOF parallel mechanism with an orthogonal \u20183\u20133 \u2032 -PSS configuration has been proposed [27 , 28] . Based on the above configuration, we proposed a three-three orthogonal parallel configuration for six-axis force/torque sensor design. Schematic of the three-three orthogonal parallel configuration is shown in Fig. 1 . The configuration is composed of a measuring platform, a lower fixed platform, an upper fixed platform and six SPS branch chains connecting the measuring platform and the fixed platform. A i ( i =1 , 2 , 3 ) are spherical joints fixed on the lower fixed platform and A i ( i =4 , 5 , 6 ) are the spherical joints fixed on the upper fixed platform, respectively. a i ( i =1 , 2 ... 6 ) are spherical joints fixed on the measuring platform. The radius of the circle on the lower fixed platform is r 1 and the radius of the circle on the upper fixed platform is r 2 , respectively", + " (3) can be rewritten as the following form through exchanging the positions of F z and M z , \u23a1 \u23a2 \u23a2 \u23a2 \u23a2 \u23a3 F x F y M z M x M y F z \u23a4 \u23a5 \u23a5 \u23a5 \u23a5 \u23a6 = \u23a1 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a3 0 0 0 r 3 cos ( \u03b2) \u2212r 2 cos ( \u03b1) \u221a r 2 2 \u2212r 2 3 r 2 sin ( \u03b1+ \u03c06 ) \u2212r 3 sin ( \u03b2+ \u03c06 ) \u221a r 2 2 \u2212r 2 3 r 2 sin ( \u03c06 \u2212\u03b1) \u2212r 3 sin ( \u03c06 \u2212\u03b2) \u221a r 2 2 \u2212r 2 3 0 0 0 r 3 sin ( \u03b2) \u2212r 2 sin ( \u03b1) \u221a r 2 2 \u2212r 2 3 r 3 cos ( \u03b2+ \u03c06 ) \u2212r 2 cos ( \u03b1+ \u03c06 ) \u221a r 2 2 \u2212r 2 3 r 2 cos ( \u03c06 \u2212\u03b1) \u2212r 3 cos ( \u03c06 \u2212\u03b2) \u221a r 2 2 \u2212r 2 3 0 0 0 \u2212r 2 r 3 sin ( \u03b1\u2212\u03b2) \u221a r 2 2 \u2212r 2 3 \u2212r 2 r 3 sin ( \u03b1\u2212\u03b2) \u221a r 2 2 \u2212r 2 3 \u2212r 2 r 3 sin ( \u03b1\u2212\u03b2) \u221a r 2 2 \u2212r 2 3 0 \u221a 3 r 1 2 \u221a 3 r 1 \u22122 0 0 0 \u2212r 1 r 1 2 r 1 2 0 0 0 1 1 1 0 0 0 \u23a4 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a6 \u23a1 \u23a2 \u23a2 \u23a2 \u23a2 \u23a3 f 1 f 2 f 3 f 4 f 5 f 6 \u23a4 \u23a5 \u23a5 \u23a5 \u23a5 \u23a6 (6) From Eq. (6) , we can find that the force mapping matrix is partial decoupling. The forces F x , F y and torque M z are only related to the axial forces f 4 , f 5 , f 6 . The force F z and torques M x , M y are only related to the axial forces f 1 , f 2 , f 3 . As shown in Fig. 1 , f 1 , f 2 , f 3 are the axial forces along the vertical branch chains A i a i ( i =1 , 2 , 3 ) and f 4 , f 5 , f 6 are the axial forces along the horizontal branch chains A i a i ( i =4 , 5 , 6 ) . The characteristic of partial decoupling makes the configuration more suitable for manufacturing sensor. Based on the schematic of the three-three orthogonal parallel configuration, a configuration of three-three orthogonal sensor is designed as shown in Fig. 2 . The measuring platform and the fixed platform of the sensor are connected by six force-sensitive elements with rectangular section. The force-sensitive element is equivalent to the SPS branch chain that as shown in Fig. 1 . The sensor uses flexible spherical joints instead of the ideal spherical joints. A O i ( i =1 , 2 ... 6 ) denote the spherical joints on the fixed platform. a o i ( i =1 , 2 ... 6 ) denote the spherical joints on the measuring platform. O \u2212 XY Z is the fixed coordinate system and located the lower surface center of the fixed platform. o \u2212 xyz is the moving coordinate system and located above the surface center of the measuring platform. The two coordinate systems have the same posture. The centers of flexible spherical A 1 and a 1 are located in plane XOZ" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000769_978-3-662-43893-0-Figure2.16-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000769_978-3-662-43893-0-Figure2.16-1.png", + "caption": "Fig. 2.16 Definition of the hand of spiral. 1 right hand, 2 left hand, 3 viewed from the pitch cone apex", + "texts": [ + " The first is referred to as face milling (FM), the second as face hobbing (FH). In order to define the hand of spiral of a bevel gear, one looks from the apex of the pitch cone toward a tooth which is at the 12 o\u2019clock position. If the tooth, seen from front to back, curves to the right, the gear has a right hand of spiral, and vice versa. The concave flank is normally on the right side of the tooth on right handed bevel gears, and on the left side of the tooth on left handed bevel gears (sees Fig. 2.16). In the special case of an inverse spiral, this situation is reversed [SEIB03]. 2.2 Gear Geometry 29 On spiral bevel gears with a positive hypoid offset, load conditions are favorable if the concave tooth flank of the pinion drives the convex tooth flank of the wheel (see Sect. 3.4.2). Usage in car axle drives has led to the choice of this flank of the pinion tooth for the drive mode, i.e. when the engine is moving the vehicle forward. Inversely, the convex tooth flank of the pinion is loaded when in coast mode, i" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000080_j.engfailanal.2013.03.008-Figure15-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000080_j.engfailanal.2013.03.008-Figure15-1.png", + "caption": "Fig. 15. Shape of cracked gear (7th order, a = 0.8, root crack).", + "texts": [], + "surrounding_texts": [ + "Assuming a crack length of a = 0.8 mm at the pitch circle, when the other parameters are fixed, and only the magnitude of the load is changed, the relationship between load and SIF can be obtained by analysis as shown in Figs. 23 and 24. From these figures, it can be seen that the relationship between load and SIF is linear, that is to say: KI = A1w, KII = A2w, where A1 and A2 are coefficients that are relevant to the main parameters of the gear, position of load, crack length, and so on. From the figures, it can also be seen that the rate of changed for KI, KII is fastest when the load acts at point E." + ] + }, + { + "image_filename": "designv10_14_0000525_s12541-016-0036-6-Figure4-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000525_s12541-016-0036-6-Figure4-1.png", + "caption": "Fig. 4 Contact lines on reference tooth surface", + "texts": [ + "14 However, the shapes of the paths of the points on the cutting edge are not straight lines but are curves in the generated facehobbed progress. It becomes complicated and impractical to describe the reference surface of a generated face-hobbed hypoid gear using a point on the cutting edge and the cutter rotation angle.15 For this reason, in this study, a method for describing the tooth surface using the theoretical contact lines is proposed,and this method is suitable to describe both the generated and non-generated hypoid gear using the grid points. As shown in Fig. 4, the point M(ugM, \u03b8gM) is the center of the gear tooth surface. The point Mt(ugt, \u03b8gt) is the intersection between the contact line through the point M and tip line.The distance from Mt to tip line can be represented as (6) The current rotation angle of the gear gt can be represented as (7) Where, \u03c8gM is the rotation angle of the gear corresponding to Mt. With Eqs. (6)~(7), the coordinate of Mt(ugt, \u03b8gt) can be solved by Newton-Raphson Method. Consider the distance from Mi(ugi, \u03b8gi) to tip line is f(ugi, \u03b8gi), and the current rotation angle \u03c8gi is g(ugi, \u03b8gi)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001661_j.triboint.2019.106131-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001661_j.triboint.2019.106131-Figure1-1.png", + "caption": "Fig. 1. (a) Schematic representation of the modified dynamic load PoD tribometer rig.", + "texts": [ + " The cam is attached with an additional electric motor with controlled rotational speed to regulate the frequency of dynamic loading. The test rig is designed for the preliminary study of effect of driving forces on WEA formation in WTG. Hence the typical contact pressure of the rig can vary between 1 and 2 GPa [5] and loading frequency between 1.5 and 4.5Hz [13]. The schematic representation of the modified dynamic load PoD tribometer rig (IP file No: 201741035045, filing date: 03-10-2017. PCT Application No: PCT/IN2018/050628, filing date: 03-10-2018) is shown in Fig. 1(a). The sliding action of the pin contacting with the rotating disc generates very high frictional energy due to \ufffd200% slippage (pure sliding) between them. Frictional energy generation leads to the intense formation of active nascent steel surface, followed by lubricant decomposition due to severe sliding and asperity friction. The hydrogen generation as a by-product of lubricant decomposition is one of the major root causes of WEAs/WECs failures in bearing [28]. Conventional RCF test rig operates cyclic loading in radial direction and distributes the load throughout the circumference of the rollers/ bearing race due to its rolling action" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000037_751476-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000037_751476-Figure2-1.png", + "caption": "Figure 2: Geometrical parameters of the tooth foundation.", + "texts": [ + " (1) Shear stiffness \ud835\udc58\ud835\udc60 is calculated by 1 \ud835\udc58\ud835\udc60 = \u222b \ud835\udf11 2 \u2212\ud835\udf11 1 1.2 (1 + ]) (\ud835\udf112 \u2212 \ud835\udf11) cos\ud835\udf11cos 2 \ud835\udf111 \ud835\udc38\u0394\ud835\udc59 [(\ud835\udf112 \u2212 \ud835\udf11) cos\ud835\udf11 + sin\ud835\udf11] d\ud835\udf11. (2) Axial compressive stiffness \ud835\udc58\ud835\udc4e is 1 \ud835\udc58\ud835\udc4e = \u222b \ud835\udf11 2 \u2212\ud835\udf11 1 (\ud835\udf112 \u2212 \ud835\udf11) cos\ud835\udf11sin 2 \ud835\udf111 2\ud835\udc38\u0394\ud835\udc59 [(\ud835\udf112 \u2212 \ud835\udf11) cos\ud835\udf11 + sin\ud835\udf11] d\ud835\udf11, (3) where \u210e, \ud835\udc65, \ud835\udc51, \ud835\udf111, \ud835\udf112, and \ud835\udf11 are shown in Figure 1. \ud835\udc38, \ud835\udc3a, and V represent Young\u2019s modulus, shear modulus, and Poisson\u2019s ratio, respectively. The stiffness of tooth foundation can be obtained by 1 \ud835\udc58\ud835\udc53 = cos2\ud835\udf111 \ud835\udc38\u0394\ud835\udc59 \u00d7 {\ud835\udc3f \u2217 ( \ud835\udc62\ud835\udc53 \ud835\udc46\ud835\udc53 ) 2 +\ud835\udc40 \u2217 ( \ud835\udc62\ud835\udc53 \ud835\udc46\ud835\udc53 ) + \ud835\udc43 \u2217 (1 + \ud835\udc44 \u2217tan2\ud835\udf111)} , (4) where \ud835\udc62\ud835\udc53 and \ud835\udc46\ud835\udc53 are given in Figure 2. The coefficients \ud835\udc3f\u2217, \ud835\udc40 \u2217, \ud835\udc43\u2217, and \ud835\udc44\u2217 can be calculated by polynomial functions [26] as follows: \ud835\udc4b \u2217 (\u210e\ud835\udc53, \ud835\udf03\ud835\udc53) = \ud835\udc34 \ud835\udc56 \ud835\udf03 2 \ud835\udc53 + \ud835\udc35\ud835\udc56\u210e 2 \ud835\udc53 + \ud835\udc36\ud835\udc56\u210e\ud835\udc53 \ud835\udf03\ud835\udc53 + \ud835\udc37\ud835\udc56 \ud835\udf03\ud835\udc53 + \ud835\udc38\ud835\udc56\u210e\ud835\udc53 + \ud835\udc39\ud835\udc56, (5) where \u210e\ud835\udc53 = \ud835\udc5f\ud835\udc53/\ud835\udc5fint. \ud835\udc5f\ud835\udc53, \ud835\udc5fint and \ud835\udf03\ud835\udc53 are defined in Figure 2 and the constants \ud835\udc34 \ud835\udc56, \ud835\udc35\ud835\udc56, \ud835\udc36\ud835\udc56,\ud835\udc37\ud835\udc56, \ud835\udc38\ud835\udc56, and \ud835\udc39\ud835\udc56 are given in Table 1. by guest on July 9, 2016ade.sagepub.comDownloaded from The Hertzian contact stiffness \ud835\udc58\u210e of two meshing teeth is constant along the entire line of action [20]. \ud835\udc58\u210e can be expressed as the following equation: \ud835\udc58\u210e = \ud835\udf0b\ud835\udc38\u0394\ud835\udc59 4 (1 \u2212 ]2) . (6) The total equivalent mesh stiffness of a tooth pair inmesh can be expressed as \ud835\udc58\ud835\udc56 = 1 \u00d7 ( 1 \ud835\udc58\u210e + 1 \ud835\udc58\ud835\udc4f1 + 1 \ud835\udc58\ud835\udc601 + 1 \ud835\udc58\ud835\udc4e1 + 1 \ud835\udc58\ud835\udc531 + 1 \ud835\udc58\ud835\udc4f2 + 1 \ud835\udc58\ud835\udc602 + 1 \ud835\udc58\ud835\udc4e2 + 1 \ud835\udc58\ud835\udc532 ) \u22121 , (7) where the subscripts 1 and 2 represent the pinion and gear, respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001901_j.mechmachtheory.2020.103823-Figure9-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001901_j.mechmachtheory.2020.103823-Figure9-1.png", + "caption": "Fig. 9. Shape of the pinion slice.", + "texts": [ + " Bending stiffness of the gear slice: 1 k b,g i = \u222b \u03d5 F,g i g g 3 {[ x iu,g i (\u03d5 F,g i ) \u2212 x iu,g i (\u03d5 iu,g i ) ] cos \u03b1F,g i \u2212 y iu,g i (\u03d5 F,g i ) sin \u03b1F,g i }2 2 [ y iu,g i ( \u03d5 iu,g i )]3 E g w [ x iu,g i ( \u03d5 iu,g i )]\u2032 d\u03d5 iu,g i + \u222b g g h g 3 {[ x iu,g i (\u03d5 F,g i ) \u2212 x tu,g i (\u03d5 tu,g i ) ] cos \u03b1F,g i \u2212 y iu,g i (\u03d5 F,g i ) sin \u03b1F,g i }2 2 [ y tu,g i ( \u03d5 tu,g i )]3 E g w [ x tu,g i ( \u03d5 tu,g i )]\u2032 d\u03d5 tu,g i (33) Axial compressive stiffness of the gear slice: 1 k a,g i = \u222b \u03d5 F,g i g g sin 2 \u03b1F,g i 2 y iu ,g i ( \u03d5 iu ,g i ) E g w [ x iu ,g i ( \u03d5 iu ,g i )]\u2032 d \u03d5 iu ,g i + \u222b g g h g sin 2 \u03b1F,g i 2 y tu ,g i ( \u03d5 tu ,g i ) E g w [ x tu ,g i ( \u03d5 tu ,g i )]\u2032 d \u03d5 tu ,g i (34) Shear stiffness of the gear slice: 1 k s,g i = \u222b \u03d5 F,g i g g 1 . 2 cos 2 \u03b1F,g i 2 y iu,g i ( \u03d5 iu,g i ) G g w [ x iu,g i ( \u03d5 iu,g i )]\u2032 d\u03d5 iu,g i + \u222b g g h g 1 . 2 cos 2 \u03b1F,g i 2 y tu,g i ( \u03d5 tu,g i ) G g w [ x tu,g i ( \u03d5 tu,g i )]\u2032 d\u03d5 tu,g i (35) 3.3.2. The pinion slice The stiffness calculation of the pinion without MEs is the same as that of the gear and is therefore not detailed in this paper. There are two types of pinion slices, corresponding to different MEs (see Fig. 9 ). The type of pinion slice in Fig. 9 (a) is defined as Type I, i.e., UDIT is located on the upper involute surface profile. The type of slice in Fig. 9 (b) is defined as Type II, i.e., UDIT is located on the upper transition surface profile. The independent variables Q uu i and Q uc i corresponding to UDIT for each type can be solved by Eqs. (36) and (37) , respectively. x aiu ,p i ( Q uu i , \u03b8 p i , e p , i ) = x atd ,p i ( b p , \u03b8 p i , e p , i ) (36) x atu,p i ( Q uc i , \u03b8 p i , e p , i ) = x atd,p i ( b p , \u03b8 p i , e p , i ) (37) According to Fig. 9 , the bending stiffness, axial compressive stiffness and shear stiffness of the pinion slice corresponding to Types I and II and represented by Eqs. (25) \u2013( 30 ) can be rewritten in a computable form by Eqs. (38) and ( 39 ), Eqs. (40) and ( 41 ) and Eqs. (42) and ( 43 ), respectively. Bending stiffness of the pinion slice: Type \u2160 : 1 k b,p i = \u222b \u03d5 F,p i Q uu i 12 {[ x aiu,p i ( u F,p i ) \u2212 x aiu,p i ( u iu,p i )] cos \u03b1F,p i \u2212 y aiu,p i ( u F,p i ) sin \u03b1F,p i }2 E p w [ y aiu,p i ( u iu,p i ) \u2212 y aid,p i ( u id,p i )]3 \u2202x aiu,p i ( u iu,p i ) \u2202\u03d5 iu,p i d\u03d5 iu,p i + \u222b Q uu i g p 12 {[ x aiu,p i ( u F,p i ) \u2212 x aiu,p i ( u iu,p i )] cos \u03b1F,p i \u2212 y aiu,p i ( u F,p i ) sin \u03b1F,p i }2 E p w [ y aiu,p i ( u iu,p i ) \u2212 y atd,p i ( u td,p i )]3 \u2202x aiu,p i ( u iu,p i ) \u2202\u03d5 iu,p i d\u03d5 iu,p i + \u222b g p h p 12 {[ x aiu,p i ( u F,p i ) \u2212 x atu,p i ( u tu,p i )] cos \u03b1F,p i \u2212 y aiu,p i ( u F,p i ) sin \u03b1F,p i }2 E p w [ y atu,p i ( u tu,p i ) \u2212 y atd,p i ( u td,p i )]3 \u2202x atu,p i ( u tu,p i ) \u2202\u03d5 tu,p i d\u03d5 tu,p i (38) Type \u2161 : 1 k b,p i = \u222b \u03d5 F,p i g p 12 {[ x aiu,p i ( u F,p i ) \u2212 x aiu,p i ( u iu,p i )] cos \u03b1F,p i \u2212 y aiu,p i ( u F,p i ) sin \u03b1F,p i }2 E p w [ y aiu,p i ( u iu,p i ) \u2212 y aid,p i ( u id,p i )]3 \u2202x aiu,p i ( u iu,p i ) \u2202\u03d5 iu,p i d\u03d5 iu,p i + \u222b g p Q uc i 12 {[ x aiu,p i ( u F,p i ) \u2212 x atu,p i ( u tu,p i )] cos \u03b1F,p i \u2212 y aiu,p i ( u F,p i ) sin \u03b1F,p i }2 E p w [ y atu,p i ( u tu,p i ) \u2212 y aid,p i ( u id,p i )]3 \u2202x atu,p i ( u tu,p i ) \u2202\u03d5 tu,p i d\u03d5 tu,p i + \u222b Q uc i h p 12 {[ x aiu,p i ( u F,p i ) \u2212 x atu,p i ( u tu,p i )] cos \u03b1F,p i \u2212 y aiu,p i ( u F,p i ) sin \u03b1F,p i }2 E p w [ y atu,p i ( u tu,p i ) \u2212 y atd,p i ( u td,p i )]3 \u2202x atu,p i ( u tu,p i ) \u2202\u03d5 tu,p i d\u03d5 tu,p i (39) Herein, u F,p i = [ \u03d5 F,p i , \u03b8 p i , e p , i ]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000494_s11665-015-1572-4-Figure4-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000494_s11665-015-1572-4-Figure4-1.png", + "caption": "Fig. 4 Temperature Contours for small substrate (Color figure online)", + "texts": [ + " The second idea was to use two different substrates of differing sizes while depositing sequentially using the exact same parameters. One of the substrates would be significantly larger, resulting in a much higher cooling rate. The dimensions of the smaller substrate are 25 * 6 * 4 mm in length, width, and thickness, respectively. The transient Journal of Materials Engineering and Performance Volume 24(8) August 2015\u20143131 temperature distribution contour plot for the first deposited layer is illustrated in Fig. 4. The average of the cooling rate along laser track is 1500 C/s calculated at 890 C. The larger substrate measured 50 * 50 * 25 mm in length, width, and thickness, respectively. Figure 5 illustrates the transient temperature distribution contour plot for the first deposited layer for the large substrate size case. The average of the cooling rate along laser track is 5000 C/s calculated at 890 C All of the deposition experiments were conducted at the Laser-Aided Manufacturing Process (LAMP) lab at the Missouri University of Science and Technology" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001874_s00170-019-04738-3-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001874_s00170-019-04738-3-Figure2-1.png", + "caption": "Fig. 2 Cutting coordinate system of pinion", + "texts": [ + " The main profile for machining the pinion working surface is as follows: r f \u00bc sp 0 f sp 1 T \u00f01\u00de where sp and f(sp) are the parameters of the main profile. Base on the coordinate transformation, the position vector and unit normal vector of cutting cone are as follows: rp \u00bc Rp \u00fe spsin\u03b11 \u00fe f sp cos\u03b11 cos\u03b8p Rp \u00fe spsin\u03b11 \u00fe f sp cos\u03b11 sin\u03b8p \u2212spcos\u03b11 \u00fe f sp sin\u03b11 1 2 664 3 775 \u00f02\u00de np \u00bc \u2212 cos\u03b11\u2212 f 0 sp sin\u03b11 cos\u03b8p \u2212 cos\u03b11\u2212 f 0 sp sin\u03b11 sin\u03b8p \u2212 sin\u03b11 \u00fe f 0 sp cos\u03b11 2 4 3 5 \u00f03\u00de where \u03b8p is the parameters of the main profile. Rp is the cutter radius. \u03b11 is the profile angle. f 0 sp \u00bc \u2202 f sp =\u2202sp \u00f04\u00de Figure 2 is the machining coordinate system, and Sp and S1 are connected to the cutter and the machined pinion. The pinion tooth surface is built based on coordinate transformation: r1 \u00bc M 1prp n1 \u00bc L1pnp \u00f05\u00de where M1p are the homogenous coordinate transformation matrix. L1p are the upper-left3 \u00d7 3sub-matrix ofM1p. M 1p \u00bc a11 a12 a13 a14 a21 a22 a23 a24 a31 a32 a33 a34 a41 a42 a43 a44 2 664 3 775 where a11 \u00bc cos\u03b31cos\u03c6p; a12 \u00bc \u2212cos\u03b31sin\u03c6p; a13 \u00bc sin\u03b31; a14 \u00bc Sr1cos\u03b31cos \u03c6p \u00fe q1 \u2212Xb1sin\u03b31\u2212XG1; a21 \u00bc cos\u03c61sin\u03c6p\u2212cos\u03c6psin\u03b31sin\u03c61; a22 \u00bc cos\u03c61cos\u03c6p \u00fe sin\u03c6psin\u03b31sin\u03c61; a23 \u00bc cos\u03b31sin\u03c61; a24 \u00bc Em1cos\u03c61\u2212X b1cos\u03b31sin\u03c61 \u00fe Sr1sinq1 cos\u03c61 sin\u03c6p \u00fe cos\u03c6p \u00fe sin\u03b31sin\u03c61 sin\u03c6p\u2212cos\u03c6p h ; a31 \u00bc \u2212sin\u03c61sin\u03c6p\u2212cos\u03c6psin\u03b31cos\u03c61; a32 \u00bc \u2212cos\u03c61sin\u03b31sin\u03c6p\u2212cos\u03c6psin\u03c61; a33 \u00bc cos\u03b31cos\u03c61; a34 \u00bc \u2212Em1sin\u03c61\u2212X b1cos\u03b31cos\u03c61\u2212Sr1 sin\u03c61sin q1\u00fe\u03c6p \u00fe sin\u03b31cos\u03c61cos q1\u00fe\u03c6p h i ; a41 \u00bc a42 \u00bc a43 \u00bc 0; a44 \u00bc 1: The arc-edge cutter of gear is shown in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002158_13506501211010556-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002158_13506501211010556-Figure3-1.png", + "caption": "Figure 3. Computational mesh for the whole domain.", + "texts": [ + "1, 6, 14\u201323 With these assumptions, the rheology of the lubricant oil and the surface elastic deformation of the elements can be neglected. The EHL film is commonly calculated by special tools, which is not the focus of this study. Several studies have revealed that the polyhedral mesh has better accuracy, lower memory demand, and faster convergence behavior than the tetrahedral mesh.37\u201339 In this paper, the computational domain is divided by a polyhedron unstructured mesh in view of its complicated construction. The flow field mesh scheme contains 2,953,781 cells and 9,750,843 nodes, as shown in Figure 3. The number of tetrahedral mesh cells under the same size as the polyhedral mesh is 8,592,525 and the number of nodes is 2,989,654. Moreover, the mesh quality of the former is higher than that of the latter. For example, the maximum aspect ratios of polyhedral and tetrahedral meshes are 40.68 and 65.89, respectively. The skewness of polyhedral and tetrahedral meshes is 0.81 and 0.93, respectively. It takes about 460 and 180 s to construct polyhedral mesh and tetrahedral mesh, respectively. However, the solution results converge to the same accuracy, and the former only consumes half of the computational resources (computation time) of the latter" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001796_j.oceaneng.2020.108033-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001796_j.oceaneng.2020.108033-Figure2-1.png", + "caption": "Fig. 2. Schematic diagram of the LVS guidance principle.", + "texts": [ + " A= \u23a1 \u23a2 \u23a2 \u23a3 w11 w12 \u22ef w1m w21 w22 \u22ef w2m \u22ee \u22ee \u22f1 \u22ee wn1 wn2 \u22ef wnm \u23a4 \u23a5 \u23a5 \u23a6 (7) In the marine practice, the planned route of the underactuated ship is usually made by the ship officer, and the waypoint information Wi = (xi, yi) will be fully utilized. It can be seen easily that the navigating reference path could be splitted into two parts: the straight line and the curve one. For this consideration, the logical virtual ship (LVS) guidance principle has been proposed to generate the smooth waypoints-based attitude reference based on the piecewise logical programming strategy (see Fig. 2). It is noted that the virtual ship (8) is assumed as an ideal ship with no inertia and no damping effects. \u23a7 \u23aa\u23a8 \u23aa\u23a9 x\u0307d = ud cos(\u03c8d) y\u0307d = ud sin(\u03c8d) \u03c8\u0307d = rd (8) Part 1: The virtual ship sails at a fixed speed ud when sailing along the straight line Ls, and the corresponding sailing time could be calculated according to ts = Ls/ud. In particular, the sailing speed of the virtual ship can be adjusted by the ship officer on the basis of the specific mission requirements. Part2: For the curve one, one needs to determine the real-time turning radius Ri (9) to obtain the related turning rate rd, where the maneuvering ability and practical heading change \u0394\u03c6i of underactuated ships will be fully considered" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000937_035001-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000937_035001-Figure1-1.png", + "caption": "Figure 1. Basic model representing the physics of a moving projectile.", + "texts": [ + " These formulas are also beneficial to specialists actively working in the field as well as undergraduate and graduate students of engineering. Although the governing equations can be found in many sources, some of which are already cited here, for a complete description of the physical phenomenon, we briefly outline the derivation of them in what follows. Let us consider the motion of a point mass m or projectile launched at an angle 0q with an initial speed V0 encountering a quadratic resistance force Eur. J. Phys. 37 (2016) 035001 M Turkyilmazoglu 2 2= . A simple sketch of the physical event is displayed in figure 1. It is assumed that the projectile is at the initial instant standing at the origin. Here g is the acceleration of gravity, k0 is the drag constant and V is the speed of the object in the tangential direction. At any angle \u03b8 (or time t), the equations of motion along the tangential direction due to Newton\u2019s second law and the principal normal direction due to the centrifugal law are written as m V t mg mgk V V t V d d sin , 0 , 10 2 0( ) ( )q= - - = = mV t mg t d d cos , 0 , 20( ) ( )q q q q= - = = The horizontal and vertical distances (x, y) are related to the motion of the projectile via the kinematic relations x t V y t V d d cos , d d sin " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002406_j.engfailanal.2021.105456-Figure7-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002406_j.engfailanal.2021.105456-Figure7-1.png", + "caption": "Fig. 7. Equivalent (von Misses) Stress distribution in the bucket wheel excavator model for the load case with inertial forces (left), the critical joint with highest stresses (top right), and magnitudes (bottom right).", + "texts": [ + " The most interesting case for this analysis was the combination of dead weight and working load. While a possibility of failure due to high inertial forces caused inadequate operation of the bucket wheel excavator boom does exist, it is extremely unlikely to happen in real S. Kirin et al. Engineering Failure Analysis 126 (2021) 105456 S. Kirin et al. Engineering Failure Analysis 126 (2021) 105456 working conditions, due to safety systems in place, and for this reason, that particular scenario was not considered in this analysis. Stress distribution is shown in Fig. 7. The locations of the highest stress can be clearly seen in the model. Their magnitudes are significantly lower than the allowed stresses for this load case, and yet failure of the whole structure occurred. It should also be noted that the entire structure got separated from its foundation (Fig. 3), instead of failing at stress concentration locations, shown in Fig. 7, which would be the expected. This suggests that the design and regular working conditions were not the cause of failure, as the stresses are on the safe side. Hence, this calculation confirms that it was indeed human error that resulted in the catastrophic failure. This is further reinforced by the fact that this load case, in theory is not the most dangerous one \u2013 the case with inertial forces produces higher loads, as shown by the the following table. The individual values in each row are related to four selected finite elements from the model" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002401_j.mechmachtheory.2021.104371-Figure6-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002401_j.mechmachtheory.2021.104371-Figure6-1.png", + "caption": "Fig. 6. Three-dimensional dynamic model of distributed meshing of a helical gear pair.", + "texts": [ + " \u03b2 is the helix angle of the helical gear pair. The instantaneous pressure angle of each tooth in a slice can be expressed as Eq. (16) [18] \u03b1i = arctan ( \u03d5 1 ,n + \u03d5 2 ,n ) (16) If the sheet teeth are engaged, the instantaneous pressure angle of each of the sheet teeth should be between the gear minimum pressure angle \u03b1L and the maximum pressure angle \u03b1U , in Fig. 5 (b). According to the above method of gear pair cutting, the three-dimensional dynamic model of a distributed helical gear pair is obtained, as shown in Fig. 6: The dynamic meshing force of a helical gear pair can be calculated by the sum of the meshing forces of the respective gear blades, as shown in Eq. (17) : F\u0304 m = N \u2211 i = i F m,i i = 1 , 2 , \u00b7 \u00b7 \u00b7 , N (17) where F m,i is the dynamic meshing force of the i th slice of the gear pair in the direction of the meshing line, which can be expressed as: F m,i = k i e \u03b4i ( b, i (t) ) + c i m \u02d9 \u03b4i ( b, i (t) ) (18) where i = 1 , \u00b7 \u00b7 \u00b7 , N and c i m are the meshing damping of the i th slice of gear pair, c i m (t) = \u03be \u221a k i e m 1 m 2 / ( m 1 + m 2 ) " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000093_j.engfailanal.2013.02.003-Figure4-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000093_j.engfailanal.2013.02.003-Figure4-1.png", + "caption": "Fig. 4. Tooth back thickness and tooth root fillet radius.", + "texts": [ + " So we concluded the engaging status might be not well, and the contact area was only on the large end of the tooth surface. As there were hundreds of successful lifespan tests equipped with the straight bevel gears, and no similar failure occurred before the replacement of machining equipment, careful examinations on structural geometric parameters were performed. Inconsistency in the structural geometric parameters was found between the failed gears and previous gears which had been working well. The chief discrepant structural parameters were tooth root fillet radius and tooth back thickness as shown in Fig. 4 and in Table 2. In Table 2, the design tolerance of tooth back thickness was from 0.85 mm to 1.25 mm. The dimensions of previous gears were all bigger than 0.95, while the dimensions of failed ones were smaller than 0.95. Although the dimensions satisfied the design tolerance, but it was still a question whether the design tolerance was suitable. The design tolerance of tooth root fillet radius was from 0.30 mm to 0.50 mm. The dimensions of previous gears all satisfied it, but the dimensions of failed gears were too small obviously" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001573_j.mechmachtheory.2018.12.031-Figure6-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001573_j.mechmachtheory.2018.12.031-Figure6-1.png", + "caption": "Fig. 6. 5-PRUR overconstrained parallel mechanism and structural decomposition of a limb.", + "texts": [ + " (45) is the stiffness model of the whole mechanism, and K w is the stiffness matrix of the mechanism. Generally, K w is nonsingular, so its inverse exists and can be expressed as C w = ( K w ) \u22121 (47) where C w is called the compliance matrix of the whole mechanism. Based on Eq. (45) , the following equation can be obtained, D = C w ( W e + G p + J ac f K\u03c4 \u2212 J r f r ) (48) which is the compliance model of the whole mechanism. 5. Numerical examples and result comparisons 5.1. Example one A 5-PRUR (U denoting a universal pair) overconstrained parallel mechanism is shown in Fig. 6 (a), in which all R pairs next to the fixed base are parallel to each other, and all R pairs next to the moving platform are parallel to each other [1] . Structural decomposition of limb i is shown in Fig. 6 (b), in which lengths of links and related gravity parameters have been marked in those figures. The diameter of each link is 0.05 m. Each angular contact ball bearing is the outside diameter of 0.052 m, inner diameter of 0.025 m and width of 0.015 m. The material of each component is structural steel with elasticity module of 200 GPa and Poisson\u2019s ratio of 0.3. At a given configuration, all actuation pairs are locked and the mechanism becomes a structure. For simplicity, here, the compliance coefficients of all actuators are considered to be zero" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000769_978-3-662-43893-0-Figure6.2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000769_978-3-662-43893-0-Figure6.2-1.png", + "caption": "Fig. 6.2 Technological angles on the cutting wedge", + "texts": [ + " More and more direct drives with high rotation speeds are being used to achieve high cutting speeds even on smaller spiral bevel gears and hence using smaller tool diameters. Work piece automatic loading into the gear cutting machines is steadily expanded. Efforts to integrate additional processes into the machines, such as deburring, are observed. The geometry of the bevel gear tooth flank is defined solely by the form and position of the cutting edge of the tool. Aside from the form, the technological angles of the cutting wedge are important for the manufacturing process. Figure 6.2 defines these angles. For the sake of simplicity, a cutting wedge is shown with a plane front face and plane relief surfaces, and without tool edge radii. 236 6 Manufacturing Process The cutting wedge consists of three sections: the primary cutting edge (1), the tip cutting edge (2) and the secondary cutting edge (3). The front face is the plane which is made by these three sections of the cutting wedge. It is inclined by the side rake angle (5) in relation to the plane normal to the direction of primary motion (4)", + " The machine then generates till the end roll position, the tool is withdrawn from the tooth slot and returned to the starting position for another plunge. Meanwhile the machine does one indexing turn and is ready for the next tooth slot (Fig. 6.17). On bevel gears cut with the continuous indexing method, the tool which is employed determines the best plunge position. With a two-blade-per-group cutter head, plunging is made near themiddle of the generating zone since it is there that the inner and outer blades are cutting under the same load and, if at all, to a slight extent with the secondary cutting edge (Fig. 6.2) which has unfavorable cutting properties. On a three-blade-per-group cutter head, plunge cutting preferentially occurs at the heel to let the pre-cutter remove as much stock as possible from the tooth slot. The pre-cutter is usually the first blade in the group to enter the tooth slot, but is somewhat shortened in height such that it cannot work effectively if it is plunged in another position. In the continuous indexingmethod, the final plunge cutting position is often not the first generating position which must then be reached by rotating the cradle" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001304_j.mechmachtheory.2019.103697-Figure4-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001304_j.mechmachtheory.2019.103697-Figure4-1.png", + "caption": "Fig. 4. Decomposed diagram of the force-sensitive element.", + "texts": [ + " This paper use the theories that the stiffness of the parallel mechanism is equal to the sum of each branch stiffness, the flexibility of the series mechanism is equal to the sum of each branch flexibility, and the stiffness and flexibility are mutual inverse matrix [32 , 33] to establish the stiffness matrix of the sensor. First, we decompose the sensor into basic elements and establish their stiffness/flexibility matrices. Then the systemlevel stiffness matrix can be established with the matrices. As shown in Fig. 4 , part1 and part3 are the flexible spherical joints with flexible thin beams that have the same sizes. Part2 is the flexible long beam with a rectangular cross-section and the strain gauges are attached to it. Part11, 13, 31, 33 are the flexible thin beams with a rectangular cross-section that have the same sizes. Part12, 32 are the flexible spherical joints. Therefore, the flexibility matrices of the flexible beam with a rectangular cross-section and the flexible spherical joint should be established, respectively", + " T j i denotes the transformation matrix from the reference coordinate system i to the coordinate system j . R j i denotes the rotation matrix that the reference coordinate system i is relative to the coordinate system j . r j i denotes the position coordinate of the reference coordinate system i in the coordinate system j . S( r j i ) denotes skew-symmetric matrix of the position coordinate. r j i = ( r x r y r z )T (11) S ( r j i ) = [ 0 \u2212r z r y r z 0 \u2212r x \u2212r y r x 0 ] (12) The force-sensitive element is decomposed into three parts: part1, part2, part3, shown as Fig. 4 . The flexibility matrices of part1 and part3 are the same, so the paper only needs to establish flexibility matrices of part1 and part2. The part1 also can be decomposed into three parts: part11, part12, part13, shown as Fig. 7 . o i x i y i z i denotes the fixed coordinate system located at the bottom center of part11. i = 1 , 2 . . . 6 denote the serial number of the force-sensitive element. o i 1 x i 1 y i 1 z i 1 denotes the local coordinate system located at the end of part11. o i 2 x i 2 y i 2 z i 2 denotes the local coordinate system located at the end of part12" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000769_978-3-662-43893-0-Figure5.2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000769_978-3-662-43893-0-Figure5.2-1.png", + "caption": "Fig. 5.2 Vibration model of a single stage cylindrical gear set", + "texts": [ + " RWTH Aachen (1971) References 195 Chapter 5 Noise Behavior The main causes of noise generation in a gearbox are the gears themselves. Gear teeth meshing periodically excite vibrations in a gear system, on the one hand causing the emission of direct air-borne noise and on the other hand conveying structure-borne noise to the surface of the housing, where it is radiated as indirect air-borne noise (Fig. 5.1). To clarify the mechanisms which effectively determine the vibration behavior of gears, Fig. 5.2 shows the torsional vibration model of a single stage cylindrical gear set. The observations which follow apply to both cylindrical and bevel gears. The \u00a9 Springer-Verlag Berlin Heidelberg 2016 J. Klingelnberg (ed.), Bevel Gear, DOI 10.1007/978-3-662-43893-0_5 197 model is essentially composed of the two gear bodies of the gear set connected by means of a coupling element. This torsional vibration model is linked to the gear housing via spring and damper elements representing the bearing and housing stiffnesses and their damping properties" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000894_tro.2015.2489499-Figure8-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000894_tro.2015.2489499-Figure8-1.png", + "caption": "Fig. 8. 3\u2013PPS robot and its hidden robot model (the gray joints denote the actuated joints). (a) 3\u2013PPS robot. (b) Its hidden robot model: a 3\u2013PPS robot with no actuators.", + "texts": [ + " Three main classes of parallel robots belong to this category (the list is not exhaustive, but groups the most usual and known robots in the community): 1) robots with legs whose directions are constant for all robot configurations: for these robots, the anchor point location of the observed links cannot be found through the use of the simplified kinematic models. This are the cases of planar 3\u2013PPR (see Fig. 7) and 3\u2013PPR robots [25], [37] and of certain spatial robots such as the 3\u2013[PP]PS robots6 (with 3\u2013PPS robots (with three dof [38] (see Fig. 8) or with six dof\u2014e.g., the MePaM [36]). It is obvious that for robots with legs whose directions are constant in the whole workspace, it is not possible to estimate the platform pose from the leg directions only; 6[PP] means an active planar chain able to achieve two dof of translation, such as PP or RR chains. 2) robots with legs whose directions are constant for an infinity of (but not all) robot configurations: this is the case of PRRRP robots with all P parallel [see Fig. 9(a)] and of Delta-like robots actuated via P joints for which all P are parallel (such as the UraneSX (see Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000457_j.ijmachtools.2014.05.009-Figure7-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000457_j.ijmachtools.2014.05.009-Figure7-1.png", + "caption": "Fig. 7. The thermal model of the high speed press.", + "texts": [ + " When the grid was meshed irregularly, it could be degraded into a five-sided prism element or a tetrahedral pyramid element. Compared with the degradation element, the hexahedral elementFig. 3. Force sketch of the slider-crank mechanism. can provide more accurate numerical solutions. Therefore, it is adopted in this dynamic analysis. The material properties of each component are listed in Table 1 and the FE model for the combined frame is established using the ANSYS software package, as shown in Fig. 7. As the combined frame is composed of several components, the contact elements CONTA174 and TARGE170, commonly used in 3-D structural contact analysis, are introduced to simulate the effect of assembly feature between solid contact surfaces. For the high speed press system, there are three major heat sources in the press system: (a) heat generated by the support bearings as a result of friction between the balls and races; (b) heat generated by the motor; (c) heat generated by the guide way. The thermal contact resistance R can be expressed as[16] R\u00bc 1=\u00f0Ahc\u00de \u00f06\u00de where A is the apparent contact area of the contact region, hc is the contact conduction coefficient" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001629_j.renene.2019.09.049-Figure20-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001629_j.renene.2019.09.049-Figure20-1.png", + "caption": "Fig. 20. The iron loss density distribution of the LHPM generator with silicon steel 50A1000.", + "texts": [ + " Therefore, the LHPM generator with SPCC core could be a good candidate if the low speed and low cost are mainly considered as the objectives. Fig. 19 shows the output power comparison results between the prediction by simulation and experiments of the prototype generator. It shows that the results between the prediction and experiments have little difference with SPCC and silicon steel. It demonstrates the FEA simulation method which is used for optimization in this paper has high precision. The iron loss density distribution of the LHPM generator with silicon steel 50A1000 is shown in Fig. 20. It shows most of the iron losses distribute in the stator. In other words, iron loss in the rotor core is little. Therefore, as one of the low cost generators, a novel LHPM generator with hybrid material cores, of which the stator is made of silicon steel and the rotor is made of SPCC is proposed. Fig. 21 shows the performance comparison of three types of LHPM generators at n\u00bc 200 rpm. Compared with the LHPM 00 150 200 eed n (rpm) MF of the LHPM generator. 0 200 400 600 800 1000 947.5 964.5 967 98" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000576_gt2016-57454-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000576_gt2016-57454-Figure2-1.png", + "caption": "FIGURE 2. Simplified computational domains: (a) gear pair configuration (b) single rotating gear configuration", + "texts": [ + " The test rig arrangement has been numerically reproduced in the present work, firstly to test the numerical methodologies for calculating the windage losses of a single wheel, secondly to decouple the windage contribution from the total power losses, allowing by difference the evaluation of meshing losses. For more details the reader is referred to [1]. 3 Copyright \u00a9 2016 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/conferences/asmep/89496/ on 05/29/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use In the numerical simulations some geometrical simplifications have been adopted both for gear pair and single gear configuration. Despite the cantilever shafts, a symmetry boundary condition has been used, as shown in Figure 2, in order to significantly reduce the computational effort since only one half of the geometry has to be simulated. Furthermore for the meshing gear pair configuration, the wheelbase has been increased, with respect to the experimental one, in order to guarantee the continuity of the computational domain by preventing the contact between the teeth. The adopted computational domain is depicted in Figure 2 (a). Figure 2 (b) shows the rotating single gear numerical domain where another simplification has been applied since the presence of the second shaft, on which no gear is mounted, was neglected as suggested by Gorla et al. [1]. Among the available moving boundaries approaches, during the current work two numerical unsteady RANS approaches have been tested in order to correctly reproduce the gear meshing: the Immersed Solid approach and the Dynamic Mesh approach. These methods use two considerably different strategies for simulating the gear meshing phenomenon, as well as the discretization methods and the computational grids to be adopted", + " The same method was also adopted for the computational grid definition in the single rotating gear configuration. A 3.1\u00b7106 elements mesh was generated with 2.2\u00b7106 extruded prisms and 9.9\u00b7105 tetrahedral elements. The flow system was treated as isothermal and incompressible, imposing a reference pressure of 6 bar and a temperature of 90\u25e6 C. In these conditions the lubricant properties assume the values reported in Table 2. A no slip condition was imposed on all the walls of the domains, while a symmetry condition was imposed on the blue surface of Figure 2 exploiting the geometry of the test-rig. This condition leads to a remarkable reduction of the computational 5 Copyright \u00a9 2016 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/conferences/asmep/89496/ on 05/29/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use costs, as only a half of the geometry has to be simulated. Despite that, such boundary condition was exploited, since its impact on the resistant torque calculation is considered negligible, as shown by previous works [20]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001687_j.engfailanal.2020.104411-Figure4-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001687_j.engfailanal.2020.104411-Figure4-1.png", + "caption": "Fig. 4. The surface of the SEM images.", + "texts": [ + " The density of the gears having this production technique was determined by taking the average of the results obtained from all gears produced by additive manufacturing. Since the gears produced by the conventional method did not evidence such changes in density, the measurements for the traditional method were repeated only three times. To better understand the gear surfaces and their damages, the surfaces before and after the wear test were examined with a ZEISS GeminiSEM 500 scanning electron microscope and a SOIF SZ780-B2/L Trinocular Stereo Zoom microscope. The surface of the SEM images in the study is shown in Fig. 4. For the determination of wearing elements, oil analysis was performed. The researchers added the oil to the gearbox, and then collected it after the test. This oil is mineral-based and consists of an excessive pressure additive. At the end of test, the researchers determined what elements were present and their quantity as revealed through an [10] standard method. To determine the efficiency of the gears, instantaneous recorded data were used in the input and output torque meters in the wear tester", + " The researchers also applied a surface polishing process to the DMLS-produced gears, and assessed how DMLS post-production surface polishing affected many properties of the gears. Also, for all tests, the gears were driven a total of 105 cycles. Surfaces, hardness and density of gears produced by the different techniques were examined. Microscope images of gear tooth surfaces produced using the hobbing machine are shown in Fig. 5a, while the gears produced by additive manufacturing are shown in Fig. 5b and c. Upon examining Fig. 5, one sees that there are clear differences in the tooth surfaces of the gears dependent upon production technique. In Fig. 4a, the force applied to the gears produced using a hobbing machine is seen on the tooth surface as being wear from the production. Also, after production, both tooth surfaces appear to take on similar properties. The DMLS gear surface is composed of sintered powders and has variable properties across the entire surface (Fig. 5b). Electron microscope images reveal that there are gaps caused by the sintering process. Surface polishing applied to the additively manufactured steel caused the tooth surfaces to change, as can be seen in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000217_s00158-017-1693-5-Figure6-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000217_s00158-017-1693-5-Figure6-1.png", + "caption": "Fig. 6 Contact stresses and crack and subsurface-initiated crack", + "texts": [ + "25 174 mm 20o 14o L 186 mm High-speed stage Gear 88 5 440 mm 20o 14o L 110 mm Pinion 22 5 100 mm 20o 14o R 120 mm Using the Dang Van\u2019s assumption that the fatigue microcrack appears when the mono-crystal reaches the elastic shakedown due to shearing, the number of cycles for crack initiation is defined by (Osman and Velex 2011): Ni \u00bc 1 2 \u03c4max \u00fe 3 \u03c4 0 f =\u03c3 0 f \u22120:5 pH \u03c4 0 f 0 @ 1 A 1=c \u00f012\u00de where \u03c4max and pH are, respectively, the maximum shear stress and hydrostatic stress in the subsurface; \u03c3 0 f and \u03c4 0 f are fatigue strength coefficients for tension/compression and shearing, respectively; and c is a fatigue strength exponent. It is assumed that the crack is initiated at a point where the ratio of the maximum shear stress to the hardness is maximal and the initial crack is parallel to the surface (Choi and Liu 2006; Jiang et al. 1993) as shown in Fig. 6. Assuming the Hertzian contact between two surfaces, the maximum contact pressure pmax can be obtained as a function of the normal contact force FN predicted by the gear dynamics simulation. The greatest maximum shear stress beneath the surface is then obtained as \u03c4max = 0.3pmax at the z0 = 0.786 b0 from the surface, where b0 is the half width of the contact patch (Budynas and Nisbett 2008). On the other hand, the number of cycles that causes the initial crack, parallel to the surface, to reach the surface is calculated using the Paris equation by (Osman and Velex 2011; Dong et al. 2013) Np \u00bc \u222baca0 1 Cp \u0394K\u00f0 \u00dem\u2212 \u0394K0\u00f0 \u00dem\u00f0 \u00de dap \u00f013\u00de where a0 is the half length of the initial crack and ac is half of the critical crack length given by ac = z0/ sin\u03b1 as shown in Fig. 6 (Osman and Velex 2011; Dong et al. 2013). In the preceding equation, \u0394K is the model II stress intensity factor that is quadratic in the ratio of maximum shear stress to the hardness, while \u0394K0 is the threshold for the crack growth (Osman and Velex 2011; Choi and Liu 2006). Cp and m are constants that are identified experimentally (Osman and Velex 2011; Kato et al. 1993). Using the 10-min gear dynamics simulation under a given wind load scenario characterized by vi10 and i j10, the contact fatigue damage for all teeth of a gear with a design variable d is evaluated by Miner\u2019s rule (Melchers 1999) as Dij 10min d; vi10; i j 10 \u00bc \u2211 nij10 k\u00bc1 1 Nijk pijkmax \u00f014\u00de where Nijk is the number of load cycles to failure for each meshing cycle k and nij10 is the number of load cycles of the gear tooth under consideration in the 10-min simulation" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001646_s11771-019-4207-3-Figure8-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001646_s11771-019-4207-3-Figure8-1.png", + "caption": "Figure 8 Spherical hob assembly schematic", + "texts": [ + " Figure 7 shows the difference among each row of teeth on the hob. The angle \u03b5 between the centerline of each gear and the bottom edge is different. The difference value between adjacent blades is: 1 1 2 2\u03c0 / | | N n (9) The cutter teeth are assembled in a certain order on the cutter body groove, and the axial position is fixed using the wedge, both ends are J. Cent. South Univ. (2019) 26: 2704\u22122716 2709 fixed with the end cover, the end cover can be pressed on the cutter body by the compression nuts, as shown in Figure 8. Figure 9 shows the structure of the face gear hobbing machine, including three moving shafts (x, y, z) and two rotating axes (B, C). Motion in the x direction is used to achieve radial feed motion. The motion in the y direction is used to achieve axial feed motion. The z-direction motion is used to realize the additional motion of the cutter relative to the face gear. The rotational motion in the B direction is used to realize the rotation of the face gear along its own axis. The rotational motion in the C direction is used to realize the rotation of the cutter around its own axis" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000520_s40684-016-0008-4-Figure5-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000520_s40684-016-0008-4-Figure5-1.png", + "caption": "Fig. 5 Lamina model", + "texts": [ + " Factors reflecting the actual operating conditions and dimensions of the gears to be used in practice are then applied. The equation for the pitting stress limit is as follows: (8) The life prediction of rolling element bearings in this study was carried out based on the ISO/TS 16281.15 The elastic deformation of the rolling elements in a radial roller bearing can be described by a thin lamina model. To calculate the elastic deformation of a roller, a roller is divided into at least 30 thin, lamina of equal size, as shown in Fig. 5. The load-deformation equation of thin lamina k of roller j is given by Eq. (9). (9) The elastic deformation of roller j with respect to the radial deformation of the raceway \u03b4r is given by Eq. (10), and the misalignment angle between the inner and outer raceways is given by Eq. (11), as shown in Fig. 6. The elastic deformation of the kth thin lamina of the j th roller can be obtained from Eq. (12) using Eqs. (10) and (11). If the right-hand side of Eq. (12) is negative (-), the deformation is zero" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000318_itec.2018.8450239-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000318_itec.2018.8450239-Figure1-1.png", + "caption": "Fig. 1. Machine design with rotor air cooling system.", + "texts": [ + " Through the rotation of the fins and the static guidance TABLE I MAIN MACHINE PARAMETERS Rated power 30 kW Yoke outer diameter 202mm Rated torque 100Nm Air gap diameter 122mm Max. speed 10 000min\u22121 Shaft inner diameter 82mm Pole pairs 3 Active length 80mm Slots 36 467978-1-5386-3048-8/18/$31.00 \u00a92018 IEEE plates of the tube inside the shaft a speed dependent air flow is achieved. In addition to the cooling of the electrical machine the air flow is used to cool the inverter system, which is mounted circularly around the intake/output tube. The system is illustrated in Fig. 1, the inverter itself is not shown. Further information on the machine can be found in [9]. A rotational electrical machine features a good radial symmetry to the rotational axis, and to some degree a symmetry to the axial center plane depending on the endwindings, shaft and endshield with bearings. In this paper it is taken advantage of these symmetries for simplification of the model structure. The LPTN is shown in fig. 3, and a sketch to illustrate the areas/volumes represented by the distinct nodes is given in fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001960_tmrb.2020.3010611-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001960_tmrb.2020.3010611-Figure2-1.png", + "caption": "Fig. 2. Graphical representation of the sagittal plane of an upper-body exoskeleton and the operator based on the mass-spring-damper model.", + "texts": [ + " The interaction forces can be modeled as a linear mass-spring-damper system, which is in fact also one of the approaches for the modeling of musculoskeletal systems [28]. In this approach, a limited number of masses represent the inertia properties of different segments of the human body including hard tissues and soft tissues. Springs and dampers represent the biomechanical properties of the different segments including bones, muscles, tendons, and ligaments. To model the motion of the human trunk while wearing a back-support exoskeleton, a linear biomechanical model is illustrated in Fig. 2. The equivalent rotational stiffness, including stiffness parameters between the human trunk and the exoskeleton, is represented by a rotational spring keh [N m rad\u22121]. Similarly, the equivalent rotational damping between the human trunk and the exoskeleton is given by a rotational damper deh [N m s rad\u22121]. The damping of the human hip, which corresponds to the rotation axis of the human upper-body, is represented by bh [N m s rad\u22121] and the center of mass locations of the human upper-body and the exoskeleton, relative to their axes of rotation, are given by CMhu [m] and CMexo [m], respectively. It is assumed that the torque is applied to the operator\u2019s upper body (back) through a back-support exoskeleton (Fig. 2). Authorized licensed use limited to: University of Liverpool. Downloaded on July 27,2020 at 17:01:34 UTC from IEEE Xplore. Restrictions apply. 2576-3202 (c) 2020 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. This produces angular displacement \u03b8h [rad] with an angular velocity \u03b8\u0307h [rad s\u22121] and an angular acceleration \u03b8\u0308h [rad s\u22122] of the human upper-body motion, as well as, angular displacement \u03b8exo [rad], angular velocity \u03b8\u0307exo [rad s\u22121] and angular acceleration \u03b8\u0308exo [rad s\u22122] in the upper-body exoskeleton" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002096_j.ifacsc.2021.100138-Figure4-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002096_j.ifacsc.2021.100138-Figure4-1.png", + "caption": "Fig. 4. Description of S\u2013F frame in steering plane.", + "texts": [ + " The control law is designed for the heading otion of AUV by ensuring the L2-gain performance in presence of some bounded perturbations. This approach provides a local solution to the nonlinear control problem for the heading control of AUV. However, when the dimension of state vector increases, the output may be conservative in nature because of the large area of convex hull. 4.2. Path guidance law The design of path following control algorithm based on the concept of Serret\u2013Frenet (S\u2013F) frame which is shown in Fig. 4. The kinematic model of AUV in terms of S\u2013F frame as reported in B\u00f8rhaug (2008) is represented as follows[ x\u0307b/r ] = [ cos(\u03c8rb) \u2212 sin(\u03c8rb) ][ us ] \u2212 [ s\u0307r ] y\u0307b/r sin(\u03c8rb) cos(\u03c8rb) vs 0 \u2212 s\u0307r [ 0 \u2212pc(sr ) pc(sr ) 0 ][ xb/r yb/r ] (25) here \u03c8rb = \u03c8s \u2212\u03c8r , is the steering angle of the body relative to he steering angle of S\u2013F frame, pc(sr ) is the path curvature along he circular path, sr is the arc length and [xb/r , yb/r ]T is the error pace vector between the body and S\u2013F frame along x and y axes. he S\u2013F frame will travel along a desired path in the horizontal lane as shown in Fig. 4. The desired yaw angle for the AUV in the orizontal plane is being found in such a way that it will track S\u2013F rame on the desired path. The error coordinates between body nd S\u2013F frames are exploited below to minimize the convergence f AUV into the path. It is represented as lim \u2192\u221e xb/r = 0, lim t\u2192\u221e yb/r = 0. o realize the path following problem, a guidance and an update aw (B\u00f8rhaug, 2008) are given as D = \u03c8r \u2212 tan\u22121 ( vs us ) \u2212 tan\u22121 \u239b\u239d yb/r\u221a \u25b32 + ( xb/r )2 \u239e\u23a0 (26) s\u0307r = \u221a u2 s + v2s \u239b\u239d \u221a \u25b32 + ( xb/r )2 + xb/r\u221a \u25b32 + ( xb/r )2 + ( yb/r )2 \u239e\u23a0 (27) here \u25b3 is a positive design parameter called as lookahead istance which may be considered as a constant, function of time, rror coordinates or any other parameters" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000769_978-3-662-43893-0-Figure2.12-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000769_978-3-662-43893-0-Figure2.12-1.png", + "caption": "Fig. 2.12 Principle of the tilted root line [ISO23509]. 1 Apex of the pitch cone", + "texts": [ + " Tooth depth and root angle determine the size and shape of the blank. \u2013 The tooth thickness is variable along the face width and is measured on the pitch cone in the transverse or normal section. \u2013 The slot width is dependent on the cutting method and is usually tapered. Only in the cases of the Palloid\u00ae method\u2014a continuous indexing process\u2014and the Completing method\u2014a single indexing process (see Sect. 2.1)\u2014is the slot width constant in the normal section along the face width. In the Completing method, this is achieved by a tilted root line as indicated in Fig. 2.12, again showing straight bevel gear teeth for simplicity. In all other cases, the slot width is determined by the structure of the tool, the point width and the tool tip radius. \u2013 Along the pitch cone of a bevel gear (i.e. in the pitch plane of the crown gear), the space width in the normal section is usually not constant. The exception is the Palloid process which has an involute as the tooth trace and is therefore selfequidistant in every normal section. Because of the tilted root line, the Completing method attains the desired constant slot width but tooth space width is tapered along the pitch cone (see Fig. 2.12). Fig. 2.11 Tooth form variables [ISO23509] 1 tooth depth 2 tooth thickness 3 slot width (crown gear) 4 space width in the pitch plane 26 2 Fundamentals of Bevel Gears Figure 2.13 shows the customary tooth depth variants, or depth wise taper, which are described briefly below. The formulae for the associated angles are given in Tables 2.12 and 2.13. Standard depth tooth depth is directly proportional to the cone distance at any particular section along the tooth. The extended root line intersects the axis of the bevel gear at the pitch cone apex", + " The extended tip line intersects the axis at a different point, defined by the root line of the mating gear plus a constant clearance. The sum of the dedendum angles of the pinion and wheel is not dependent on the tool radius. Most straight bevel gears are of the standard depth type. Constant slot width (Duplex) this tooth depth form occurs when the root line has to be tilted as required in the case of the Completing method, to obtain a constant slot width in the normal section of the pinion and wheel (see Fig. 2.12). The formulae in Tables 2.12 and 2.13 indicate that the tool radius rc0 has a significant 2.2 Gear Geometry 27 effect on the tilt angle of the root line. Too large a tool radius being produces unreasonably small tooth depth at the toe and excessive tooth depth at the heel. As a result, tooth tips become too thin at the heel and there is a danger of undercutting at the root. It is therefore recommended that the tool radius rc0 should be no larger than the mean cone distance of the wheel Rm2. If the tool radius is too small, the opposite effect occurs, and therefore the selected value should not be lower than 1" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001535_icems.2019.8921448-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001535_icems.2019.8921448-Figure3-1.png", + "caption": "Fig. 3. The distribution of magnetic flux density under no-load condition.", + "texts": [ + " The amplitude of the cogging torque for the prototype motor with straight slot-opening is 82.3 mNm. And after the stepped slot-opening shift, the amplitudes of the cogging torque are reduced to about 21.4, 18.9 and 8.7 mNm respectively. As shown in Fig. 2(b), with the proposed stepped slotopening shift method employed, the fundamental and 2nd order components of the cogging torque can be effectively reduced for the prototype motor. But the 3rd order harmonic component increases slightly for 3 steps shift scheme, which is a minor shortcoming of the proposed method. As shown in Fig. 3, the distribution of magnetic flux density under no-load condition has been investigated by 2D FEM. It can be seen that due to the local magnetic saturation in the rotor core, (2) cannot accurately express the cogging torque with stator slot-opening shift effects. Furthermore, since the stator slot width is limited, the outermost slotopening steps are shifted to the stator teeth when the number of steps exceeds 4 for the prototype. Therefore, two different optimization measures are proposed to correct the proposed method and reduce the cogging torque further in this section" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000943_1350650115611155-Figure4-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000943_1350650115611155-Figure4-1.png", + "caption": "Figure 4. Inner ring displacement under combined loads.", + "texts": [ + " Then the resultant moment along with y-axis is X Miy \u00bc Xn i\u00bc1 Rp \u00f0Fi F\u00de sin 2 n \u00f0i 1\u00de \u00f03\u00de the resultant moment along with z-axis is X Miz \u00bc Xn i\u00bc1 Rp \u00f0Fi F\u00de cos 2 n \u00f0i 1\u00de \u00f04\u00de The 5-DOF quasi-static model When the ball bearing is working under the combined loads of an axial force Fx,, two radial forces Fy, Fz and two momentsMx,Mz, there will be a relative displacement between the outer ring and inner ring, which can be expressed as the axial displacement x, radial displacement y, z and angular displacement y, z for the inner ring (see Figure 4). And the angular position of each ball in bearing is shown in Figure 5. Assuming that the number of balls in bearing is Z and the contact force acting on ball in position angle j is Q j, then the force balance equations of inner ring are18 Fx \u00bc Xj\u00bcZ j\u00bc1 Q j sin \u00f05\u00de Fy \u00bc Xj\u00bcZ j\u00bc1 Q j cos j cos \u00f06\u00de Fz \u00bc Xj\u00bcZ j\u00bc0 Q j sin j cos \u00f07\u00de My \u00bc Xj\u00bcZ j\u00bc1 di 2 Q j sin j cos \u00f08\u00de Mz \u00bc Xj\u00bcZ j\u00bc1 di 2 Q j cos j cos \u00f09\u00de where di \u00bc dm D 2 cos In Figure 6, for a single ball with an azimuth of j under an applied static load, the action line between the inner and outer raceway groove curvature centers is OiOe" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000893_j.ymssp.2015.10.028-Figure11-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000893_j.ymssp.2015.10.028-Figure11-1.png", + "caption": "Fig. 11. Representation of the mesh in the area of the modeled teeth of the second model.", + "texts": [ + " In the employed gearbox bolted joints connect both sides of the ring gear with the supporting gearbox structure. Each bolt on every side reaches to roughly one quarter of the depth of the ring gear while the tapped hole reaches to one third of the depth. There are no through holes. However, a potential influence of these holes should be investigated to study the sensitivity of various measurement positions in relation to the ring gear width. Hence, in a second model additional drill holes of the bolted joints at the three modeled tooth meshes, as shown for example in Fig. 9, are modeled (Fig. 11). These holes reach from both sides to one third of the ring gear depth. The bolts are not modeled since the stiffness of the bolt connection in radial direction is unknown. Hence, the model underestimates this stiffness and the real results will be between both extremes represented by no holes and holes to one third of the depth. The contact boundary-conditions between the ring gear and the gearbox housing are again realized with tie-constraints. The diagram in Fig. 12 shows a 90\u00b0-section of the simulation results of both models" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000302_b978-0-12-813691-1.00006-3-Figure6.23-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000302_b978-0-12-813691-1.00006-3-Figure6.23-1.png", + "caption": "FIGURE 6.23", + "texts": [ + " RDEs are mostly used in hydrodynamic techniques, for example, hydrodynamic amperometry or voltammetry. RDEs can\u2019t be applicable for reversal techniques because the product of the electrochemical reaction is continuously swept away from the surface of the disk A schematic representation of TCO and GO/FTO. 1936.3 Electrochemistry, Electrochemical Cell and Techniques (Bard and Faulkner, 2001). Therefore, a concentric-ring electrode can be added to surround the disk, which produces the rotating ring-disk electrode (RRDE) (Fig. 6.23). Since in RRDE experiments, two potentials (that of the disk, ED, and that of the ring, ER) and two currents (iD and iR) are measured and reported. Moreover, the ring itself can be used alone as an electrode, which is called the rotating ring electrode (RRE). There are only three main sources of analytical signals, which include potential, current and charge (Wang, 2006; Harvey, 2000). However, a wide variety of experimental designs can be utilized. There is a simple division, consisting of (1) bulk methods, measuring properties of the solution (e" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000804_s40435-015-0203-0-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000804_s40435-015-0203-0-Figure1-1.png", + "caption": "Fig. 1 a Aeropendulum (photo), b schematic physical model", + "texts": [ + " The estimation of stability is performed through a linearization of the system that, in spite of its simplicity, is effective for the purposes of this work. In the last part of the paper, stability charts obtained with different controllers (continuous time, standard discrete time, AaW) are compared, showing the larger stability region obtainable under proper conditions with the AaW controller, as predicted by the theory. The system under study consists of a light carbon rod, connected perpendicularly to the shaft of a low-friction potentiometer (Fig. 1). At one of its edge, the rod is rigidly connected with a small DC electric motor, driven by a 5-V pulse-width modulated (PWM) signal. The potentiometer is fixed to a plastic frame, such that the rod and the motor form a rigid planar pendulum, freely rotating around the axis of the potentiometer. The motor drives a propeller, able to give a control force to the pendulum, linearly proportional to the PWM signal within a certain range. The system is equipped with a custom-designed circuit board, which controls the voltage supplied to the motor with a resolution of 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000925_s11044-016-9500-4-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000925_s11044-016-9500-4-Figure2-1.png", + "caption": "Fig. 2 Schematic diagram of a general passive overconstrained PM", + "texts": [ + " For example, the 3-PRRR orthogonal 3-DOF translational overconstrained PM [35], each supporting limb of which supplies an actuated force and two constraint couples to the moving platform. In the following, we will investigate whether the weighted Moore\u2013Penrose generalized inverse can be directly used to the force analysis of the general passive overconstrained PMs. The schematic diagram of a general passive overconstrained PM that consists of a moving platform, a fixed base and t supporting limbs is shown in Fig. 2. Assume that the ith limb supplies Ni (i = 1,2, . . . , t ) constraint wrenches (including the actuation wrenches, which are not reciprocal to the twist of the driving joint, but are reciprocal to all the twists of the other joints within the ith limb [36]) to the moving platform, as denoted by /S i j (j = 1,2, . . . ,Ni ), respectively. For a general passive overconstrained PM, the inequality N1 + N2 + \u00b7 \u00b7 \u00b7 + Nt > 6 is always satisfied. For convenience of analysis, a reference coordinate system O-XYZ is attached at a central point O on the moving platform" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001925_978-3-030-46817-0_30-Figure25-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001925_978-3-030-46817-0_30-Figure25-1.png", + "caption": "Fig. 25. Stress intermediate principal validation of FEA model.", + "texts": [], + "surrounding_texts": [ + "Figs. 23, 24, 25 and 26. 142 I. Alagi\u0107 In case of ball joint, the results of stress distribution and allowed displacement are presented in Figs. 27, 28, 29, 30, 31 and 32. Finite Element Analysis (FEA) of Automotive Parts Design as Important Issue 143 The displacement achieved as a result of finite element analysis doesn\u2019t much differ from results of the laboratory-test performed by control device MR 96. The maximum displacement appeared into Z direction 0,145 [mm]. On the basis of conducted simulations were possible to affirm that the magnitude of deformations depends on model geometry. The largest concentration of stresses appeared in places near of cover of ball joint. Through change of model geometry it was possible to influence on expansion of stresses and displacement distribution. To analysis of maximum warp angle of ball joint using FEA solver was very low compared to its allowed value 58 \u00b1 6\u00b0." + ] + }, + { + "image_filename": "designv10_14_0001761_012028-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001761_012028-Figure1-1.png", + "caption": "Figure 1. Sectioned view of the AC homopolar motor/generator", + "texts": [ + " This paper describes 2 MW, 25000 RPM concept designs for machines employing an HTS field excitation winding energized by a DC dynamo. Table 1 summarises the specifications for the 2 MW 25,000 RPM motor excited with a DC dynamo. k) Stator laminations are 0.1 mm thick Japanese JNEX-Core (model 10JNEX900) The design and analysis for such machines is based on reference [15]. A similar machine was previously designed for a flywheel energy storage system [16] but did not incorporate excitation by a dynamo. A cutaway diagram of the AC homopolar motor/generator is shown in Figure 1. The shaft and rotor have not been sectioned so as to show the 6-pole layout, with 60\u00b0 rotationally offset poles. The three armature coil colors illustrate the three phases winding scheme. The HTS coil cryostat and its ICMC 2019 IOP Conf. Series: Materials Science and Engineering 756 (2020) 012028 IOP Publishing doi:10.1088/1757-899X/756/1/012028 surrounding insulation have been omitted for clarity, as has an electromagnetic (EM) shield located between the armatures and the superconducting coils" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002015_j.asr.2020.09.040-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002015_j.asr.2020.09.040-Figure2-1.png", + "caption": "Fig. 2. Layout of the spiral folding pattern on a flat membrane and its folding process.", + "texts": [ + " Generally speaking, given a flat membrane and a regular polygon positioned in its core (or hub), creases on the flat membrane are tangential to the core polygon, and the fold behaviour transforms the flat membrane into a hollow cylinder having a regular polygon in its core. Along the transition process from flat to cylinder configuration, zigzag patterns are generated in the cylindrical shape which is due to the shape and configuration of the mountain and valley folds of the membrane, as show in Fig. 1. An example of the layout structure of the spiral fold pattern is shown by Fig. 2(a), and an example of a wrapped configuration using paper is shown by Fig. 2(b). In order to portray the governing equations of the spiral folding pattern, we first refer to Fig. 3, which describes the key elements involved in the rendering of the crease geometry and the overall folding structure. By observing Fig. 3, given a regular polygon with center at origin O, the crease Fold A is tangential to the edge of the core polygon. Considering a regular polygon with N sides, the spiral folding pattern is expected to have N creases of the type Fold A. Then, let P be an arbitrary point in Fold A, it is possible to describe an infinitesimal region bounded by an inner polygon with apothem R and an outer polygon with apothem R\u00fe dR, in polar coordinates, as shown by Fig", + " Since it is difficult to predict where the buckling effect is to occur, for simplicity we use reasonably small holes at the corners of creases (Fig. 27(c)) to avoid the kind of buckling effect due to winding, thus enabling to make a compact folding while ensuring the virtual thickness to be within the upper bound t (Fig. 27(d)). In line with the above-mentioned concepts, we evaluated the following types of planar layouts, as shown by Fig. 28: Model 1. Spiral fold with t \u00bc 0:1 mm.; this model is inspired by the conventional formulation of the spiral folding with planar membrane as shown by Fig. 2\u20136. Model 2. Multiple spiral folding, with t \u00bc 0:1 mm.; this model takes into account the multi-spiral folding pattern of the above layout, in which superimposition of the clockwise and counterclockwise orientations is realized by two concentric spirals. Model 3. Spiral fold, with t \u00bc 1 mm.; this model considers the formulation of archimedean-type spiral fold segments originating at (being normal to) vertices (edges) of the polygon hub(Guest and Pellegrino, 1992; Nojima, 2001a, 2007; Zirbel et al" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000213_s12206-017-0202-5-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000213_s12206-017-0202-5-Figure1-1.png", + "caption": "Fig. 1. Fault and degradation experiment of the planetary gear.", + "texts": [ + " (19) and the eigenvalue is from small to large, then if the eigenvectors that correspond to eigenvalue 0 are removed, then the next d eigenvectors 1 2, , , df f fL are the best projection results, as follows: 1 2( ) ( ( ), ( ), , ( )) .dX i f i f i f i\u00ae L (20) Planetary gear fault and degradation experiments were simulated in the DDS mechanical fault comprehensive simulation bench made by Spectra Quest Company in the USA. The experimental system is composed of a two-stage planetary gearbox, a programmable control motor, a fixed-axis gearbox, a programmable magnetic brake component, a data acquisi- tion system and a portable computer. The simulation bench is shown in Fig. 1. All types of gear vibration signals are measured by acceleration sensors, the layout of which is shown in Fig. 2. The proposed method is verified by the measured vibration signal, and the analysis flowchart of the proposed method based on DT-CWT threshold denoising and LE is shown in Fig. 3. Given that the sun gear meshes with the planet gears several times in a revolution and works at a higher speed, it is prone to fault. Therefore, in the experiment, the fault and degradation of sun gear are selected as the examples to simulate" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000769_978-3-662-43893-0-Figure3.11-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000769_978-3-662-43893-0-Figure3.11-1.png", + "caption": "Fig. 3.11 Flanks of tooth pair k in meshing position \u03c61 [BAER91]", + "texts": [ + " What is simulated is the load-free meshing of the tooth flanks (described, for example, by fitting surfaces) of a tooth pair at a theoretical transmission ratio (not including the subsequently calculated transmission error). Mating tooth flanks are separated by distances which vary for each meshing position i at a given meshing interval expressed by means of the pinion rotation angle \u03c61i or the wheel rotation angle \u03c62i. In meshing position i, arc distance \u03b6 (r2, \u03c52, \u03c61, k) exists between wheel flank 2 on tooth k, given by variables r2, \u03d12,, and pinion flank 1 (Fig. 3.11). The function of variables r2, \u03d12 per meshing position i is termed the (instantaneous) ease-off function of \u03c61 and k. The enveloping surface over all instantaneous ease-offs represents the minimum of all gape distances during one complete meshing cycle of a tooth pair. This distance is termed the contact distance or ease-off. 3.3 Tooth Contact Analysis 75 This approach is by no means restricted to bevel gears but is also suitable to describe the contact parameters of any three-dimensional surfaces in rolling contact" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001765_ffe.13295-Figure4-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001765_ffe.13295-Figure4-1.png", + "caption": "FIGURE 4 Geometry of the fatigue specimens (dimensions in mm)", + "texts": [ + " The mean average roughness (Ra) of the tensile samples in the \u201cAs built\u201d and \u201cMachined\u201d configuration was 25.8 and 2.8 \u03bcm, respectively. The tensile tests were performed by using an MTS 641 Hydraulic Wedge Grips with load cell of 250 kN, in displacement control with a rate of separation of the two heads of the testing machine during the test equal to 2 mm/min. The strain was measured by a contact-type extensometer directly connected to the specimen. Experimental tests were performed using dog-bone round specimens, with final geometry and dimensions compliant to the standard ASTM E46631 shown in Figure 4. To limit the number of variables that can affect the mechanical response of the specimens, in this work, a single grow orientation and the same set of parameters for all specimens were considered. In particular, according to some research works,19,32 the samples were tested in the \u201cweak\u201d z direction, i.e., perpendicular to the layers, to get minimum fatigue properties. Three batches of specimens were produced with 90 grow orientation with respect to the start plate in the x-y plane (Figure 2B)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000894_tro.2015.2489499-Figure19-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000894_tro.2015.2489499-Figure19-1.png", + "caption": "Fig. 19. Hidden robots involved in the tested visual servoings of the 3\u2013PRR robot. (a) When all leg directions ui are observed (Case 1): a 3\u2013PRR robot. (b) When all leg directions ui and the Plu\u0308cker coordinates of the line passing through the leg 1 are observed (Case 2): a PRR\u2013{2\u2013PRR} robot. (c) When all leg directions ui and the Plu\u0308cker coordinates of the lines passing through the legs 1 and 3 are observed (Case 3): a PRR\u2013{2\u2013PRR} robot.", + "texts": [ + " 18(b)]: They may appear for any x and y if and only if the robot reaches constant platform orientations defined by cos \u03c6 = a2/(d + l) or cos \u03c6 = a2/ |d \u2212 l|. 2) Analysis of the Possible Hidden Robot Models: Case 1: Let us now assume that we want to control the 3\u2013PRR robot depicted at Fig. 17(a) by using the observation of its leg directions ui (see Section II). From Section III, we know that using such a control approach involves the appearance of a hidden robot model. This hidden robot model can be found by straightforwardly using the results of Section III and is a 3\u2013PRR robot shown in Fig. 19(a). This robot is known to be architecturally singular (it can freely move along the y-axis) and cannot be controlled by using only the observation of its leg directions ui . Case 2: As a result, one would logically wonder what should be the necessary information to retain in the controller to servo the robot. By using the results of Section IV-D, we know that, from the projection of the cylindrical leg in the image plane, it is not only possible to estimate the leg direction, but also the Plu\u0308cker coordinates of the line passing through the axis of the cylinder, i.e., the direction and location in space of this line. Let us consider that we add this information for the estimation of the leg 1 position only. Modifying the hidden robot model according to Fig. 16(a), the corresponding robot model hidden in the controller is depicted in Fig. 19(b): This is a PRR\u2013{2\u2013 PRR} robot that is not architecturally singular. In other words, using the Plu\u0308cker coordinates of the line for leg 1 involves to actuate both the first P and R joints of the corresponding leg, i.e., the virtual leg is a PRR leg. For the PRR\u2013{2\u2013PRR} robot, it is possible to prove that two assembly modes exist which are separated by a Type 2 singularity at \u03c6 = 0 or \u03c0 (for any x and y). For both assembly modes, the end-effector position is the same, while the orientation is different", + " Case 3: From the result that, using the Plu\u0308cker coordinates of the line passing through the axis of the cylinder, the leg of the virtual robot becomes a PRR leg; it is possible to understand what is the minimal set of information to provide to the controller to fully control the robot in the whole workspace: we need to use the Plu\u0308cker coordinates of the lines passing through legs 1 and 3 and the direction of the leg 2. In such a case, the hidden robot model is a PRR\u2013{2\u2013PRR} robot depicted in Fig. 19(c). It is possible to prove that this robot has no Type 2 singularity and can freely access its whole workspace. 3) Simulation Results: Simulations are performed on an Adams mockup of the 3\u2013PRR robot with the following values for the geometric parameters: l = 1 m, d = 0.4 m, a1 = 0.4 m, and a2 = 0.25 m. This virtual mockup is connected to MATLAB/Simulink via the module Adams/Controls. The controller presented in Section II is applied with a value of \u03bb assigned to 20. The initial configuration of the robot end-effector is x0 = 0" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000233_hsi.2017.8005020-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000233_hsi.2017.8005020-Figure2-1.png", + "caption": "Fig. 2. The geometry\u2019s 2-DOF robot manipulator.", + "texts": [ + " SIMULATION Simulation studies were conducted on Matlab-Simulink and the mechanical model of the 2-link robot manipulator to verify the Proposed scheme control. Its dynamic parameters are considered into Table 1. Links Parameter of each link Length (m) (m) Weight (kg) Link inertia ( ) Link 1 = 1 = 1 = 1 = 0.8 Link 2 = 1 = 1 = 1 = 0.8 Table. 1. The parameter of the robot manipulator. The mechanical model of the 2-link robot manipulator is designed by SimMechanics on Matlab-Simulink. The geometry\u2019s 2-DOF robot manipulator is illustrated in Figure 2. The dynamic of the robot is described as in (1), where Inertia term ( ) = + ( + + 2 ( )) + + ( ) = ( ) = + ( ) + ( ) = + Velocity and Coriolis term (1) = \u22122 ( ) \u2212 ( ) (2) = ( ) Gravitational term (1) = ( ) + ( ( ) + ( + )) (2) = ( + ) The friction ( ) and disturbance are assumed as (1) = 1 + 3 (3 ) (2) = 1.1 + 4 (2 ) (1) = 0.2 ( ) (2) = 0.1 ( ) The desired trajectory is given by = [ 5 \u2212 1; 5 + 2] The parameter of controller = (55,55), =(10,10) , the sliding surface parameter is chosen as =(10,10)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000769_978-3-662-43893-0-Figure2.1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000769_978-3-662-43893-0-Figure2.1-1.png", + "caption": "Fig. 2.1 Bevel gears with variable and constant tooth depth", + "texts": [ + " With constant tooth depth, the face and root angles are of equal value, such that the depth of the tooth remains the same over the entire face width. On bevel gears with variable tooth depth, also known as tapered teeth, the face and root angles differ, causing a proportional change in tooth depth along the face width. At the small diameter of the bevel gear (toe), tooth depth is less than that at the large diameter (heel). Constant tooth depth may be regarded as a special case of tapered teeth (Fig. 2.1). \u00a9 Springer-Verlag Berlin Heidelberg 2016 J. Klingelnberg (ed.), Bevel Gear, DOI 10.1007/978-3-662-43893-0_2 11 Other bevel gear criteria are the type and form of the tooth trace on the basic crown gear (see Sect. 2.2.2). Depending on the type of tooth trace, bevel gears may be differentiated according to Fig. 2.2 into: \u2013 straight bevel gears \u2013 skew bevel gears \u2013 spiral bevel gears On spiral bevel gears, it is possible to draw a further distinction in terms of the form of the tooth trace, which may be: \u2013 a circular arc, \u2013 an elongated epicycloid, \u2013 an involute or \u2013 an elongated hypocycloid Bevel gears may likewise be classified with respect to their hypoid offset" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000203_1.g002222-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000203_1.g002222-Figure1-1.png", + "caption": "Fig. 1 Schematic of flexible aircraft frames.", + "texts": [ + " Section IV presents themain results that show how the LPV FTC controller can be synthesized by an LMI formulation. Simulation results are given to illustrate the effectiveness of the proposed design in Sec. V. Section VI provides concluding remarks. A flexible aircraft model that can describe the rigid-body motions of the aircraft and the relatively small elastic deformations of the flexible wings, as well as the coupling between them, is briefly introduced in this section. To describe the equations of motion of flexible aircraft, three frames are defined, as shown in Fig. 1: the inertial frame, the fuselage frame attached to the undeformed fuselage, and the wing frame attached to the undeformed wing. The six rigid-body degrees of freedom are the translations Rf and rotations \u03b8f of the fuselage frame, and the elastic deformation is defined as the displacement of each point on the wing relative to the wing frame uw. The translational and angular velocity vectors of the fuselage frame and D ow nl oa de d by U N IV O F C A L IF O R N IA S A N D IE G O o n Ja nu ar y 24 , 2 01 7 | h ttp :// ar c" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002033_j.jmatprotec.2020.116973-Figure5-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002033_j.jmatprotec.2020.116973-Figure5-1.png", + "caption": "Fig. 5. Oblique cutting model for calculation of effective rake angle.", + "texts": [ + " Journal of Materials Processing Tech. 290 (2021) 116973 the cutting edge can be calculated from Sc(u,v), as follows: Vc(t) = \u2202Sc(u, v)/\u2202t To calculate the depth of the crater wear KT, the effective rake angle \u03b3ei on the rake face must be clarified during a single cut. From Eq. (8), it is evident that the tangential vector nt(u) is not decisively perpendicular to Vc; thus, we can simplify the skiving process as a serial oblique cutting scenario for cutting edge points at one moment. As defined in Fig. 5, the effective rake angle \u03b3ei for a cutting point on the tool edge can be calculated using the method proposed by Usui and Masuko (1973), as follows: \u03b3ei = sin\u2212 1(sin\u03b7csin\u03b1b + cos\u03b1bsin\u03b3cos\u03b7c), (9) where \u03b7c is the chip discharge angle, and can be obtained from equation proposed by Kaneko et al. (2017) as follows: \u03b7c = tan\u2212 1(tan(\u03b1b)/(sin(\u03b1) + cos(\u03b1)) ), (10) where \u03b1b is the tool inclination angle, and can be expressed as follows: \u03b1b = \u03c0/2 \u2212 cos\u2212 1(vc(t)\u2219nt(u)/(\u2016vc(t)\u2016\u2016nt(u)\u2016)). (11) To obtain the temperature and stress distributions on the rake face for calculating crater wear, the oblique cutting model in Fig. 5 was converted to a 2D-orthogonal cutting simulation by projecting the effective rake angle, depth-of-cut, and cutting velocity to a plan defined by the chip flow direction and normal vector of the rake face nr. Fig. 6(a) shows the extracted cutting characteristics for establishing the 2D-FEM model shown in Fig. 6(b) by AdvantEdge\u2122 (Third Wave). A database was created for the stress and temperature distributions under various machining conditions. The temperature distribution in quasi-steady state was employed for each machining condition" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001212_tpel.2019.2906431-Figure4-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001212_tpel.2019.2906431-Figure4-1.png", + "caption": "Fig. 4. Winding circuit diagram for parallel operation of 1-phase (left) and 3-phase motor (right).", + "texts": [ + " The voltage equation in the stationary reference frame is given as follows: , , , , , , , , , , , , sin cos s s s ds k s k ds k s k ds k r f k r s s s qs k s k qs k s k qs k r f k r d v R i L i dt d v R i L i dt (1) where Rs,k , Ls,k, and \u03bbf,k denote the stator resistance, the stator synchronous-inductance per phase, and the flux linkage of the rotor, respectively. id,k and iq,k are the d- and q-axis currents, and vd,k and vq,k are the d- and q-axis voltages. Superscript s denotes the stationary reference frame. Subscript k represents the 3-phase and the 1-phase motor with 3\u03d5 and 1\u03d5, respectively. \u03c9r and \u03b8r refer to the electrical rotating speed and rotor angle. Because the healthy and faulty 3-phase motor are connected in parallel, the q-axis voltages applied to each motor are the same, as shown in Fig. 4. Applying Kirchhoff\u2019s voltage law to the left loop in the 1-phase motor, the steady state current of each phase can be derived as ,1 ,1 ,1 ,1 ,1 3 2 s s s qs bs cs bs s v e e i R , ,1 ,1cs bsi i (2a) where ias, ibs, and ics are the phase currents of a-, b-, and c-phase, respectively. In addition, eas, ebs, and ecs represent the back EMFs of the a-, b-, and c-phase in the stationary reference frame, respectively. In the same manner, the phase current in the upper left and lower left loop of the 3-phase motor are given below: ,3 ,3 ,3 ,3 ,3 1 1 3 2 2 s s s bs ds qs bs s i v v e R , ,3 ,3 ,3 ,3 ,3 1 1 3 2 2 s s s cs ds qs cs s i v v e R . (2b) Because the healthy and faulty 3-phase motor are connected in parallel, the q-axis voltages applied to each motor are the same, as shown in Fig. 4. Applying Kirchhoff\u2019s voltage law to the left loop in the 1-phase motor, the steady state current of each phase can be derived as Under the assumption that the harmonic current of high order term and the leakage current are negligible, the d- and q-axis stator currents can be formulated as 1 1/ 2 1/ 22 3 0 3 / 2 3 / 2 ass ds bss qs cs i i i i i . (3) From (3) and (3), the d- and q-axis stator currents flowing through each motor are determined by each phase currents, as shown in the following equations" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001851_jsen.2020.3043999-Figure4-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001851_jsen.2020.3043999-Figure4-1.png", + "caption": "Fig. 4. Deformed rod. The coordinate of the rod\u2019s centerline at arc length s is described in world frame by vector p(s). At this point, external forces n(s) and torques m(s) act on the rod.", + "texts": [ + " We then compare our strain curvature model to two other approaches commonly found in the literature. All algorithms are implemented in Matlab 2019b. Systems of ordinary differential equations are solved with Runge-Kutta (4,5) as implemented in Matlab\u2019s ode45 function. To apply the shooting method presented in section III-A, we rely on the trust-region dogleg optimization algorithm as implemented in Matlab\u2019s fsolve function. A. Cosserat Rod Model In this section we give a brief overview of the kinematics of a simple Cosserat rod with a circular cross-section as depicted in Fig. 4. A more detailed discussion can be found in [30]. The rod shape is described with respect to the arc length s \u2208 [0, L] by its centerline p(s) \u2208 R 3 and a corresponding material frame with the orientation R(s) = [e1, e2, e3] \u2208 SO(3). The vector e3 is coaxial to the rod centerline\u2019s tangent vector. Furthermore, u(s) and v(s) are the angular and translational rate of change of this frame with respect to s. The initial straight rod is deformed by distinct forces n(s) and torques m(s) which act on the rod at s (distributed loads are neglected)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000750_j.mechmachtheory.2018.11.002-Figure9-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000750_j.mechmachtheory.2018.11.002-Figure9-1.png", + "caption": "Fig. 9. Surface deviations for: (a) Case 3, and (b) Case 4.", + "texts": [ + " 8 (b) shows the surface deviations for Case 2 for which the cutter profile tilt angle has been increased to 35 \u00b0, and the cutter radius has been kept equal to that of Case 1. Considering the new cutter profile tilt angle of 35 \u00b0, the deviations at the bottom boundary of the active surface of the face gear have been reduced to \u22124.0 \u03bcm and \u22122.33 \u03bcm at the inner and outer sections, respectively. However, deviations at the top edge of the face gear tooth surfaces have been increased to \u221278.06 \u03bcm at the inner section and to \u221260.02 \u03bcm at the outer section, and with it, the effective longitudinal crowning of the face gear tooth surfaces has been increased. Fig. 9 (a) shows the surface deviations for Case 3. This case uses a larger cutter radius (600 mm) compared with Case 2 that causes the amount of longitudinal crowning applied to the face gear tooth surfaces to be decreased. Deviations of -88.67 \u03bcm at the top edge of the inner radius section and \u221230.19 \u03bcm at the outer radius section were found. Deviations at the bottom boundary of the active surface are of \u22121.42 \u03bcm at the outer section and \u22122.13 \u03bcm at the inner section. Case 4 uses a parabolic profile for the rack-cutter of the shaper and introduces profile crowning for the profile of the reference shaper that is used as a reference profile for the circular cutters. Surface deviations for Case 4 are shown in Fig. 9 (b). In this case, more material is left at the top and the bottom parts of the shaper profile than for the involute profile and therefore, more material will be removed from the face gear tooth surfaces at the top and bottom portions of the active tooth surface. Compared to Case 1, the positive deviations on the bottom boundary of the active surface for the inner and outer sections are reduced to about 3.42 \u03bcm and 4.83 \u03bcm which are very close to the results obtained in Case 3. At the same time, the maximum deviations on the top boundary reach about \u221282.8 \u03bcm and \u221239.75 \u03bcm for the inner and outer sections, respectively. Finally, Fig. 10 shows the surface deviations for Case 5. Here, toprem-type profiles are applied for the generating profile of the shaper together with a parabola coefficient of 0.0015 1/mm. It can be seen from Fig. 10 that only the material of the flank near the bottom part was removed gradually compared with Case 4 shown in Fig. 9 (b). The deviations at the bottom of the inner and outer sections are of -35.98 \u03bcm and -34.71 \u03bcm, respectively, and this value can be adjusted by the parabolic coefficient of the toprem portion. Fig. 11 shows the maximum deviations obtained at the top and bottom boundaries of the active surface of the face gear when it is generated by circular cutters with mean cutter radii varying from 100 mm to 10 0 0 mm. A constant cutter profile tilt angle \u03b4 = 30 \u25e6 was considered for all cases investigated. The absolute values of the obtained deviations for both the top and bottom boundaries of the active surface are reduced with the increment of the mean cutter radius" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001597_s11012-019-00996-3-Figure7-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001597_s11012-019-00996-3-Figure7-1.png", + "caption": "Fig. 7 Finite element analysis model of the bearing pairs", + "texts": [ + " The finite element analysis (FEA) model was established by using the hexahedral element (SOLID186) to mesh the simplified model. In order to obtain accurate simulation results, the contact regions between the balls and races were meshed more refined. In this contact model, the contact pressure was calculated for the node on the contact element and the augmented Lagrangian formulation was used to solve the nonlinear contact problem. The FEA model consists of 2698 solid elements, 3048 contact elements and 10958 nodes. Figure 7 shows the FEA model of the bearing pairs. As demonstrated, the external load Fa was applied on a facet of the left inner race through a uniform pressure. The preload FP was applied by decreasing the clearance between the left and right outer races so as to increase the compressive deformations of the balls. The constraint conditions of the bearing pairs were simulated by fixing the outer races and restricting the movements of the inner races along the y and z directions. The wear simulation flow chart of the bearing pairs is shown in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000963_physreve.94.013002-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000963_physreve.94.013002-Figure1-1.png", + "caption": "FIG. 1. The planar confined e-cone. (a) A thin disk of paper card is cut radially and splayed open onto a flat surface. The deformed shape is held down by adhesive tape along the open edges. (b, c) Finite element analysis using the commercial software package ABAQUS [14]. A split disk made of elastic S4R5 shell elements sits slightly above a larger, rigid disk, which performs as the underlying surface. A linear eigenvalue analysis of the elastic disk is performed to generate a stress-free transverse imperfection of 0.1%, in order to seed out-of-plane buckling as the split is opened, which is just visible. Zero friction contact between the two surfaces prevents interpenetration as the buckle moves away from the rigid base with differently colored height contours; in panel (c), the outline wedge has been added to encapsulate transverse displacements larger than the initial imperfection, and subtends approximately 120\u25e6. (d) A discretely folded version of (a) using three fold lines symmetrically located with respect to the radial cut. The central fold line is a mountain fold, the other two are valley folds.", + "texts": [ + " INTRODUCTION Surface disclinations are associated with a rapid change in the orientation, or rotation, of a surface, usually along a straight line; a dislocation occurs when there is a jump in positional order [1]. Disclinations are typically formed by first adding or subtracting material from within an originally flat surface. If the surface is very thin, this change in material content is accommodated by elastic, out-of-plane bending in preference to in-plane straining according to Gauss\u2019s Theorema Egregium [2]: and, in structural terms, equilibrium must be satisfied in this deformed or buckled state while respecting the support conditions. Figure 1 shows how this behavior can be achieved using a disk of ordinary paper card. The disk has been cut radially toward its center before being splayed into a flat, open wedge of around 15\u25e6. Later, we insert extra card for continuity, to form a positive disclination, but we note for now that the split edges have to be held down by adhesive tape, Fig. 1(a). Approximately two-thirds of the disk lie naturally flat without surface fixity; the remainder has lifted off to form a distinctly conical vertex connected to the flat parts by narrow transition zones on each side. As the wedge angle is manually increased, the cone becomes taller but not wider, and the angle subtended around the vertex rises above 360\u25e6. Such angular excess connotes the term \u201ce-cone\u201d for this shape [3]. For much larger wedge angles, the width of the cone also changes as more of the disk detaches from the underlying surface because of increasing geometrical nonlinearity: although interesting, this regime is beyond the scope of study", + " Figures 1(b) and 1(c) show a repeat of the card experiment using a geometrically nonlinear finite element analysis. Its *kas14@cam.ac.uk details are described in the caption and contours of transverse displacement are shown above the level of imperfection needed to induce buckling, which evidently include the transition zones. The conical width is clearly in line with the 120\u25e6 from our crude card experiment; analytical studies that consider a perfectly flat disk initially, suggest angles closer to 180\u25e6 [5]. The shape in Fig. 1(a) is mostly developable because the paper card is thin, but close to the vertex the surface must be doubly and negatively curved because Gauss\u2019s theorem tells us that the angular excess imposed at the vertex produces an exact amount of Gaussian curvature. If we then imagine the card thickness being reduced to zero, the Gaussian curvature becomes concentrated at the vertex tip as a \u201cpoint charge.\u201d The same concentration occurs when the disk is folded instead along three radial lines, in order to emulate the e-cone in a discrete manner; see Fig. 1(d). These lines are symmetrically located with respect to the radial cut: along the tallest meridian and where the e-cone of Fig. 1(a) begins to lift off on either side. When the wedge angle is opened horizontally, adjacent plate regions slide on the surface as per the continuum case. The other plates move upwards under fold line rotations to form approximately the original conically shaped buckle. The meridian now forms as a ridge, or mountain fold, and the lift-off lines as valley folds. Assuming that the vertex is perfect and that the plates, or facets, remain flat, the folded shape is described entirely by the fold line rotations and the wedge angle. This is a simpler kinematical specification, cf. the continuum case, without detriment to the accuracy of shape: both e-cones from Fig. 1 have roughly the same conical widths and wedge angles, with very similar proportions of shape. Opening the wedge angle right up to when the cone facets become vertical and contact each other, and closing it back to zero, requires little effort when the fold lines are repeatedly flexed for compliancy. The shape, however, becomes more difficult to fold when the valley folds in plan are separated by nearly 180\u25e6. At this value and beyond, the disk \u201clocks\u201d and the wedge cannot be opened. This locking is not due to the material stiffening but to inadmissible fold line rotations, as will be shown", + " When they are separated by more than 180\u25e6, folding can proceed only when disk is made to overlap at the cut\u2014when angular deficit is imposed at the vertex, and this 2470-0045/2016/94(1)/013002(8) 013002-1 \u00a92016 American Physical Society case is treated momentarily. Otherwise, feasible folded shapes are made stiff by fixing the wedge angle, say, by gluing the open edges to the surface. Extra deformation in this case is promoted by flexibility of the facets alone. A wedge of material is inserted into the gap of Fig. 1(d) and, by adjusting the fold line rotations, the folded shape can be seated naturally upright; see Fig. 2(a). If the insertion is repeated for the continuum case and no other forces are applied, the disk will bend everywhere, including the wedge itself and attempt to form a \u201cfree-standing\u201d e-cone [3], which displaces upwards and sits above the original surface. Bending of the wedge in this case suggests folding the wedge in the discrete case as an extra mountain fold so that all facets move upwards, Fig. 2(b). As for the simple e-cone of Fig. 1(d), the valley folds cannot be separated by 180\u25e6 exactly, otherwise folding is impossible: in other words, our free-standing e-cone must be asymmetrically shaped. Other e-cones made with more fold lines can, of course, exhibit more symmetrical shapes but our simplified rendition is closest in behavior to the simplest continuum case of Fig 1(a). It is possible to recover the first shape of Fig. 2(a) by simply flattening the wedge fold of Fig. 2(b). Furthermore, its folding direction can be reversed to form another valley fold, requiring the disk to be simply supported on a rim as shown in Fig. 2(c). This shape can also be free-standing because the fold lines can be worked into a set of compatible rotations but we show it being held in equilibrium by a central point force\u2014as for the continuum case, which must be forcibly displaced from its free-standing position", + " Hence, x = \u03c6 2 tan \u03b2 2 , y = \u03c6 tan \u03b2 \u2212 x; \u21d2 d2 d1 = 2 [ tan \u03b2 tan \u03b2/2 \u2212 1 ] . (2) For \u03b2 = 60\u25e6, d2/d1 = 4, which tallies well with Fig. 3(b). The expression is independent of fold rotations provided they remain small, and it shows that the displacement ratio can increase asymptotically as \u03b2 approaches 90\u25e6, which we observe in other paper card models. Conversely, as \u03b2 tends to zero\u2014in the limit of the fold lines not overlapping in practice, the displacement ratio approaches two. III. DISCRETE E-CONES We begin with the simplest discrete e-cone of Fig. 1(d) equivalent to the \u201cplanar confined\u201d continuum case, as it is called, in Ref. [5]. Recall that the disk opens up along a radial wedge with adjacent facets, A and B, remaining horizontal. We define the angular excess to be the angle, \u03b1, subtended by the open wedge in Fig. 5(a). Its magnitude has been exaggerated for clarity, for otherwise, the original plan view of fold lines cannot alter during deformation. There are three folds, and the facets, C and D, move upwards. The Gauss mapping is also drawn in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000208_s12541-017-0045-0-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000208_s12541-017-0045-0-Figure1-1.png", + "caption": "Fig. 1 Hysteresis curve of a precision reducer3", + "texts": [ + "3,4 The performance of reducers can be estimated with a hysteresis curve, rotational transmission error and efficiency.5 Hysteresis curve is an integrated characteristic of two important performances: torsional stiffness and backlash. A procedure of measuring a hysteresis curve is as follows: when torque is applied to the output shaft while the input shaft is fixed, the output shaft has a small rotational motion, according to the torque applied. However, the torque-output rotation characteristics are not the same for loading and unloading, which is called the hysteresis curve in Fig. 1. Many important characteristics such as torsional rigidity, lost motion and backlash are defined from the hysteresis curve.3 There was previous research on the hysteresis characteristics of high precision reducers. In a harmonic drive, torsional rigidity was first studied by mathematical methods,6-8 FEM9 and experiments.10 In addition, the torsional rigidity of a two-stage cycloid drive without tolerance was analyzed by an iteration FE analysis and Hertz theory.11 Moreover, the lost motion of a cycloid reducer was analyzed with the iterative procedure of kinematic and FE analyses" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000894_tro.2015.2489499-Figure11-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000894_tro.2015.2489499-Figure11-1.png", + "caption": "Fig. 11. 3\u2013P RP robot and its hidden robot model (the gray joints denote the actuated joints). (a) 3\u2013P RP robot. (b) Its hidden robot model: a 3\u2013P RP robot known to be uncontrollable.", + "texts": [ + " This robot is well known to be architecturally singular as there is no way to control the translation along the axis of the parallel P joints. This result can be easily extended to the cases of the hidden robots of the UraneSX and the I4L (see Fig. 10). 3) robots with legs whose directions vary with the robot configurations but for which all hidden robot legs contain active R joints but only passive P joints: the most known robot of this category will be the planar 3\u2013PRP robot for which the hidden robot model is a 3\u2013PRP, which is known to be uncontrollable [25], [37] (see Fig. 11). The hidden robot model can be used to analyze and understand the singularities of the mapping and to study if a global diffeomorphism exists between the space of the observed element and the Cartesian space. However, not finding a global diffeomorphism does not necessarily mean that the robot is not controllable. This only means that the robot will not be able to access certain zones of its workspace (the zones corresponding to the assembly modes of the hidden robot model that are not contained in the same aspect as the one of the robot initial configuration)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001478_j.acme.2019.06.005-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001478_j.acme.2019.06.005-Figure1-1.png", + "caption": "Fig. 1 \u2013 Structure of", + "texts": [], + "surrounding_texts": [ + "Linear guides are the structural equivalents of linear kinematic pairs whose task is to maintain the relative displacement of individual machine body elements in a specific direction. These carriage systems are able to absorb the loads caused by cutting forces and gravity. The stiffness of the linear guide is a key factor for the stiffness of the machine tool, which is why it is necessary to define this in the design of the load-bearing system. Using the finite element method (MSC Nastran/Patran) [23], a complex model of stiffness was built for a MannesmannRexroth 1651-25 linear roller guide, implicitly taking into account phenomena such as the change in contact angle between the balls and raceways, changes in stiffness depending on the direction of the applied load as well as preload. The guide consisted of a carriage, a rail and four rows of balls with a contact angle of 458. In each row there were 13 balls in the active contact area. The construction of the guide and its stiffness are presented in Figs. 1 and 2, respectively. The first stage of building the linear guide stiffness model was construction of the geometric model. Due to the fact that the manufacturer did not provide full technical documentation, it was necessary to carry out additional geometric measurements, especially within the contact surfaces (ball radius, raceway radius). Then, the geometric model underwent simplifications, consisting of omitting elements that were not significant from the point of view of the analyzed phenomenon, such as small roundings, phases, mounting holes and inclinations. In the next stage, the model was discretized with the use of solid elements (CHEXA type, six-sided isoperimetric elements with 8 nodes), material data was defined for individual elements (Young's modulus, Poisson's number) and the surface-to-surface contact between the balls and raceways was modeled corresponding to Hertzian contact model. The model building process is presented in Fig. 3. An extremely important issue affecting the accuracy of computations in a contact model is selection of the appropriate mesh density [24,25]. Bearing in mind both the aforementioned accuracy of computations and the time needed for their implementation, it was decided to perform local mesh refinement in the vicinity of the contact areas. Preliminary tests for various finite element mesh densities were performed, and a density was chosen that made it possible to obtain best agreement of the actual stress distribution with Hertzian theory (the maximum stress in the contact area in the constructed model was 3620 MPa, compared to the theoretical value of 3660 MPa). The detailed procedure including mesh dependency was presented in the paper [12]. The next step was to determine the stiffness characteristics of the guide, achieved by means of a two-stage non-linear analysis. In the first stage, the preload was modeled (0.08 of dynamic load-bearing capacity, C = 22,800 N) [22] by applying a 77.8 8C heat load to the balls, resulting in an increase in size and thus the creation of stresses with the assumed value. The results of the calculations from this stage served as boundary conditions imposed in the second step, in which the deformation of the carriage was determined for an incrementally rising load applied to the upper surface of the carriage in ten steps from zero to the dynamic load-bearing capacity of 22,800 N, while lower surface of the rail was fixed. As a result of this analysis (MSC Nastran Nonlinear Static \u2013 SOL 106), the stiffness characteristics of the carriage were the linear guide. obtained and compared to the data provided by the manufacturer, determining that the model was a good representation of its real counterpart. A comparison of the characteristics is shown in Fig. 4. 3. Building a simplified model Despite the good compatibility of the model presented in Section 2, the long computation time (about 8 h) and necessarily large processing power was less than desirable, so simplifications which would balance the accuracy of the model and shorten the computation time, and tools that would enable an effective use of the model as an element of a complex structure. Therefore, it seems justified to use substructuring and static reduction, which has a certain limitation in that all elements of the model within the reduced component must be characterized by a linear model. Bearing in mind the limitations of the method and the purpose of the model (determination of stiffness in technological machines), the new model of linear guide stiffness had the following assumptions: the linear guide is preloaded, displacements in the system are small, loads acting on the system do not cause plastic deformation, there is no kinematic input (connected with inaccurate mounting of the guide), and there are no errors in the system related to the geometry of the balls (the balls are perfectly spherical and have the same diameter) [26]. This idealization was based on the range of loads during machine tool operation. It can be seen that as the load is increased uniformly, the increasing rate of deformation declines. Therefore, above the inflection point of the loaddeformation curve, the stiffness characteristic is almost linear. It should also be noted that an increase in preload is accompanied by a decrease in the argument of the stiffness function (load value) for which the inflection point occurs [22,27], thereby increasing the range in which stiffness characteristic can be described by a linear model. In technological machines, linear guides are usually selected in such a way that they work in the range of working loads above the inflection point of the stiffness characteristic. The non-linearity of the stiffness characteristic of the model is caused by the model of contact between the raceways and the ball. This model can be considered as the contact between a deformable ball and cylinder, which is schematically shown in Fig. 5. This issue is described by Hertzian equations, allowing determination of the state of deformation and the state of stress in the areas of contact between two bodies. The contact surface area changes depending on the value of the pressing forces, which results in a change in the values and distributions of stresses as well as deformations and displacements in the contact area. After the deformation of the bodies caused by their mutual pressure (Fig. 5), in a general case the contact area formed is an ellipse with semi-axes a and b (a > b), which can be determined from the equations: a \u00bc a3 ffiffiffiffiffiffiffiffi P m n r ; b \u00bc b3 ffiffiffiffiffiffiffiffi P m n r (1) m \u00bc 4 \u00f01=r1\u00de \u00fe \u00f01=r01\u00de \u00fe \u00f01=r2\u00de \u00fe \u00f01=r02\u00de ; n \u00bc 8 3 E1E2 E2\u00f01 v1\u00de \u00fe E1\u00f01 v2\u00de (2) where a and b \u2013 coefficients depending on the value of the ratio B/A where: A \u00bc 2 m ; B \u00bc 1 2 ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 1 r1 1 r01 2 \u00fe 1 r2 1 r02 2 \u00fe 2 1 r1 1 r01 1 r2 1 r02 cos2\u2019 s (3) Based on the above formulas, assuming material data analogous to the model developed in Section 2 (E1 = E2 = 210,000 N/mm2, v1 \u00bc v2 \u00bc 0:3, r1 \u00bc r01 \u00bc 2:375 mm, r2 \u00bc 2:5 mm; r02 \u00bc 1) and the load P = 10 kN, the dimensions of the contact ellipse were: a = 0.524 mm, and b = 0.183 mm. This theoretical contact ellipse was used to eliminate the non-linear surface-to-surface contact model. A substitute linear model was developed where the ball was replaced with four rod elements of equivalent stiffness, stretched between the opposite ends of the face of cuboidal finite elements (the dimensions of the face correspond to the rectangle described in the contact ellipse). A schematic representation of the replacement of the contact model is shown in Fig. 6. For such a developed model, the stiffness characteristic was determined in a manner analogous to that shown in Section 3. The results are shown in Fig. 7. In addition, Fig. 8 presents a comparison of displacements of the carriage with respect to the rail under load, and for comparison, results for the methods constructed in accordance with [12] and [9], taking into account only vertical stiffness for a similar number of degrees of freedom. Table 1 shows the computation times. It can be concluded that the proposed model shows good compatibility with its real counterpart. Replacing the ball with rod elements and the surface-to-surface contact model with an equivalent contact model, resulted in a significant reduction in computation time, with only a slight decrease in accuracy. In addition, it should be noted that the developed model does not include elements with non-linear stiffness characteristic, and unlike other models it does not include multipoint constraints. Therefore, it is possible to reduce the degrees of freedom using a Guyan reduction [28] and implementation in the substructuring method. 4. Simplified model application in substructuring method This section presents an example of the developed model in determining the stiffness of a machine table load-bearing system. As mentioned before, machine tools are complex mechanical systems consisting of many elements interconnected with a variety of connections forming kinematic constraints. Due to the complexity of the model, determining the stiffness of the machine using the classic finite element method can be very time consuming and hence inefficient. The computation time can be significantly shortened by substructuring. Substructuring is based on the idea of dividing a complex model built in the convention of the finite element method (FEM) into smaller component-based models known as substructures. This approach, apart from significantly shortening the computation time, has a number of other advantages, enabling analysis of individual elements to identify local parameters describing their properties, changes to the model of a single component without changing the global model, creating a library of the most frequently used components, and thus using the same model of an element occurring several times in the structure (e.g. linear guides). In the case of structures whose elements are made by various manufacturers, it is possible to implement models of individual components provided by them, without the need to build a separate model, which allows a significant reduction of preparation time of the entire structure. The models presented in the article were established using the finite element method (MSC Nastran/Patran) [23]. The essence of this method is to replace the continuous model of the mechanical system described by means of partial differential equations, by a discrete model comprising a system of algebraic equations:" + ] + }, + { + "image_filename": "designv10_14_0002175_tie.2021.3084172-Figure8-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002175_tie.2021.3084172-Figure8-1.png", + "caption": "Fig. 8. Flux density harmonic distributions. (a) CWC. (b) CWoC.", + "texts": [ + " See http://www.ieee.org/publications_standards/publications/rights/index.html for more information. The amplitudes of post-fault back EMF voltages in the two cases are analysed by FEM, as shown in Table II. It can be seen that the back EMF voltage amplitudes in CWoC are lower than those in CWC, which indicates that the reconstructed armature MMF contains demagnetized MMF for the proposed strategy. Flux density distributions of the investigated machine in the two cases are analysed by FEM, as shown in Fig. 8. It can be seen that the stator core flux density in CWoC is obviously lower than that in CWC, which is beneficial for improving post-fault performances. Losses and efficiencies in the two cases are analyzed by FEM, as shown in Table III. Copper losses in the two cases are the same. In CWoC, the iron loss is reduced obviously for lower flux density in stator iron core, and lower PM eddy loss is obtained. The slight difference of PM eddy current loss is associated with the characteristics of armature MMF, including orders, amplitudes, and rotating directions of the harmonics, etc" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000215_s11740-017-0734-7-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000215_s11740-017-0734-7-Figure1-1.png", + "caption": "Fig. 1 Geometric CWR-tool parameters of a three-zone single wedge tool", + "texts": [ + " CWR is a preforming process for reshaping circular cylindrical billets into rotationally symmetrical workpieces with variable diameter in axial direction using two oppositely moving wedge-shaped tools. In CWR the wedges of the tools distribute the material which results in an unequal mass distribution along the main axis [7]. CWR is a favourable preforming operation especially due to the high material utilization [8]. The main parameter describing a cross wedge rolled part is the cross section area reduction \u0394A (Fig.\u00a01). The main tool parameters in CWR processes are the forming- and wedge-angle (see Fig.\u00a01). Process parameters for typical CWR processes with a monolithic material are the billet and tool temperature and the forming velocity. Although CWR offers a lot of advantages, it has not been widely accepted throughout the forging community for a long time. One of the main reasons is the complexity of CWR tools and process design. However, latest researches in the field of CWR enable more simplified developments of CWR processes as suitable simulation parameters were identified [9] and a part geometry can be designed automatically [10, 11]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001447_j.mechmachtheory.2019.03.002-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001447_j.mechmachtheory.2019.03.002-Figure3-1.png", + "caption": "Fig. 3. Kinematic diagram of PGT-II.", + "texts": [ + " The kinetic relations will be established after the causality definition. As shown in Fig. 2 There are three ports which connect to the external system, namely, sun gear, planet carrier and ring gear. Planetary gear is a transmission link in the system, which has no direct contact with the external system. The kinetic equations of the model can be obtained after causality definition which will be discussed later. PGT-II is a planetary gear trains driven by two planetary gears between the ring gear and the sun gear. Its diagram is shown in Fig. 3 . The planetary gear contacting with the sun wheel is set to P1, and the outer planetary gear contacting with the gear ring is set to P2. Other symbols refer to the same thing as PGT-I Type. The Kinematic Equations is shown as follows: \u03c9 r1 = \u03c9 r \u2212 \u03c9 c (5) \u03c9 p2 r p2 = \u03c9 r1 r r (6) \u03c9 s 1 = \u03c9 s \u2212 \u03c9 c (7) \u03c9 p1 r p1 = \u03c9 s 1 r s (8) \u03c9 p1 r p1 = \u03c9 p2 r p2 (9) Similar to PGT-I, the system also has three ports connected to the external system, namely sun, carrier and ring. Bond graph theory has strict rules for causality definition" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002061_tec.2020.3048442-Figure35-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002061_tec.2020.3048442-Figure35-1.png", + "caption": "Fig. 35. Summary of overall characteristics.", + "texts": [ + " The efficiency maps along the torque-speed curves of the three machines are further compared in Fig. 34. Since the field winding produces additional copper loss in the HE machine, the two DSHE machines exhibit lower efficiency compared with the DSPM machine. As stated above, the conventional DSHE machine employs overlapped field winding, which tends to increase the machine volume and copper losses. As the result, the conventional DSHE machine exhibits lowest efficiency among the comparison. Based on the above calculation, the characteristics of three machines are summarized in Fig. 35. The DSPM machine Authorized licensed use limited to: Western Sydney University. Downloaded on June 14,2021 at 20:12:10 UTC from IEEE Xplore. Restrictions apply. 0885-8969 (c) 2020 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. > REPLACE THIS LINE WITH YOUR PAPER IDENTIFICATION NUMBER (DOUBLE-CLICK HERE TO EDIT) < 13 exhibits compact machine structure with non-overlapped winding, relatively high torque density and efficiency, whereas the fault tolerance and flux weakening capability is not comparable with the HE machines" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002049_1464419320972870-Figure5-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002049_1464419320972870-Figure5-1.png", + "caption": "Figure 5. Schematic of load zone of ball bearing.", + "texts": [ + " When the shaft rotates with xs, the angular position of the ith defect on inner raceway changes as equation (13), hi di \u00bc xst\u00fe hi0 \u00fe i 1\u00f0 \u00dehIAD i \u00bc 1 : NDI (13) where NDI is the number of defects on inner raceway. When the outer ring is fixed on bearing pedestal, the angular position of the i-th defect on outer raceway remains constant, which is shown as equation (14), hi do \u00bc ho0 \u00fe i 1\u00f0 \u00dehOAD i \u00bc 1 : NDO (14) where NDO is the number of defects on outer raceway. When radial load Fr is applied on Y direction, there would be a load zone in ball bearing, which is shown in Figure 5. Radial displacement and contact force will change when the balls roll over multiple defects on outer raceway. Every localized defect on inner and outer raceway would cause an additional displacement did-i and did-o, and they are calculated as equations (15) and (16), respectively, did i \u00bc db 2 1 cos hdi 2 0 < mod\u00f0hi hi di; 2p\u00de < hid (15) did o \u00bc db 2 1 cos hdo 2 0 < mod hi; 2p\u00f0 \u00de hi do < hod (16) where db is the diameter of the ball, hdi and hdo are the angle of defect on inner and outer raceway" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002436_tmag.2021.3087460-Figure4-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002436_tmag.2021.3087460-Figure4-1.png", + "caption": "Fig. 4 Experimental set up", + "texts": [ + " Some techniques are then applied to reduce the cogging component of the AFPM motor to low levels. Since measuring the cogging torque require decoupling the effect of drive and the device under test (DUT). At the same time the dynamometer is be chosen with minimal impact to the cogging generated by the DUT. In addition, the sensors and measurements systems should be capable of measuring the changes in the cogging torque accurately. Considering these requirements, the test set up is configured, that is given in Fig. 4. The cogging torque of each AFPM motor before and after the proposed method are presented to illustrate the capability and the sensitivity of the proposed method. The cogging torque variations with respect to position and their harmonics contends for each motor before and after the proposed compensation method are given in Fig. 5, Fig. 6, Fig. 7 and Fig. 8. To be able to demonstrate the impact of the proposed method for different motors, both motors were operated at 10\ud835\udc5f\ud835\udc5d\ud835\udc5a with 5.8\ud835\udc34\ud835\udc5f\ud835\udc5a\ud835\udc60 current. Since each motor has different torque constants, the resulting torque for each is different" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000769_978-3-662-43893-0-Figure4.32-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000769_978-3-662-43893-0-Figure4.32-1.png", + "caption": "Fig. 4.32 Load and stress distribution on the bevel gear", + "texts": [ + "4 Stress Analysis 169 The load distribution determined along the contact lines forms the basis for subsequent calculation of local stresses. Numerical, analytical and partly analytical methods may be used for the calculations. Loads act on a structural assembly or a component thereof in the form of torques, bending moments, external forces, centrifugal and frictional forces, speeds, velocities and temperatures. As a result of these loads, strains occur in the component, manifesting them- selves as deflections, stresses or heating (Fig. 4.32). According to the hypothesis of cumulative damage, each stress leads to a progressive damage state (see Sect. 4.2.7). If a certain limit is exceeded, this causes damage such as wear, pitting, scuffing etc. (see Sect. 4.1). The limit is usually a strength value which, when related to the stress, will produce a safety factor. The main external loads acting on a gear system are torques and speeds of rotation. Over a particular period of time, these may be regarded either as a constant nominal load or as a collection of discrete loads" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000894_tro.2015.2489499-Figure12-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000894_tro.2015.2489499-Figure12-1.png", + "caption": "Fig. 12. G\u2013S platform and its hidden robot model. (a) G\u2013S platform from DeltaLab: a 6\u2013UP S robot. (b) Its hidden robot model: a 3\u2013UP S robot (when three legs are observed).", + "texts": [ + " This is, of course, a problem if the operational workspace of the real robot is fully or partially included in these zones. Robots belonging to this category are probably the most numerous. They are those for which the hidden robot models have several possible assembly modes, whatever is the number of observed leg directions. Presenting an exhaustive list of robots of this category is totally impossible because it requires the analysis of the assembly modes of all hidden robot models for each robot architecture. However, some examples can be provided. Examples of such types of robots [the G\u2013S platform (see Fig. 12) and the Adept Quattro (see Fig. 13)] have been presented in [19], [21], and [22]. More specifically, in [21] and [22], it was shown (numerically but also experimentally) that the Adept Quattro [41] controlled through leg direction observation has always at least two assembly modes of the hidden robot model, whatever the number of observed legs. As a result, some areas of the robot workspace were never reachable from the initial configuration. Fig. 14 shows a desired robot configuration that was impossible to reach even if all robot legs were observed" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000130_j.mechmachtheory.2013.07.010-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000130_j.mechmachtheory.2013.07.010-Figure3-1.png", + "caption": "Fig. 3. Coordinate systems and common moving frames with errors.", + "texts": [ + " From a technical point of view, the worm surface is modified, namely that the radius of the tooth profile of the worm is designed to be a little bigger than that of the steel ball. As a result, the line contact of conjugate teeth is altered into the point contact. Nonetheless, the contact properties of teeth can still be analyzed in the normal plane to the directrix of the worm surface, and it is advisable to apply the moving frame of the worm surface as a commonmoving frame of conjugate tooth surfaces at the contact position. As shown in Fig. 3, at the predesigned contact point PR1, the commonmoving frame {PR1;\u03b1s1\u03b1s2\u03b1s3} for conjugate tooth surfaces is represented by: \u03b1s1 \u00bc \u03b11 \u03b1s2 \u00bc \u2212\u03b12 sin\u03b1n1\u2212\u03b13 cos \u03b1n1 \u03b1s3 \u00bc \u03b12 cos\u03b1n1\u2212\u03b13 sin\u03b1n1 8< : \u00f016\u00de \u03b1s1 is the unit tangent vector of conjugate tooth surfaces along the tooth length direction; \u03b1s2 is another unit tangent where vector along the tooth depth direction; \u03b1s3 is the unit normal vector pointing to the center of the tooth profile; \u03b1n1 is the normal pressure angle of tooth. 3.2. Error and variation 3" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002479_j.matdes.2021.110025-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002479_j.matdes.2021.110025-Figure2-1.png", + "caption": "Fig. 2 Schematic diagram of sampling. (a) cross-section specimens, (b) tensile specimens", + "texts": [ + " Then two pieces of substrates were cleaned to serve as the experimental group while the other two do not do cleaning treatment to serve as the control group in the additive forging experiment. Ultrasonic cleaning machine (JP-060, China) was employed in cleaning experiment, and cleaning steps are as follows: (1) Gently wipe the substrate surface with a dust-free cloth stained with acetone solution. (2) Put the substrate into the ultrasonic cleaning machine and cleaned the substrate for 15 minutes within To study the bonding effect of the interface, twenty cross-section specimens passing through the interface were cut out from T1 and T2 respectively, as shown in Fig. 2(a). The sampling region should stay as close to the bonded samples center as possible to eliminate the effects of uneven strain due to friction at the interfaces. The sampling rule of the tensile specimens is similar to the sampling rules of the cross-section specimens, and the dimension of the tensile specimens is shown in Fig. 3. Fig. 3 Dimension of the tensile specimen (Unit: mm) Cross-section specimens were mechanically ground and polished by diamond abrasive with the particle size of 2.5 \u03bcm and then were processed by electro etching with the parameters of 12V and 2A for 120 seconds in a 10 % H2C2O4 (vol%) solution", + " By contrast, the heated substrates were forged at 1200\u2103 with deformation strains of 40% and were maintained for 24 hours at 1200\u2103 in this paper, which is more conducive to the decomposition and diffusion of oxide layers. Besides, the exclusive variable designed in the experiment was cleaned substrate or uncleaned, so the variation in the bonding performance can be attributed to substrate cleaned or uncleaned. Hence, the influence of surface oxide layers on the interface bonding can exclude. 3.2.1 Cross-section morphology Twenty cross-section specimens were cut out from T1 and T2 respectively as shown in Fig. 2(a). Olympus microscope (MX40, Japan) was used to observe the cross-section morphology. Fig. 10 shows the image of cross-section specimens got from T1. Fig. 10 Images of cross-section specimens got from T1, (a) specimen 1(5\u00d7), (b) specimen 2(5\u00d7), (c) specimen 1(50\u00d7), (d) specimen 2(50\u00d7) In Fig. 10 (a) and (b), the interface appears, but only part of the interface appears rather than the whole interface run through the specimen. This means that the substrates were not bonded together perfectly. Fig", + " Surface cleaning has a great influence on the bonding performance by comparing the cross-section morphology of bonded samples produced by the cleaned and uncleaned substrates respectively in additive forging. The cleaned substrate is more conducive to the bonding of the interface, and to obtain a nearly perfect bonding state. On the contrary, the uncleaned substrate is suffering the obstruction effect brought by contaminants in the interface bonding process, resulting in poor interface bonding performance. 3.2.2 Tensile property Three tensile specimens were cut out from T1 and T2 respectively as shown in Fig. 2(b). Fig. 13 exhibitions the stress-strain curves of tensile specimens at room temperature, it can be seen from the figure that no macroscopically brittle fracture occurs in all the tensile specimens, but the maximum stress and strain of the tensile specimens (T1-1, T1-2, T1-3) derived from T1 were slightly better than that of the tensile specimens (T2-1, T2-2, T2-3) derived from T2. Table 3 shows the stress and strain values of tensile specimens, the stress of T1-1, T1-2, and T1-3 stabilize at about 435 MPa, while 420 MPa were got from the tensile specimens derived from T2" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000419_1.3656641-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000419_1.3656641-Figure2-1.png", + "caption": "Fig. 2 Precision rolling-disk machine and X-ray system", + "texts": [ + " The five-ball fatigue tester was modified to measure the temperature near the contact area of a modified test specimen during operation. Fig. 1(c) shows the tost specimen and mounting assembly, which is inserted into the drive spindle of the five-ball rig [see Fig. 1(a)). Each test specimen had a thermocouple attached with the tip at one edge of the running track. An axial hole was drilled through the drive spindle to accept the thermocouple wire. The thermocouple emf was taken out through a slipring brush assembly mounted at the top of the drive spindle. The rolling-contact disk machine is shown in Fig. 2. This machine, described in detail in references [3 and 4], was used to measure lubricant film thickness and contact deformation under dynamic conditions. Essentially, this method of measuring film thickness consists of directing a monochromatic , collimated, square beam of high-energy X rays between two rolling-disk surfaces. The amount of radiation passing between the disk parallel to the flat contact regions is related to the thickness of the lubri- cant film separating the surfaces. A particular wavelength X - r a y was selected that penetrated lubricants quite readily but did not penetrate steel significantly" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000885_ecc.2015.7331093-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000885_ecc.2015.7331093-Figure1-1.png", + "caption": "Fig. 1: Forces and moments on the UAV.", + "texts": [ + " Using rigid body assumption, a generic unmanned air vehicle system dynamics is given as below:[ mI3x3 03x3 03x3 Ib ] [ V\u0307w \u2126\u0307b ] + [ 0 \u2126b \u00d7 (Ib\u2126b) ] = [ Ft Mt ] (1) where m and Ib represent the mass and the inertia matrix in the body frame and Vw and \u2126b represent the linear velocity 978-3-9524269-3-7 \u00a92015 EUCA 3623 with respect to world frame and the angular velocity with respect to body frame of the vehicle, respectively. The net force and the moment applied on the vehicle are represented by Ft and Mt, respectively (see Fig. 1). It is noted that for tilt-wing quadrotors, these forces and moments are functions of the rotor trusts and wing angles. Using vector-matrix notation, (1) can be rewritten as follows: M\u03b6\u0307 + C(\u03b6)\u03b6 = G+O(\u03b6)\u03c9 + E(\u03be)\u03c92 +W (\u03be) (2) where, \u03b6 = [X\u0307, Y\u0307 , Z\u0307, p, q, r]T , \u03be = [X,Y, Z,\u03a6,\u0398,\u03a8]T (3) and where X,Y and Z are the coordinates of the center of mass with respect to the world frame, p, q and r are the angular velocities in the body frame and \u03a6,\u0398 and \u03a8 are the roll, pitch and yaw angles of the vehicle expressed in the world frame" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000769_978-3-662-43893-0-Figure6.18-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000769_978-3-662-43893-0-Figure6.18-1.png", + "caption": "Fig. 6.18 Elements of a generating process without plunge cutting", + "texts": [ + " In any generating process, very different relative positions occur between the work piece and the tool along the path of generation, Therefore, quite different chips are produced as all points of the cutting edge successively come into contact with the work piece during the generating process, the top of the blades being the most solicited. In the single indexing method, the tool rapidly cuts the blank to full tooth depth, at the start-of-roll position, and then performs the generating process until the endof-roll position is reached. The tool is then withdrawn from the finished tooth slot and is traversed back to the starting position. After the indexing operation the next tooth slot is generated (Fig. 6.18). This sequence can be shortened on modern machines by simultaneously moving a number of axes. With the continuous indexing method, the start-of-roll position is the same as before, but once the generating process is completed, not only one but all the tooth slots are finished and the machine can return to its start-of-roll position. Depending on the tool and the required tooth flank surface quality, a single generating cycle may not be sufficient. For example, coarse chips may scratch the tooth flanks and another generating cycle with a small finishing tool feed may be required" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002220_s00521-021-06314-x-Figure5-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002220_s00521-021-06314-x-Figure5-1.png", + "caption": "Fig. 5 Isometric and front view of joint definition", + "texts": [ + " Since the goal is to determine how easy it is to identify cracks in the sun gear due to the rotation of the carrier, four different conditions were considered for the PGT, which are: 1. Healthy condition with no cracks, 2. Sun gear with 0.02 mm elliptical crack, 3. Planet pinion with 0.02 mm elliptical crack, 4. Both sun and planet with 0.02 mm elliptical cracks. For the single-stage planetary gear system modeled in MSC ADAMS, it is critical to define joints and constraints. Table 1 describes the joints created in MSC ADAMS for each of the components. Figure 5 illustrates the joint definition of the PGT model. In this arrangement, the input motion is applied to the sun shaft and the carrier shaft shall be the output. The ring gear is fixed and does not rotate. To determine the dynamic behavior of the PGT, the fixed joint between the ring and the ground shall be used for our analysis, and to simulate a real-world scenario, the magnitude of the force at this joint shall be measured. For each of the PGT healthy conditions, certain input conditions on MSC Adams were applied as shown in Table 2" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000130_j.mechmachtheory.2013.07.010-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000130_j.mechmachtheory.2013.07.010-Figure2-1.png", + "caption": "Fig. 2. Theoretical contact area on the worm-gear tooth.", + "texts": [ + " The worm surface is generated by sweeping the circular-arc tooth profile along the directrix, so the distribution of contact lines on the worm surface is simple, namely that at every instant the contact line coincides with the normal tooth profile of the worm, and the total contact area covers on the whole worm surface. Toward the worm-gear tooth, its instantaneous contact line must be the line of section of the steel ball surface intersected by the normal plane to the directrix of the worm surface. Because the tangential direction of the directrix varies with the lead angle, it is sure that the contact line on the worm-gear tooth swings about the e3 (d)-axis, and a theoretical contact area is generated on the worm-gear tooth, as shown in Fig. 2. Obviously, the lead angle of the directrix of the worm surface is an important factor to determine the distribution of contact lines on the worm-gear tooth. This result is obtained with the assumption that the steel ball is rigidly fixed on the worm gear, but because of the friction action the steel ball could perform a redundant rotation about the e3 (d)-axis in the finger-like sustaining socket of the worm gear. Then, an actual contact area on the worm-gear tooth should cover the half of the steel ball surface" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001901_j.mechmachtheory.2020.103823-Figure4-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001901_j.mechmachtheory.2020.103823-Figure4-1.png", + "caption": "Fig. 4. Coordinate transformation of gear pair.", + "texts": [ + " E l,m i = \u23a1 \u23a2 \u23a2 \u23a2 \u23a2 \u23a3 x td,m i ( \u03d5 td,m i ) x id,m i ( \u03d5 id,m i ) x a,m i ( \u03d5 a,m i ) x iu,m i ( \u03d5 iu,m i ) x tu,m i ( \u03d5 tu,m i ) y td,m i ( \u03d5 td,m i ) y id,m i ( \u03d5 id,m i ) y a,m i ( \u03d5 a,m i ) y iu,m i ( \u03d5 iu,m i ) y tu,m i ( \u03d5 tu,m i ) z td,m i ( i ) z id,m i ( i ) z a,m i ( i ) z iu,m i ( i ) z tu,m i ( i ) 1 1 1 1 1 \u23a4 \u23a5 \u23a5 \u23a5 \u23a5 \u23a6 (8) 2.4. The surface profile equation of the gear pair in the meshing CS The model for calculating TSCS is defined at the meshing CS O e X e Y e Z e , in which the surface profile equation of the gear and pinion can be obtained by coordinate transformation on the basis of E l,m i , as shown in Fig. 4 . First, the gear and pinion are converted to their actual working CSs O wg X wg Y wg Z wg and O awp X awp Y awp Z awp , with the transformation matrices M lg2wg and M alp2awp from their actual local CSs, respectively. Next, pinion transforms from the actual CS O awp X awp Y awp Z awp to the ideal CS O wp X wp Y wp Z wp are obtained with the transformation matrix M awp2wp , and the ideal working CS of the gear coincides with the actual CS. Moreover, the gear and pinion transforms from the ideal working CSs O wg X wg Y wg Z wg and O wp X wp Y wp Z wp to the meshing CS O e X e Y e Z e are obtained with M wg2e and M wp2e , respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001990_s40516-020-00128-w-Figure5-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001990_s40516-020-00128-w-Figure5-1.png", + "caption": "Fig. 5 Representation of tensile test sample", + "texts": [ + " After milling to a depth of 2mm, HRA hardness tests (EN-ISO 6508) were performed on samples from the second set of tests along the center of the largest lateral face at regular intervals in the vertical direction. By determining the hardness along the entire height of each sample, it was possible to determine the average hardness and uniformity of the resulting hardness profile. The locations of hardness tests are shown schematically in Fig. 4. Tensile tests were performed on samples that had been machined to the standard geometry defined in ISO 6892- 1:2016, shown in Fig. 5. Tests were performed with an Instron 8032 machine with a deformation rate of 0.00025 s\u22121 as per the aforementioned standard. The yield strength was calculated based on the point at which each sample reached 0.2% permanent deformation, while the ultimate tensile strength was based on the maximum load. A number of process variables were monitored during production of the second set of samples, including the substrate temperature, deposition efficiency, production time and specific energy input. The substrate temperature was monitored with 1mm diameter K-type thermocouples placed in two 1" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000654_j.isatra.2017.09.012-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000654_j.isatra.2017.09.012-Figure1-1.png", + "caption": "Fig. 1. The schematic representation of quadrotor.", + "texts": [ + " UAV modeling The UAV model considered in this article takes account of drag forces acting on the rigid body, gravity and the thrust generated by DC motors [37]. Consider an Earth-fixed inertial reference frame Please cite this article as: Chang J, et al. Analysis and design of sec estimation. ISA Transactions (2017), http://dx.doi.org/10.1016/j.isatra. and a body fixed frame attached to the quadrotor. Let {\u2192 \u2192 \u2192}e e e, ,x y z be the unit basis vectors of earth frame and {\u2192 \u2192 \u2192 }e e e, ,bx by bz be the unit basis vectors of the body frame. The body frame is aligned along the quadrotor's principal axis of symmetry as seen in Fig. 1. The change of coordinates from the body frame to inertial frame is governed by three Euler angles \u03d5 \u03b8 \u03c8[ ], , , for roll, pitch and yaw respectively while \u03c9 = [ ]p q r, , T combines the angular velocity vector for the same sequence. Thus, the rotational kinematics relating these angular velocities to the Euler angles is expressed as \u23a1 \u23a3 \u23a2 \u23a2 \u23a2 \u23a4 \u23a6 \u23a5 \u23a5 \u23a5 \u23a1 \u23a3 \u23a2 \u23a2 \u23a2 \u23a4 \u23a6 \u23a5 \u23a5 \u23a5 \u23a1 \u23a3 \u23a2 \u23a2 \u23a4 \u23a6 \u23a5 \u23a5 \u03d5 \u03b8 \u03c8 \u03d5 \u03b8 \u03d5 \u03b8 \u03d5 \u03d5 \u03d5 \u03b8 \u03d5 \u03b8 \u0307 \u0307 \u0307 = \u2212 ( ) p q r 1 sin tan cos tan 0 cos sin 0 sin sec cos sec 1 The transformation from the body frame to earth frame is given by the rotation matrix b e and is defined based on a 3-2-1 rotation sequence as follows \u23a1 \u23a3 \u23a2 \u23a2\u23a2 \u23a4 \u23a6 \u23a5 \u23a5\u23a5 \u03b8 \u03c8 \u03d5 \u03b8 \u03c8 \u03d5 \u03c8 \u03d5 \u03b8 \u03c8 \u03d5 \u03c8 \u03b8 \u03c8 \u03d5 \u03b8 \u03c8 \u03d5 \u03c8 \u03d5 \u03b8 \u03c8 \u03d5 \u03c8 \u03b8 \u03d5 \u03b8 \u03d5 \u03b8 = \u2212 + + \u2212 \u2212 c c s s c c s c s c s s c s s s s c c c s s s c s s c c c b e where s" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000771_978-3-319-16823-4-Figure8.10-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000771_978-3-319-16823-4-Figure8.10-1.png", + "caption": "Fig. 8.10 Oscillations in mean degree during dynamic reconfiguration caused by changing node concentration. (a) Plot of mean degree with indicative network structures, (b) 3D plot of node weight concentration throughout the experiment.", + "texts": [ + "9 shows the evolution of one network which follows a similar evolution to those with regular node arrays: A \u2018backbone\u2019 of outer nodes is the first to stabilise and then the configuration of inner nodes is incorporated into the network structure. Unlike the regular arrays, however, the network configuration never fully stabilises. Certain configurations are generated (for example those shown on the bottom row of Fig. 8.9), only to be remodelled (because one or more nodes has mean degree of 3). This pattern of evolution was found to be very repetitive with the network oscillating between semi stable states. Fig. 8.10a illustrates the effect of the DR method on the mean degree of node connectivity throughout an experimental run. The mean degree fluctuates as the network configuration changes in response to the dynamical adjustment of node concentration. The images above the plot (dashed boxes) show configurations at the time of increasing mean degree (i.e. sub-optimal networks at the crests of the plot). These configurations are typically transient and occur during major shifts of network patterns. The lower images below the plot (solid boxes) coincide with local minima of mean degree. Note that although these local minima have mean degree close to the desired specification (2), the patterns are not stable (due to isolated nodes or occasional degree of 3) and the network soon transforms into another configuration. The evolution of the network only halts when stability of configuration has been achieved. Also shown in Fig. 8.10b is an illustration of the changing node concentration for each of the 11 nodes throughout the same experiment (time direction is arrowed). The highest point on each node \u2018lane\u2019 corresponds to maximal node concentration (10, no inhibition of node weight at all) and the lowest point corresponds to the minimum concentration (zero, maximum inhibition of node weight, i.e. no projection of attractants). The pattern of node concentration shows the independent nature of the node weight adjustment by the feedback mechanism, and it is only when the network stabilises into the final pattern that the node weights are synchronised" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000794_978-1-4939-2065-5_7-Figure7.2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000794_978-1-4939-2065-5_7-Figure7.2-1.png", + "caption": "Fig. 7.2 (a) Two-dimensional waving sheet in a viscous fluid illustrating the traveling wave of velocity c progressing in the x-direction and the forward swimming speed (U) in opposite direction. (b) Resistive force theory (RFT) diagram illustrating the normal and tangential components of the velocity U and force F , and the resulting net propulsive force", + "texts": [ + " These properties illustrate that swimming at low Reynolds numbers can seem at first as a highly confined phenomenon, yet microorganisms have found a variety of ways to overcome the constraints of the scallop theorem. In what follows, we briefly review some of the classical theories that have shed light on the hydrodynamic mechanisms leading to net propulsion at low Reynolds numbers. The discussion will be limited but the reader can find a more thorough review in [2, 5, 8, 30] as well as in Chaps. 1 and 8 of this book. Over half a century ago, Taylor [27, 28] beautifully demonstrated that an infinite waving sheet (see Fig. 7.2a) could swim in an incompressible, Newtonian fluid by generating traveling waves in the absence of inertia or vanishing Reynolds numbers. Note that the hydrodynamics of Taylor\u2019s waving sheet is governed by Eq. (7.1). In Taylor\u2019s work, the planar sheet oscillates in time in a prescribed form according to y(x, t) = asin(kx\u2212\u03c9t), where a is the traveling wave amplitude, \u03c9 is the frequency, \u03bb = 2\u03c0/k is the wavelength, c = \u03c9/k is the traveling wave speed, and k is the wave number. Taylor found that the sheet oscillations induce a forward velocity U = \u03c9a2k/2+O(ka)4 [27], where the sheet is propelled in the direction opposite to that of the propagating wave (Fig. 7.2a). Many important investigations followed Taylor\u2019s landmark contribution. Of particular relevance, we highlight the well-known resistive force theory (RFT) introduced by Gray and Hancock in analyzing the locomotion of sperm cells [3]. There, the authors assumed that the hydrodynamic forces experienced by the organism would be approximately proportional to the local body velocity such that the force exerted by a body or flagellar segment is given by F = CNUN +CT UT, where C corresponds to the local drag coefficient per unit length (dependent on geometry and fluid viscosity) and N and T are the normal and tangential components, respectively (see Fig. 7.2b). Hence, the total thrust can then be obtained by integrating the propulsive force over the entire body or flagellum length. It is namely the anisotropy between the normal and tangential drag coefficients, with CN >CT , that lies at the origin of the drag-based thrust. Using RFT, Gray and Hancock obtained (for the case of large-amplitude displacements) a closed-form solution for the swimming speed of an undulating filament given by the expression U = \u03c0c(a/\u03bb )2(CN/CT \u2212 1)/ (1 + 2\u03c02(CN/CT )(a/\u03bb )2)", + " Elastic effects are expected to dominate for El > 1. Because of the nonlinear (squared) dependence of El on the (swimmer) length scale L, one anticipates the effects of fluid elasticity to become increasingly important for swimming microorganisms. In 1979, Chaudhury [87] attempted to incorporate the effects of fluid elasticity on swimming using a second-order fluid and a series of expansions similar to Taylor\u2019s analysis. It was then predicted that fluid elasticity could either increase or decrease the propulsion speed of the waving sheet (Fig. 7.2), depending on the value of Re. Later, inspired by experimental observations of spermatozoa swimming in mucus [81, 97], the effects of elasticity on beating flagellar structures were considered in Stokes flow using the Maxwell model [104]. It was shown that self-propulsion was not affected by viscoelasticity even at large Deborah numbers (De), where De = \u03bb f and \u03bb is the fluid relaxation time and f is the beating frequency. However, the total work decreased with increasing De. It was then suggested that a microorganism could swim faster in a viscoelastic fluid with the same expenditure of energy compared with a Newtonian fluid. More recently, Lauga [83] showed that, for a 2D waving sheet (Fig. 7.2), elastic stresses could significantly alter the organism speed and the work required to achieve net motion. Using nonlinear viscoelastic fluid models such as the Oldroyd-B and the FENE-P models (see Chap. 1), Lauga [83] showed that the sheet\u2019s forward speed U in a purely elastic fluids is given by U UN = 1+De2(\u03b7s/\u03b7) 1+De2 , (7.2) where UN is the swimming speed of the sheet in a viscous Newtonian fluid (i.e., Taylor\u2019s original result) and \u03b7s is the solvent viscosity; note that the solution viscosity \u03b7 is assumed to be the sum of the solvent viscosity and the polymer viscosity such that \u03b7 = \u03b7s + \u03b7p" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002165_j.matchar.2021.111183-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002165_j.matchar.2021.111183-Figure2-1.png", + "caption": "Fig. 2. Dimensions of the fatigue short crack sample (unit: mm, thickness = 5 mm).", + "texts": [ + " The short fatigue crack samples were prepared according to the standard of ASTM E467-13, whose size was 62.5 \u00d7 40 \u00d7 5 mm3, as shown in Fig. 1. For simplification, the horizontal and vertical samples were named as H-sample and V-sample, respectively. Each direction was tested with three samples. The short fatigue crack growth tests were carried out by Instron 8801 with a stress of 4 kN, a frequency of 30 Hz, a stress ratio (R) of 0.1, and at room temperature according to the standard of ASTM E467-13. The dimension of short fatigue crack samples was shown in Fig. 2. The fatigue test was employed with a constant \u0394K = 6.5 MPa m1/2, and the fatigue test was stopped when the crack reached 2 mm to examine the sample through EBSD. The crack open displacement (COD) was used to test and calculate the short fatigue crack length. In addition, the three forged samples were employed to short fatigue crack growth tests along the forged sample. The test conditions kept the same as the WAAMed part. The microstructure observation surface was XZ plane, and the samples were etched with 1 mL HF, 6 mL HNO3 and 100 mL H2O for 30 s" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000255_s10033-017-0199-9-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000255_s10033-017-0199-9-Figure1-1.png", + "caption": "Figure 1 Schematic of a general passive overconstrained PM", + "texts": [ + " At present, the approaches proposed to solve this problem have different characteristics. It is difficult to quickly find a suitable method for force analysis of the corresponding overconstrained system. In this paper, the methods for force analysis of both active overconstrained and passive overconstrained PMs are reviewed and discussed in detail, to provide an important reference for researchers and engineers who would like to solve the statically indeterminate problem of overconstrained systems. The schematic of a general passive overconstrained PM with n DOFs is shown in Fig. 1. Assume that the t supporting limbs supply m constraint forces/moments to the moving platform in total. For a passive overconstrained PM, there exists m[ 6\u2013n. Let At, Bt, Ct, \u2026, denote the joints of the tth (t = 1, 2, \u2026, t) supporting limb from the moving platform to the base in sequence. Assume that the friction in the kinematic joints is ignored, and the stiffness of the moving platform is much greater than that of the supporting limbs. Owing to the existence of redundant constraints, the force and moment equilibrium equations of a passive overconstrained PM are insufficient to determine all the driving forces/torques and constraint forces/moments", + " Discussion: Based on the constraint Jacobian matrix of a passive overconstrained mechanism, several methods were proposed to isolate the joint reactions that can be uniquely determined. Those methods were proposed from a purely mathematical perspective, i.e., the corresponding physical interpretation was not considered. Besides, the analytical expressions of joint reactions cannot be obtained by this kind of method. Main ideas: Assuming that the tth (t = 1, 2, \u2026, t) supporting limb of a passive overconstrained PM contains Nt driving forces/torques and constraint forces/moments in total, as shown in Fig. 1, the elastic deformations generated at the end of the tth limb by the Nt driving forces/torques and constraint forces/moments are considered to be decoupled to each other [53\u201356]. In this case, the stiffness of each supporting limb can be expressed as a scalar quantity or a diagonal matrix. The steps of this method can be summarized as follows: The force and moment equilibrium equations of the moving platform of a passive overconstrained PM can be formulated as 6SF\u00f0 \u00de6 1\u00bc G6 n\u00fem\u00f0 \u00def n\u00fem\u00f0 \u00de 1; \u00f07\u00de where 6SF denotes the six-dimensional external load imposed on the moving platform, G is the coefficient matrix mapping the driving forces/torques and constraint forces/moments to the external loads, and f is the vector composed of the magnitudes of the n driving forces/torques and m constraint forces/moments" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001810_s40430-020-02642-6-Figure20-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001810_s40430-020-02642-6-Figure20-1.png", + "caption": "Fig. 20 Offset in planet gear", + "texts": [ + "\u00a017, where the sun gear center has an offset from its original center (Of). In such a condition, the transformation matrix of Eq.\u00a0(8a) can be replaced by: where \u03b4es is the offset error. Figure\u00a018 shows the resulting kinematic error with an offset error of \u03b4es = 0.01\u00a0mm. It can be also noted that the kinematic error caused by a sun gear offset error can be reduced when the speed reduction ratio increases, as shown in Fig.\u00a019a,b, where the maximum kinematic errors are 6 and 3 arcsec, and where the speed reduction ratio m are 40 and 79 for Fig.\u00a019a,b, respectively (Fig.\u00a020). Similar to the sun gear, the offset error in the planet gear may also result in kinematic errors. As shown in Fig.\u00a019, the center of the planet gear has an offset distance \u03b4ep from its ideal center O2\u2032. This error can be simulated by replacing the transformation matrix Eq.\u00a0(8b) with: Figure\u00a021 shows the kinematic error resulting from an offset error in the planet gear of \u03b4ep = 0.01\u00a0mm. Likewise, (28) f1 = \u23a1\u23a2\u23a2\u23a2\u23a3 cos 1 sin 1 0 es cos 1 \u2212 sin 1 cos 1 0 \u2212 es sin 1 0 0 1 0 0 0 0 1 \u23a4\u23a5\u23a5\u23a5\u23a6 , (29) f2 = \u23a1 \u23a2\u23a2\u23a2\u23a3 cos 2 \u2212 sin 2 0 ( 1 + 2) cos out + ep cos 2 sin 2 cos 2 0 \u2212( 1 + 2) sin out+ ep sin 2 0 0 1 0 0 0 0 1 \u23a4 \u23a5\u23a5\u23a5\u23a6 " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000626_02640414.2017.1329547-Figure5-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000626_02640414.2017.1329547-Figure5-1.png", + "caption": "Figure 5. A sagittal plane view of the seam elevation angle, whereby \u03c6Seam is the angle measured between the plane of the seam and the horizontal plane of the global CS.", + "texts": [ + " The spin axis\u2019 elevation angle \u03c6~\u03c9 was the angular excursion of the ball\u2019s angular velocity vector from the horizontal (x-y) plane (90\u00b0 = pointing up; 0\u00b0 = coincident with horizontal; \u221290\u00b0 = pointing down) (Figure 4). Figure 2. Static ball markers attached evenly around the plane of the seam. Novel to this study were the calculations of the seam relative to the global coordinate system and the spin axis. The seam elevation angle \u03c6Seam was measured as the angle between the plane of the seam and the horizontal (x-y) plane (0\u00b0 = seam plane \u201cflat\u201d and coincident with horizontal; 90\u00b0 = seam plane \u201cupright\u201d and perpendicular to horizontal) (Figure 5). After defining a vector originating at the centre of the ball and directed toward the most superior aspect of the seam, the seam azimuth angle (\u03b8Seam) was calculated as the excursion of this vector from the x-axis of the global CS, in the horizontal plane (Figure 6). The acute angle between the ball\u2019s angular velocity vector and the plane of the seam was used to define seam stability (\u03c9Seam) (Figure 7) and expressed as a percentage value according to: A value of 100 denoted the theoretical condition of perfect stability, whereby the ball\u2019s x-axis and angular velocity vector were coincident" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002401_j.mechmachtheory.2021.104371-Figure17-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002401_j.mechmachtheory.2021.104371-Figure17-1.png", + "caption": "Fig. 17. Installation diagram of a vibration displacement sensor.", + "texts": [ + " Under different working conditions, the lateral vibration displacement RMS and errors of the experiment test and the simulation results of gear are shown in Table 1 . The variation trend of the simulation results is basically the same as that of the test results. The maximum relative error between the simulation results and the test results is 12.79% in each condition. Vibration displacement sensors were installed on the boxes closest to the edges of the input gear and the output gear. The sensors were arranged in the horizontal direction (Z-axis direction) of each gear, as shown in Fig. 17 . Five equidistant measurement points were chosen along the tooth width direction (Z-axis direction) at the root of the tooth, and the strain gauges were pasted evenly along the tooth width direction, as shown in Fig. 18 . The direction of the resistance wire of each strain gauge is consistent with the direction of tooth root deformation. Under steady-state working conditions with an input speed of 1800 r/min and a load torque of 500 Nm, the swing vibration displacement signals of the unmodified gear pair and the modified gear pair were collected and analysed" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000769_978-3-662-43893-0-Figure3.13-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000769_978-3-662-43893-0-Figure3.13-1.png", + "caption": "Fig. 3.13 Transmission error curves of three consecutive tooth pairs", + "texts": [ + " This difference may be calculated if the associated pinion rotation angle, for a given wheel rotation angle, is corrected by an angular difference \u03c6korr until the pair of tooth flanks meet exactly at one point. \u03c6korr, i \u00bc \u03c62, i z1 z2 \u03c61, i \u00f03:33\u00de This calculation of the difference angle of rotation is done iteratively and is performed for each mesh position i. This yields the variation in the angle of rotation or transmission error. If pitch deviations are neglected, the transmission error of each tooth pair of a given gear set has the same shape. The transmission error curves of neighboring tooth pairs are then translated one pitch apart, as shown in Fig. 3.13. 76 3 Design Contact pattern calculation The contact pattern is a representation of all the contact lines during a complete mesh of a tooth pair. Load-free motion produces a different contact pattern than motion under load. The contact pattern which occurs under load is one result of the load distribution calculation (see Sect. 4.4.3.4). The following points should be considered when calculating the load-free contact pattern: \u2013 the tooth contour limiting the contact pattern, \u2013 the simultaneous engagement of several teeth, \u2013 specified or calculated deviations in relative position Other parameters which may be considered are: \u2013 pre-mesh or edge contact, \u2013 specified pitch deviations To determine the contact pattern, the bevel gear set rotates in such a way that at each moment one tooth flank pair is in contact" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001121_j.triboint.2018.06.035-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001121_j.triboint.2018.06.035-Figure1-1.png", + "caption": "Fig. 1. Schematic of the simulated bodies in contact.", + "texts": [ + " The analysis highlights transient effects in the lubricated finite line EHL contact that become especially important to consider while analysing systems that are continuously subjected to transient loading conditions during operation. The methodology and model described in this section can be used to study finite length EHL contacts under transient conditions. In the model, the contact simulated represents the contact between a rolling element with logarithmic crowning and the inner ring in a cylindrical roller bearing, shown schematically in Fig. 1. In order to simplify the model, the contacting bodies are described as an equivalent geometry, where the influence of the chamfered zone on the inner ring is neglected and the rounded corner on the rolling element is not included. Instead the logarithmic crowning is assumed to reach all the way out to the end of the roller. Moreover, isothermal conditions, Newtonian fluid behaviour and smooth surfaces are assumed, which are all parameters that have been studied and included in simulation models presented in literature" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000999_iet-cta.2016.1277-Figure7-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000999_iet-cta.2016.1277-Figure7-1.png", + "caption": "Fig. 7 Coupled inverted pendulums", + "texts": [ + " We choose the state vectors as x1(t) = \u03b81(t) \u03b8\u03071(t) \u22a4 x2(t) = \u03b82(t) \u03b8\u03072(t) \u22a4 Then, the systems can be described by dx1(t) dt = 0 1 g l 0 + \u0394A1(\u03c51, t) x1(t) + 0 1 m1 \u2217l2 + \u0394B1(\u03bd1, t) \u03931(u1(t)) + \u2211 j = 1 2 A1 j(\u03be1, t)x j(t \u2212 h1 j) + q1(\u03b61, t) (50) dx2(t) dt = 0 1 g l 0 + \u0394A2(\u03c52, t) x2(t) + 0 1 m2 \u2217l2 + \u0394B2(\u03bd2, t) \u03932(u2(t)) + \u2211 j = 1 2 A2 j(\u03be2, t)x j(t \u2212 h2 j) + q2(\u03b62, t) (51) where \u0394Ai(\u03c5i, t) = 0 1 mi \u2217l2 \u2212 mi \u2217 mi(t) k(t)a2(t) 0 := Bi \u03c5i(t) 0 \u0394Bi(\u03bdi, t) = 0 1 mi \u2217l2 \u2212 \u0394mi(t) mi(t) := Bi \u03bdi(t) Ai j(\u03bei, t) = 0 1 mi \u2217l2 mi \u2217 mi(t) k(t)a2(t) 0 := Bi \u03bei(t) 0 qi(\u03b6i, t) = 0 \u03b6i(t) (52) where \u03c5i(t) := \u2212 mi \u2217 mi(t) k(t)a2(t) \u03bdi(t) := \u2212 \u0394mi(t) mi(t) \u03bei(t) := mi \u2217 mi(t) k(t)a2(t) and where qi( \u22c5 ) is any external disturbances. It should be pointed out that the uncertain interconnected time-delay inverted pendulum systems described by (50) and (51) are closer to practical problem 1368 IET Control Theory Appl., 2017, Vol. 11 Iss. 9, pp. 1360-1370 \u00a9 The Institution of Engineering and Technology 2017 than the uncertain interconnected inverted pendulum systems considered in [30\u201333], because some degree of time lags and deadzone inputs always exist in the practical control systems (Fig. 7). Actually, in the case that hi j = 0 and \u0393i(ui) = ui, i, j = 1, 2, the systems described by (50) and (51) will be reduced to the uncertain systems in [30\u201333]. Here, for some given constants \u03b71 = \u03b72 = 1.0, and the matrices Q1 = Q2 = diag{8, 8}, it can be obtained that Pi = 12.0 4.0 4.0 4.0 , i = 1, 2 For the decentralised local adaptation laws, we select the following parameters: \u03c3i = 0.1, \u03b3i = 0.5, i = 1, 2 Thus, for the large-scale interconnected time-delay system given in (50) and (51) with the uncertainties and external disturbances, from (9) and (10) we can obtain the decentralised local adaptive robust state feedback controllers, which can guarantee the uniform ultimate boundedness of the uncertain large-scale systems" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000894_tro.2015.2489499-Figure15-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000894_tro.2015.2489499-Figure15-1.png", + "caption": "Fig. 15. Orthoglide and its hidden robot leg. (a) Kinematic chain. (b) The hidden robot leg.", + "texts": [ + " Robots of this category are those for which there exists a global diffeomorphism between the leg direction space and Cartesian space for all workspace configurations. Their hidden robot models have only one possible assembly mode. Once again, presenting an exhaustive list of robots of this category is totally impossible because it requires the analysis of the assembly modes of all hidden robot models for each robot architecture. However, we show here for the first time robots belonging to this category. Let us consider the Orthoglide [44] designed at IRCCyN [see Fig. 15(a)]. This robot is a mechanism with three translational dof of the platform. It is composed of three identical legs made of PR\u03a0R architecture, or also with PUU architecture, the P joint of each leg being orthogonal. Let us consider the second type of leg which is simpler to analyze (even if the following results are also true for the first type of leg). If the link between the two passive U joints is observed, from Section III, the hidden robot leg has a PUU architecture with, of course, two degrees of actuation. As a result, for controlling the three dof of the platform, only two legs need to be observed. For a fixed configuration of the actuated U joint, each leg tip has the possibility to freely move on a line directed along the corresponding P joint direction: This line corresponds to the free motion of the platform due to the virtual passive P joint of each leg, when other legs are disconnected [see Fig. 15(b)]. Then, estimating the robot pose is equivalent to finding the intersection of two lines in space (three lines if the three legs are observed). As a result, in a general manner, the forward kinematic problem (fkp) may have: 1) zero solutions (impossible in reality due to the robot geometric constraints); 2) an infinity of solutions if and only if the P joints are parallel (not possible for the Orthoglide as all P joints are orthogonal); 3) one solution (the only possibility). Moreover, a simple singularity analysis of all the possible hidden robot models of the Orthoglide could show that they have no Type 2 singularities (which is coherent with the fact that the fkp has only one solution)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000769_978-3-662-43893-0-Figure3.4-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000769_978-3-662-43893-0-Figure3.4-1.png", + "caption": "Fig. 3.4 Cut-off or fin (schematic on a virtual crown gear)", + "texts": [ + " With spiral bevel gears, and especially with flat wheels, the cutter head may hit the gear rim at another point than the intended tooth slot, thus damaging the teeth. This effect, named as back cutting or interference at the rear, is more likely to occur with a diminishing tool radius. However, it depends on many parameters and must therefore be calculated differently for each manufacturing method. 64 3 Design Apart from leaving a fin on the bottom of the tooth slot, cut-off may also occur on the tooth flanks (see Fig. 3.4). This is caused by the variable width of the tooth slot in the normal section, creating a minimum at one location and a maximum at another which are not necessarily at the inner and outer ends of the slot. To avoid cut-off, the point width sa0 of the tool must be smaller than or equal to the slot width at its narrowest point. On the other hand, the largest slot width efn,max must be covered completely by the point widths of all blades in a group. A feature to be noted is that a number of manufacturing methods use standardized tools whose point widths cannot simply be changed as can be done when using stick blades or grinding wheels" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001390_icuas.2016.7502680-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001390_icuas.2016.7502680-Figure1-1.png", + "caption": "Fig. 1: The goal of MORUS project is to design a cooperative autonomous robotic system comprised of a UAV and an AUV, working together on maritime inspection.", + "texts": [ + ". INTRODUCTION The work presented in this paper fits within the scope of an ongoing project called MORUS that aims to build a cooperative autonomous robotic system able to work both in air and underwater (Fig 1). To that end an unmanned aerial vehicle (UAV) and an autonomous underwater vehicle (AUV) are brought to work together on a common goal, maritime surveillance. In the envisioned scenario, the UAV has to be capable of lifting the AUV and carrying it to a designated drop-off point. Ever since they were introduced to the mainstream [1], quadrotors have emerged as a number one research platform used in aerial robotics. Since then, numerous linear and nonlinear control algorithms have been proposed and tested", + "00 \u00a92016 IEEE 1327 Moving Mass Control (MMC) is a concept that relies on the change of CoG of the vehicle to ultimately distribute torque around the body in order to control its attitude. In total, a mutlirotor has 6 degrees of freedom (DOF), and in our implementation MMC controls only two of them, roll and pitch angle. Since the underlining physics of MMC cannot effectively control its other degrees of freedom, classical rotor speed control is applied to yaw angle and height control. The only two DOFs left (i.e. x-axis and y-axis) are controlled through high level controllers, and thus fall out of scope of this paper. The control structure is presented in Fig 1. The vehicle consists of a static quadrotor body that holds everything together. On top of that, there are 4 rotors symmetrically placed around the central body in a pattern known as + configuration. For an ideal configuration we can bring this components in a common denominator mass mb. Building upon this classic quadrotor construction we add 4 moving masses, placed within each UAV arm, that are controlled to move linearly and thus vary the overall CoG of the UAV. Given that each body part i\u2019s displacement from the body frame is denoted as roi and observed in the body frame L0, one can write the CoG equation: ro,c = mbro,b + \u2211 4 i=1 miroi mb + \u2211 4 i=1 mi = \u2211 4 i=1 miroi M , (1) where mb is the mass of the quadrotor rigid body (without moving masses), and M is the total mass of the vehicle" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002574_pime_auto_1949_000_009_02-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002574_pime_auto_1949_000_009_02-Figure1-1.png", + "caption": "Fig. 1. Typical Constant-load Synchronizer", + "texts": [ + "comDownloaded from THE T H E O R Y OF G E A R - C H A N G I N G 7'2 CPIGP.GE - -___ 31 u Disengaged position. Cone clutch CC is free : balls rest in detent groove in clutch ring R. b Engaged position. Cone clutch has functioned and is about to become free again : balls now rest in smaller detent in clutch ring R. The meaning of the foregoing terms is self-evident, with the exception of those specifically referring to synchronizing devices. These are best explained by reference to a particular example rather than by formal definition. Examples of Synchronizers. Fig. 1 shows a simple constantload synchronizer from a Panther tank, but of the type fitted to many passenger cars, particularly of the cheaper varieties. I t consists of the gear to be engaged G mounted freely upon its shaft and having external involute dogs formed integrally at one side; a bronze cone clutch, CC, attached by a circlip to the gear G, and driven by four square-cut jaws in its rear end-face which engagc conjugate recesses in the gear; a hardened steel external clutch C, having splined connexion to its shaft, an internal cme to suit the mating cone clutch, and external involute dogs j and a hardened steel operating ring R actuated by conventional fork mechanism and having internal involute dogs to suit the mating parts C, G", + " baulk and any cone angle due to crashing through. With 60 deg. baulk angle, some sticking occurred with 5 deg. cone, but no crashing; 10 deg. cone angle was too readily crashed and 75 deg. cone angle was generally much the best but could be crashed by attempting a very rapid change. at Gazi University on January 15, 2016pad.sagepub.comDownloaded from THE THEORY O F E X A M P L E 4. F A I L U R E T O C H A N G E DUE T O LOSS O F WAY Constant-load synchronizer, \u201cZahnfabrik Friedrichshafen\u201d gearbox for German Panther tank (see Fig. 1). Data. Vehicle weight, W . . 98,560 lb. Governed engine speed . . 3,000 r.p.m. Fixed reduction after gearbox . 12.42 Sprocket pitch circle diameter . 32.5 inches Mean diameter of synchro-cones . 5 4 inches Total included angle of cone, 2a . 13 deg. Total tractive demand on good, level, cross-country, p = (0*0545+0.00145V) where V = speed in m.p.h. By experiment the axial force to overcome the synchronizer spring baulk = 245 Ib. On the assumption that p = 0.08, T, = 468 1b.-in., and from equation (8), Table 9 is compiled neglecting drag torque" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001924_j.jmps.2020.103959-Figure5-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001924_j.jmps.2020.103959-Figure5-1.png", + "caption": "Fig. 5. Definition of coordinates \u03b8 and r : (a) Euler coordinate system ( \u03b8 , \u03c9, r ) for a spherical shell; (b) modified Euler coordinate system ( \u03b8 , \u03c9, r ) for shell with large imperfection. The thick line refers to the shell middle surface.", + "texts": [ + " We assume the constitutive relation to be linear due to the small strains involved, and present a formulation that enables us to write the shell buckling equations without any restrictions on the magnitude of displacements and rotations ( Hutchinson, 2016 ; Niordson, 1985 ). Furthermore, we follow the tensor analysis given by Niordson (1985) as well as Koiter and van der Heijden (2009) (see Appendix B ) to derive the nonlinear buckling equations of the middle surface for axisymmetric deformations. The following section presents first the theory for spherical shells with perfect geometry, while Section 3.2.2 provides the formulation for shells with large imperfection. Appendix C reports the numerical method used to obtain the solutions. Fig. 5 a shows the Euler coordinates ( \u03b8 , \u03c9, r ) for a perfect spherical shell. \u03b8 is the meridional angle, \u03c9 is the circumferential angle (not shown), and r is the distance from the origin O 1 . The meridional angle \u03b8 is measured from the equator ( \u03b8 = 0 ) to the upper pole ( \u03b8 = \u03c0/ 2 ). R is the radius of the shell and ( \u03b8 , \u03c9, R ) represents the coordinates of a material point on the middle surface of the shell. For the deformed shell, the location of a material point on the middle surface is r\u0304 = u \u03b8 i \u03b8 + u \u03c9 i \u03c9 + ( R + w ) i r (2) where ( u \u03b8 , u \u03c9 , w ) are the displacements tangent and normal to the undeformed middle surface, and ( i \u03b8 , i \u03c9 , i r ) are the corresponding unit vectors", + " The resultant membrane stresses ( N \u03b8\u03b8 , N \u03c9\u03c9 , N \u03b8\u03c9 ) and the bending moments ( M \u03b8\u03b8 , M \u03c9\u03c9 , M \u03b8\u03c9 ) for a shell with isotropic linear elastic material are N \u03b1\u03b2 = Et ( 1 \u2212 v 2 )[ (1 \u2212 v ) E \u03b1\u03b2 + v E \u03b3 \u03b3 \u03b4\u03b1\u03b2 ] M \u03b1\u03b2 = D [ (1 \u2212 v ) K \u03b1\u03b2 + v K \u03b3 \u03b3 \u03b4\u03b1\u03b2 ] (7) where E is the Young\u2019s modulus, t is the shell thickness, \u03bd is the Poisson\u2019s ratio, and D = E t 3 / [12(1 \u2212 \u03bd2 )] is the bending stiffness. The subscripts in Eq. (7) take on values 1 or 2, and the Einstein summation convention applies. The non-vanishing components in the membrane stress and bending moments are ( N \u03b8\u03b8 , N \u03c9\u03c9 ) and ( M \u03b8\u03b8 , M \u03c9\u03c9 ). The sum of the stretching and bending energy gives the elastic strain energy (SE) expressed as: SE ( u \u03b8 , w ) = 1 2 \u222b S ( M \u03b1\u03b2K \u03b1\u03b2 + N \u03b1\u03b2E \u03b1\u03b2 ) d S (8) where S is the area of the perfect spherical surface in its undeformed state ( Fig. 5 a). The potential energy of the uniform external pressure is PE = p V (9) where V is the volume change. For small deformations, the volume change can be approximated with the pressure acting on the initial middle surface in the direction normal to the initial middle surface, namely the dead pressure: V ( u \u03b8 , w ) = \u222b S w d S (10) For large axisymmetric deformations, the volume change V is obtained with the pressure acting on the deformed middle surface in the direction normal to the deformed middle surface, i", + ", 2016b ), the latter cannot be neglected for a shell with pole displacement comparable to its radius, which experiences large nonlinear deformation, as is the case of the shell examined here. The total potential energy of the spherical shell is given by the sum of the elastic strain energy SE and the potential energy PE of the external pressure ( u \u03b8 , w ) = SE + PE (12) To account for the large geometric imperfection (axisymmetric circular-arc), we introduce the modified Euler coordinates ( \u03b8 , \u03c9, r ) shown in Fig. 5 b. Here, the difference from the Euler coordinate system ( Fig. 5 a) pertains to the coordinates \u03b8 and r , which are defined with respect to the center of the imperfection O 2 , rather than the origin of the Euler coordinate system O 1 . Following the geometry parameters introduced in Section 2 , the coordinates of O 2 ( \u03b82 , \u03c9, R 2 ) are given by \u03b82 = \u03b8m R 2 = R cos \u03b8w 2 \u2212 2 R sin \u03b8w 2 h l + R sin \u03b8w 2 ( l 4 h + h l ) (13) In the general case, the coordinates of a material point on the middle surface of the imperfection is ( \u03b8 , \u03c9, R I ), where the radius of the imperfection, R I ( Fig. 5 b), is given by: R I = R sin \u03b8w 2 ( l 4 h + h l ) (14) The location of a material point on the deformed middle surface of the imperfection is r\u0304 = R 2 i 2 + u \u03b8 i \u03b8 + u \u03c9 i \u03c9 + ( R I + w ) i r (15) where R 2 i 2 is the location of the center of the imperfection, ( u \u03b8 , u \u03c9 , w ) are the displacements tangent and normal to the undeformed middle surface, and ( i \u03b8 , i \u03c9 , i r ) are the corresponding unit vectors. For axisymmetric deformations, the circumferential displacement equals to zero u \u03c9 = 0 , and the other displacements, u \u03b8 and w , are independent of the circumferential angle \u03c9" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001613_j.engfailanal.2019.06.084-Figure8-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001613_j.engfailanal.2019.06.084-Figure8-1.png", + "caption": "Fig. 8. Propeller shaft phase out", + "texts": [ + " Table 1 Bench tests and vehicle tests Bench tests Vehicle tests \u03bc-comp synchronizer endurance test Vehicle GSD test Transmission GSD Test Driver driving pattern measurement TV test rig Vehicle angular acceleration measurement Engine Dyno test High way durability test Fig. 5. \u03bc-comp test rig setup The TV (Torsional Vibration) test rig consists of gearbox, drive motor, propeller shaft & adopter pivot as shown in Fig. 7. The clutch disc is mounted on the gearbox input shaft. The electric motor is connected to the gearbox output shaft through the coupling and the propeller shaft. The adopter plate to the mount is offset by 15\u00b0-20\u00b0 on the vertical axis and 10\u00b0-15\u00b0 in the horizontal axis. The propeller shaft yokes have a phase difference of 90\u00b0 as shown in Fig. 8. This gives a phase difference during shaft rotation. A magnetic pickup speed sensor is used to measure the input shaft speed. The acceleration is processed using sensor electronics. The gearbox is engaged in the 5th gear position (1:1 gear ratio). The motor is continuously driven at 1100 rpm. This rpm with propeller shaft phase difference creates 2000 rad/s2 angular acceleration on the input shaft. This set-up vibrates the synchronizer ring assembly. The synchronizer life in the vehicle can be directly correlated with the TV test rig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000525_s12541-016-0036-6-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000525_s12541-016-0036-6-Figure3-1.png", + "caption": "Fig. 3 Blades and generating cones for gear generating tool: (a) Generating tool cones for concave side and (b) Generating tool cones for convex side", + "texts": [ + " 2, the cutting edge of head-cutter blade is divided into four sections as the edge, toprem, profile, flankrem. The cutting surfaces of the head-cutter are generated by rotation of the blade about the Zt-axis, the rotation angle is \u03b8g. Most of the generating of gear tooth surface is done by the profile section that is a straight line with the profile angle \u03b1g. The fillet of the gear tooth surface is generated by the edge section with corner radius \u03c1w. Referring to Fig. 2, an arbitrary point Mb on the cutting surface of blade is determined by ug and g (Fig. 3). The generating surface \u03a3g about the profile section of the head-cutter blade is represented by vector function rt(ug, \u03b8g) as (2) Where, rG is the head-cutter point radius. \u03b1g is the blade angle. ug and \u03b8g are tooth surface parameters. The upper signs in Eq. (2) correspond Ra \u03c9g \u03c9c ----- d\u03c8g dt -------- d\u03c8c dt --------\u2044= = rt ug \u03b8g,( ) rG ug \u03b1gsin\u00b1( ) \u03b8gcos rG ug \u03b1gsin\u00b1( ) \u03b8gsin ug \u03b1gcos\u2013 = to generation of concave side of the gear tooth surface, while the lower signs correspond to convex side" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000723_0959651818791027-Figure4-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000723_0959651818791027-Figure4-1.png", + "caption": "Figure 4. A gantry crane with flexible cable.", + "texts": [ + " As a transportation tool for effectively moving massive payloads, crane systems are widely used in many industrial locations such as building sites, workshops, warehouses, and shipyards. Crane systems include the following systems: two-dimensional (2D) overhead cranes, three-dimensional (3D) overhead cranes, tower cranes, gantry cranes, and cart\u2013crane systems. A typical gantry crane system,43 as an example of a crane system, has a bottom payload, a flexible cable, and a top trolley, as shown in Figure 4. The main task of the gantry crane system with flexible cable is to deliver payloads quickly and accurately from an initial position to the desired position. In this process, the payload swing must be as small as possible to avoid collisions with other objects or personnel around the crane. The dynamic model is written as r x\u00f0 \u00de\u20acy x, t\u00f0 \u00de T x, t\u00f0 \u00dey00 x, t\u00f0 \u00de T0 x, t\u00f0 \u00dey0 x, t\u00f0 \u00de l0 x\u00f0 \u00dey0 x, t\u00f0 \u00de3 3l x\u00f0 \u00dey0 x, t\u00f0 \u00de2y00 x, t\u00f0 \u00de=0 m\u20acy 0, t\u00f0 \u00de T 0, t\u00f0 \u00dey0 0, t\u00f0 \u00de l 0\u00f0 \u00dey0 0, t\u00f0 \u00de3 d1 t\u00f0 \u00de=u1 t\u00f0 \u00de \u00f012\u00de M\u20acy L, t\u00f0 \u00de+T L, t\u00f0 \u00dey0 L, t\u00f0 \u00de+ l L\u00f0 \u00dey0 L, t\u00f0 \u00de3 d2 t\u00f0 \u00de= u2 t\u00f0 \u00de where T(x, t)=T0(x)+ l(x)y0(x, t)2, ( )0= \u2202( )=\u2202x, and ( )= \u2202( )=\u2202t" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002064_tte.2020.3047516-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002064_tte.2020.3047516-Figure1-1.png", + "caption": "Fig. 1. 3-D view of the proposed MCTLSRM.", + "texts": [], + "surrounding_texts": [ + "2332-7782 (c) 2020 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.\nIndex Terms\u2014Modular and lateral translator, crooked-tooth angle, linear switched reluctance motor.\nI. INTRODUCTION\nIN later decades, the requirement of taller buildings and dense cities have heightened due to the urban migration increasing rate. Elevators are among the foremost components of the skyscrapers, and they have to be designed and implemented efficiently to satisfy the demands of both building users and owners. However, conventional elevator technology hinders the practical construction and usage of buildings, since they take quite a few of the time to reach their target and occupy a significant area of each floor space. To overcome the conventional elevators\u2019 problems, the rope-free multidirectional elevator systems, called MULTI, have already been presented. This system benefits from the decreasing elevator wait times, optimizing cost, and increasing energy efficiency. Unlike the conventional elevator system that uses a rotary electric motor and mechanical interfaces, the MULTI employs a linear electric motor to provide the required propulsion force.\nGenerally, the essential required specifications for an electric motor applied in MULTI are smooth motion, high thrust density, and inherent fault tolerance. Although the linear permanent-magnet synchronous motors (LPMSMs) and linear induction motors (LIMs) have been employed for driving the MULTI so far, high-price of the LPMSMs and temperature\nMatin Vatani and Mojtaba Mirsalim are with the Electrical Machines and Transformers Research Laboratory (EMTRL), Department of Electrical Engineering, Amirkabir University of Technology (Tehran Polytechnic), Tehran 15916, Iran. E-mail: (Matinvatani@aut.ac.ir and Mirsalim@aut.ac.ir).\nissues of the LIMs confine their usages [1]\u2013[4]. Recently, linear switched reluctance motors (LSRMs) have become a noticeable candidate for the rope-free elevators applications. The LSRMs offer low manufacturing cost, high fault tolerance, simple cooling, robust structure, and high-speed ranges [4]\u2013 [7]. However, they still suffer from low power density and high thrust ripple. Therefore, this article aims to introduce a new enhanced-thrust LSRM for MULTI applications. To this end, the literature of LSRMs and rotary SRMs (RSRMs) concerning new topologies are reviewed, and the features as well as qualities of each topology are clarified. In the following, the literature review is classified into two staple assortments, including the literature of LSRMs and RSRMs.\nA novel high-thrust, double-sided, and segmental-translator LSRM is introduced in [8]. Although the structure increases the output thrust, it incurs the construction complexity of the motor. Reference [9] suggests a new LSRM topology for ship elevator applications. Because of the significant impacts of a lightweight translator on the performance of LSRMs, the authors have presented a yokeless translator. A comparative study of four longitudinal LSRM configurations through the finite element analysis (FEA) simulation and experimental results is done in reference [10]. The overall outcomes indicate that double-sided topology diminishes the normal forces between the translator and stator. Also, the segmental translator improves the thrust density compared to the non-segmental one. A new tubular LSRM (TLSRM) is recommended in [11], [12]. Although the output power and thrust are enhanced, it increases the manufacturing cost noticeably, especially in the case of long-distance propulsion applications. Reference [13] compares the magnetic characteristics of double-sided longitudinal flux LSRMs (LFLSRMs) and transverse flux LSRMs (TFLSRMs). The results demonstrate the superior operation of the LFLSRM over TFLSRM in terms of thrust density, power density, and efficiency. In [14], [15], the skewed teeth technique is used to mitigate the acoustic noise and stator deformation of LSRMs. However, it results in a decrease in the thrust and efficiency of the machine. A novel double-sided segmental stator with a toroidally wound translator LSRM is proposed, designed, optimized, and fabricated in [16]\u2013[18]. Their results are indicative of the improved flux linkage, thrust density, and thrust ripple of the motor than that of the conventional ones. Reference [19] proposes a new four-phase modular-stator LSRM (MSLSRM). A comprehensive study, including flux density, inductance waveforms, thrust density, and average thrust, is adopted between three structures. Which\nAuthorized licensed use limited to: Raytheon Technologies. Downloaded on May 19,2021 at 05:48:08 UTC from IEEE Xplore. Restrictions apply.", + "2332-7782 (c) 2020 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.\nthe advantages of MSLSRM over its counterparts are evident. Additionally, a portion of the literature is dedicated to designing procedures and optimization of the LSRMs [20]\u2013[22]. Furthermore, a sensitivity analysis of geometrical parameters on a bilateral LSRM under thermal constraints is performed in [23], [24].\nOwing to the identical essence of linear and rotary SRMs, the employed methods to improve the RSRMs specifications can be applied to the linear ones, likewise. Thus, the literature of RSRMs, especially innovative topologies, are further perused. A strategy to develop the output characteristics of RSRMs is shortening the length of the flux path, which is discussed in [25]\u2013[28]. The strategy decreases the requirement of the magnetomotive force to achieve the setpoint torque. So, it results in a notable reduction of copper wire consumption, core losses, and eccentric forces between stator and rotor poles in comparison to conventional RSRMs [29], [30]. Applying the modular structure to RSRMs (MRSRMs) is another vital method. Modular structure benefits the flux linkage augmentation, flux leakage reduction, being devoid of flux-reversal, and flux path length shortening. Also, thanks to the high faulttolerance capability, the MRSRMs are an appropriate applicant for use in safety-critical applications [31]\u2013[33].\nIn this paper, the modular structure is applied to reach the ultimate capability of the LSRM. Also, a new hypothesis, named crooked-tooth angle, is introduced results in an improvement of the machine characteristics. The main objective of this article is to present a novel double-sided, 8/6 pole number configurations, and modular as well as crooked-tooth translator linear switched reluctance motor (MCTLSRM) with high-thrust per weight capability and low thrust ripple.\nIn this paper the comprehensive research of the proposed LSRM, along with FEA and experimental results, are given. In section II, the proposed motor topology is presented, and an examination of crooked-tooth angle based on FEA is fulfilled. To obtain and compare the electromagnetic specifications of the MCTLSRM with its counterparts, a comparative study, including static and steady-state analysis using 2-D and 3- D FEA, is given in section III. A prototype of the proposed motor is manufactured, the experimental results are educed, and compared with those from FEA in section IV. Ultimately, a conclusion is given in section V.\nFigures 1 and 2 illustrate the 3-D and 2-D view of the proposed three-phase, longitudinal flux, double-sided, lateral, modular, and crooked pole translator linear switched reluctance motor (MCTLSRM), respectively. The motor is composed of six magnetically independent U-shaped modules as a translator, three modules on each side, and a classical doublesided stator. The concentric windings are individually wound around each module yoke, and facing modules are connected in series to form a phase. The main contributions of this article are:\n\u2022 Lateral translator: The translator of the MCTLSRM is located in the side of the motor, which is analogous to the outer rotor for the rotary SRMs. Moreover, owing to the lengthy vertical propulsion application of the MCTLSRM, windings are transferred from stator to the translator. Hence, it offers the advantages of economical copper wire consumption, weight reduction, and costeffectiveness. \u2022 Modular translator: According to the literature review of RSRMs, the modular structure offers lower magnetomotive force requirements, lesser copper loss, no fluxreversal, and shortening the flux track. Accordingly, the modular structure is exploited for the translator. \u2022 Translator crooked-tooth: The crooked-tooth angle is determined in Fig. 2. Deflecting the translator\u2019s tooth from the vertical axis, up to a specific degree, causes the improvements of motor characteristics in terms of flux linkage and average thrust.\nIn order to propose an enhanced-thrust structure with low thrust ripple, the advantages of applying modular structure on RSRM is combined with the new crooked-tooth translator strategy.\nThe pole number configurations of the MCTLSRM are identical to the 8/6 rotary SRM. It means that the least number of stator poles is eight. Whereas, in most cases, the stator is composed of a higher number of poles to satisfy the required length.\nAuthorized licensed use limited to: Raytheon Technologies. Downloaded on May 19,2021 at 05:48:08 UTC from IEEE Xplore. Restrictions apply.", + "2332-7782 (c) 2020 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.\nThis section deals with the design procedure of the proposed MCTLSRM, which is utilized to obtain the motor\u2019s dimensions. The following main steps play a crucial role in the design process:\n1) The main design parameters, including total translator width (Lt) and stack length, is calculated by the standard design formulas [20]. 2) The undefined parameters are reduced using the calculated Lt in step 1 and geometric relations among the motor parameters. 3) By employing the finite element method (FEM), a sensitivity analysis is adopted to obtain MCTLSRM\u2019s ultimate dimensions.\nFirst, the initial dimensions of the proposed MCTLSRM can be obtained from basic design formulas [20]. The total translator width can be can be obtained from:\nLt =\n\u221a P\u03c0\n2 (60\u03b7kdk1k2kBAsvm) , (1)\nwhere P is the rated power, \u03b7 is efficiency, kd is duty cycle, k1 = \u03c02\n120 , k2 is a variable dependant on the operating point, k is the stack length factor, B is secondary-side flux density, As is specific electric loading, and vm is the rated velocity. Also, the stack length can be calculated as:\nL = kLt. (2)\nSecond, to reduce the undefined parameters, the following assumptions and equations are developed. In the primary design procedure, the translator and stator pole width is assumed to be equal. Also, to avoid saturation, yoke heights are set equal to pole widths. Then, based on the assumptions and calculated value of Lt as well as \u03b1, the following equations among motor parameters can be derived.\nFor the proposed three-phase LSRM with a modular translator, pole configuration defines as follows:\n(6n+ 2) /6n where n = 1, 2, 3, ... (3)\nTo determine the maximum value of n, the obtained Lt from equation 1 has to be valid in the following equation:\nLt = 2 (11n\u2212 1)\n3 wtp +\n(19n\u2212 3)\n2 wts + 2nhtp tan\u03b1. (4)\nFor the next step, the area of the conductor, magnetic field intensity, and number of turns per module can be obtained as:\nHg = B\n\u00b50 (5)\nTmIn = Hg (2lg)\u21d2 Tm = Hg (2lg)\nIn , (6)\nwhere In is the rated current, so the required winding area can be achieved by:\nAw = Tm \u00b7 ac Ff = htp \u00b7 wts, (7)\nwhere Ff is the fill factor, and ac is the appropriate conductor\u2019s cross-section area for the specified rated current. Hence, the translator pole height is given by:\nhtp = Aw wts . (8)\nThe translator slot width can be calculated as:\nwts = 1\n6 (19n\u2212 3) ( 6Lt \u2212 4wtp (11n\u2212 1)+\u221a\n(4wtp (11n\u2212 1)\u2212 6Lt) 2 \u2212 144 (19n\u2212 3)nAw tan\u03b1 ) .\n(9)\nAlso, the stator slot width can be derived as:\nwss = 3\n2 wts + wtp \u2212 wsp. (10)\nAccording to the presented equations and assumptions, all MCTLSRM\u2019s parameters are related to wtp and \u03b1. In other words, by determining these two parameters, others can be calculated through equations 3 to 10.\nThird, a sensitivity analysis using the 2-D FEM can be adopted to select the translator pole width and crooked tooth angle values with the highest thrust producing capability. Figure 3a indicates that the proposed MCTLSRM produces the highest average thrust at wtp = 15mm and \u03b1 = 9degree. Moreover, to guarantee the superior performance of the obtained dimensions over other possible ones, a FEM based sensitivity analysis with the variation of translator and stator pole width has to be employed. Figure. 3b shows wtp = 15mm and wsp = 15mm result in superior MCTLSRM\u2019s output characteristics. Finally, the ultimate parameters\u2019 dimension and specification of the proposed motor are listed in Table I.\nAuthorized licensed use limited to: Raytheon Technologies. Downloaded on May 19,2021 at 05:48:08 UTC from IEEE Xplore. Restrictions apply." + ] + }, + { + "image_filename": "designv10_14_0001687_j.engfailanal.2020.104411-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001687_j.engfailanal.2020.104411-Figure1-1.png", + "caption": "Fig. 1. Technical drawing of the test gear (All dimensions are in mm) [8]", + "texts": [ + " [7] performed tensile tests on stainless martensitic steel samples fabricated using direct metal laser melting and found that post-production may be required for the mechanical and surface properties to meet a certain standard. Various tests were used to analyse the behaviour of gears subject to variable torque and rotation speeds during operation. The gears used in this study were operated on a wear test machine. The modulus of 420 steel gear was 1, whereas the number of teeth were 20. In addition, the pressure angle was 20\u00b0. Fig. 1 shows the technical drawing of the test gear. Production techniques, details and nomenclatures are given in Table 1. Four hundred twenty steel cold-drawn rods were shaved down to appropriate diameter using a lathe. The hob was threaded over the entire length to achieve the desired tolerance and quality. The gears were machined to suitable width, slots were made, and the rods were then ground down. The terms \u2018conventional\u2019 or \u2018traditional\u2019 production are used to describe the production technique of hobbing and machining without computer-aided design [8,9]. SolidWorks was used to draw 3D solid models based on the technical drawing in Fig. 1. After the design was completed, computeraided design models were converted into \u2018.stl\u2019 files, thus enabling the 3D model to be converted into G code using auxiliary software. The data of the ready-to-manufacture model was then transferred to an additive manufacturing machine. Production was made using DMLS production steps, which are shown in Fig. 2. The gears produced by additive manufacturing were positioned horizontally on the device table during production. The reason for the vertical positioning of the gears in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001808_j.jsv.2020.115766-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001808_j.jsv.2020.115766-Figure2-1.png", + "caption": "Fig. 2. Schematic drawing for the bearing waviness.", + "texts": [ + " 414 \u2217 10 12 F ( \u03c1) 4 \u2212 2 . 817 \u2217 10 12 F ( \u03c1) 3 + 2 . 806 \u2217 10 12 F ( \u03c1) 2 \u2212 1 . 398 \u2217 10 12 F ( \u03c1) + 2 . 784 \u2217 10 11 (8) Bearing waviness is an unavoidable imperfection as it comes from the machining process or even during bearing operation. Thus, it is important to consider its effect on the induced micro-vibrations. It is acceptable to describe the waviness as a sinusoidal wave [17,30] . Waviness may exist on all bearing surfaces (balls and inner/outer races) and in all directions (radial and axial) [12,14,17,30] . Fig. 2 is a schematic drawing that illustrates the waviness in bearing components in axial and radial directions. Fig. 3 shows a general wave representation on any bearing component, while Eq. (9) indicates its mathematical representation. w = A sin [ 2 \u03c0 ( j \u2212 1) N b \u2212 N w (\u03c9 race \u2212 \u03c9 c ) t ] (9) where, w is the amplitude of the waviness at any time t, A is the maximum amplitude of the waviness, N b is the total number of balls of the bearing, j is the number of the ball in contact with the bearing race, N w the waviness order, \u03c9 race is the angular velocity of the bearing races, and \u03c9 c is the angular velocity of the cage" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000301_j.nahs.2018.05.008-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000301_j.nahs.2018.05.008-Figure1-1.png", + "caption": "Fig. 1. Two inverted pendulums connected by a spring.", + "texts": [ + " To solve this problem, a new single Lyapunov function method combined with dynamic gain technique is proposed in this paper, which not only reduces the conservativeness of control methodology but also simplifies the form of controllers of subsystems. 4. Illustrative examples In this section, two examples will be presented to illustrate the effectiveness of the proposed control scheme. Example 4.1. As shown in [30],we consider amechanical systemof two inverted pendulums connected by a spring exhibited in Fig. 1. The equations of motion for the pendulums are described by x\u03071 = x2, x2 = (m1gr J1 \u2212 kr2 4J1 ) sin x1 + kr 2J1 (l \u2212 b) + u1 J1 + kr2 4J1 sin x3, x\u03073 = x4, x\u03074 = (m2gr J2 \u2212 kr2 4J2 ) sin x3 + kr 2J2 (l \u2212 b) + u2 J2 + kr2 4J2 sin x2, y = x1, (41) where (x1, x3)T = (q1, q2)T and (x2, x4)T = (q\u03071, q\u03072)T are the angular displacements of the pendulums from vertical and angular rates, respectively. m1 and m2 are the pendulum end masses, J1 and J2 are the moments of inertia, k is the spring constant of the connecting spring, r is the pendulum height, l is the natural length of the spring, b is the distance between the pendulum hinges, and g is gravitational acceleration" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001872_tmag.2019.2947611-Figure17-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001872_tmag.2019.2947611-Figure17-1.png", + "caption": "Fig. 17. Fields of the proposed HPDPMSM. (a) Flux line at rated torque. (b) Flux density at rated torque. (c) Flux density at maximum speed. (d) Flux line at maximum speed.", + "texts": [ + "18% and 9.73% respectively. Fig.16 shows the rated output torque of two HPDPMSMs. The average output torque of the proposed HPDPMSM is almost equal to that of the conventional one. The torque ripple of the traditional and proposed HPDPMSM is 1.03Nm and 0.83Nm respectively. The reduction of the torque ripple indicates that not only the constant-power range is extended but also the torque ripple can be suppressed by the bypass-rib [15]. The field distributions in the typical operations are shown in Fig.17 and 18. As can be seen, the field distributions in the stators of the two HPDPMSMs are almost same. The bypassribs are saturated in the flux-weakening operation which means the additional paths for the d-axis armature field are provided by the bypass-ribs in the rotor. Those additional paths make obvious influence on the flux-weakening ability and constant-power operation range of the HPDMSM. To obtain the torque and power at different speeds, the toolkits in the FEA software are utilized. The torque-speed curves of the two HPDMSMs are shown in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000750_j.mechmachtheory.2018.11.002-Figure22-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000750_j.mechmachtheory.2018.11.002-Figure22-1.png", + "caption": "Fig. 22. Surface deviations of the gear tooth surface when generated by circular cutters with profile geometry and kinematics related to a shaper with different number of teeth from the pinion.", + "texts": [ + " In this section, the influence of using an increased number of teeth for the reference shaper when obtaining the profile of the circular cutter will be investigated. Four new study cases will be investigated considering a shaper with 25, 26, 27 and 28 teeth, respectively. We recall that the pinion of the gear drive has 24 teeth. Other parameters common to all the study cases include a mean cutter radius R c = 600 mm, a cutter profile tilt angle \u03b4 = 35 \u25e6, and no surface modifications applied to the face gear tooth surfaces for these cases. The parameters of the mating pinion are those presented in Table 1 . Fig. 22 shows the surface deviations for the four study cases. The deviation distributions for those cases are similar to that of Cases 2 and 3 (see Figs. 8 (b) and 9 (a)), but the deviations at the four corners of the active surface increase with the increase of teeth number of the reference shaper. Therefore, the effective crowning along the longitudinal direction of face gear tooth surface will be increased when the difference of the tooth number of the shaper and the pinion is increased. Fig. 23 shows the contact patterns for the four study cases" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001874_s00170-019-04738-3-Figure10-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001874_s00170-019-04738-3-Figure10-1.png", + "caption": "Fig. 10 Fifth-tooth finite element model", + "texts": [ + " Through the analysis of TCA results, it is found that edge contact may occur in the original gears but not in the modified gear; the locations of contact paths are also the same. In summary, the tooth surface modification method can effectively avoid tooth edge contact without affecting the contact path location of high contact ratio spiral bevel gear. To verify the rationality of the tooth surface modification method, the finite element analysis method is carried out on the working surface of high contact ratio spiral bevel gear. As shown in Fig. 10, the fifth-tooth FEA model is built. According to the model, the FEA on the working surface of the original gear and the modified gear are carried out using Abaqus software with the load of 20,000 N\u00b7m. The FEA results on the same meshing position are shown in Fig. 11. From the comparison of Fig. 11, it has been found that the maximum contact stress on the tooth surface of the original gear is located at the gear tip and pinion root, and the maximum contact stress on gear tooth surface is 1456 MPa and 1510MPa on pinion tooth surface" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002152_13506501211010030-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002152_13506501211010030-Figure2-1.png", + "caption": "Figure 2. Computational model.", + "texts": [ + " The parameters of the spiral bevel gears in the intermediate gearbox are presented in Table 1. During the operation process, the spiral bevel gears rotate at high speed, splashing the oil in the gearbox housing. An air vent is used to balance the atmospheric pressure inside and outside the gearbox housing. The simulation model contains only the gearbox housing, a pair of spiral bevel gears and the air vent. Based on the above analysis, a simplified simulation model obtained by Boolean operation is shown in Figure 2. The computational model was divided by a tetrahedral unstructured mesh using ANSYS meshing software, as shown in Figure 3. To prevent negative volume mesh appearing near gears during reconstruction, the mesh is refined locally. To maximise the calculation accuracy and reduce the challenge in mesh generation, the simulation model is simplified by (1) removing the chamfer and transition fillet of the gear and (2) combining the small fragmented surface and narrow slit of the inner wall of the intermediate gearbox housing" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000769_978-3-662-43893-0-Figure3.5-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000769_978-3-662-43893-0-Figure3.5-1.png", + "caption": "Fig. 3.5 Virtual crown gear in a bevel gear generator. 1 basic crown gear, 2 bevel gear axis, 3 cutter head axis, 4 cradle axis \u00bc basic crown gear axis", + "texts": [ + " efny1,2 \u00bc \u03c0 2 myn 2 xsm1,2 mmn khf p mmn \u00fe RPy RPm tan \u03b8f1,2 tan \u03b1nD \u00fe tan \u03b1nC\u00f0 \u00de \u00f03:12\u00de On cylindrical gears, only a few variables (number of teeth, face width, pitch diameter, whole depth, profile shift, helix and pressure angles) are sufficient to determine exactly the relative motions between the tool and work piece. In generating methods, the relative motions are based on a basic rack with straight tooth profiles meshing with the work piece. If the rack is replaced by a hob, it is possible to determine the manufacturing motion which will replicate the motion of the rack. On bevel gears, the same principle applies. Instead of a basic rack, a virtual basic crown gear is used which, like the rack for cylindrical gears, has straight tooth profiles. Figure 3.5 shows the basic crown gear also termed generating gear in a bevel gear manufacturing machine. The generating motion takes place when the cutter, rotating about its own axis, is simultaneously rotating about the axis of the generating gear while the work piece rotates about its own axis at a given gear ratio with the generating gear. The generating gear can be flat, as the one displayed in Fig. 3.5, or take a conical shape. For the different gear manufacturing methods described in Sect. 2.1, there are various ways in which a cutter can be substituted to 3.2 Manufacturing Kinematics 65 the relevant virtual generating gear. Figure 3.6, for example, represents the Spirac\u00ae method in which a conical generating gear, used for a plunge cut wheel, is replaced by a tilted face cutter. Determining the respective relative motions between the cutter and the bevel gear is therefore by no means a trivial concern, particularly since additional motions are superimposed to modify and improve tooth flank topography", + " On a conventional mechanical cutting machine, the amount of tilt and direction of the tilt axis are set by means of the tilt and swivel angles (see Fig. 3.7). 66 3 Design The next three subsections provide a universal description of manufacturing kinematics for a virtual gear-cutting machine which can perform all important production processes without any geometrical restriction. A bevel gear machine is designed such that it imposes to the tool and bevel gear blank the relative motions of the generating gear meshing with the bevel gear. For this purpose one tooth of the generating gear is replaced by the straight flank profile of the tool. Figure 3.5 shows a tool used in the single indexing method. In order to roll the tool with the blank, i.e. to machine one tooth slot, and return to the starting position, the generating gear shown in the figure must rotate by a relatively small angle. On many mechanical machine tools, the cutter moves to and fro like a cradle, from the start of roll angle to the stop roll angle and back again. Thus, conceptually, a virtual gear-cutting machine closely resembles the former mechanical machine, but with the significant advantage that the virtual machine is not subject to any limitation in terms of penetration, stability, damping, assembly, accessibility etc" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001822_tie.2020.3034859-Figure11-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001822_tie.2020.3034859-Figure11-1.png", + "caption": "Fig. 11. Proposed voltage command compensation for ZOH reflected one step delay.", + "texts": [ + " Different from FOC with vector control, the torque vector direction is not always perpendicular to the flux vector. Authorized licensed use limited to: University of Gothenburg. Downloaded on December 19,2020 at 13:25:15 UTC from IEEE Xplore. Restrictions apply. 0278-0046 (c) 2020 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. Thus, more consideration on voltage compensation is desired due to six-step commutation. This insufficient sample can be explained by Fig. 11. In this case, there are only 18 sample points per cycle. Assuming at sample instant t=0T, the voltage output is 1 with per unit value and the voltage vector is 90deg leading to the rotor position. For the next step at t=1T, the BLDC actually executes this 1 per unit command. However, the rotor has rotated 20deg. The corresponding torque vector is now 20deg leading to the rotor position. This ZOH one step delay degrades the torque output to 1sin(90+20)=0.9397. It results in the stability issue for sixstep commutation" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000935_tia.2014.2311503-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000935_tia.2014.2311503-Figure2-1.png", + "caption": "Fig. 2. Different winding connections of the six-stator/four-rotor pole SR machines under ac sinusoidal bipolar with a dc bias excitation. (a) Symmetrical connection. (b) Asymmetric connection.", + "texts": [ + " The operation principles of such excitation can be simply summarized as follows: The dc bias flux, which is equivalent to the magnet flux in PM machines, interlinks with the ac coils of each phase and induces voltages, which equivalently will be called the back electromotive forces (EMFs). Consequently, similar to PM machines, output torque will be generated if the ac coils are excited by ac currents. Further illustrations and investigations will be presented in the following sections. Moreover, the coils under the ac sinusoidal bipolar with a dc bias excitation have the possibility to be connected either symmetrically, as in Fig. 2(a), or asymmetrically, as in Fig. 2(b). However, it should be mentioned that, in the case of an asymmetric winding connection, only four short-pitched dc coils are required, and there is no need for dc coils on the Phase B teeth, i.e., there are no dc coils on teeth 2 and 5 in Fig. 1. This is because such coils do not contribute to the output torque, but they only generate extra copper loss, i.e., different dc polarities in the same slot. It should be also mentioned that the asymmetric connection can be also equivalently obtained by two full-pitched dc coils (or actually one full-pitched dc coil), as illustrated in II in Fig. 2(b), which are actually woundfield doubly salient machines [27]. In this case, the effective length of the machine becomes larger due to the longer dc-coil end winding. However, the machine efficiency for both shortpitched and full-pitched dc-coil-excited machines depends on the total copper loss of all dc coils, including the end winding. In this study, only the short-pitched concentrated dc coils will be considered since both asymmetric winding connections, as shown in Fig. 2(b), are electromagnetically equivalent. It is also worth mentioning that, under the ac sinusoidal bipolar with a dc bias excitation, the ac frequency is determined by the rotor pole number, i.e., similar to the switched-flux PM (SFPM) machines [28]. In other words, one electric cycle completes when the rotor mechanically rotates one rotor pole pitch. As aforementioned, the operation principle of the ac sinusoidal bipolar with a dc bias excitation is similar to its counterpart in the PM machines", + " The open-circuit torque, i.e., the cogging torque, of the two connections are predicted at a dc current density of 30 A/mm2 and compared in Fig. 6(a). It shows that the cogging torque period of the symmetrical winding connection is 30 mechanical degrees since the least common multiple between the rotor poles and the equivalent stator slots is 12; thus, the cogging torque period is 360\u00f7 12 = 30. On the other hand, for the asymmetric connection, the dc stator polarity repeats once over 360 mechanical degrees, as shown in Fig. 2(b). This is equivalent to two stator slots from the cogging-torque-producing point of view. Thus, the cogging torque is produced by the interaction of the four rotor poles with these slots. This means that the cogging torque waveform is 90 mechanical degrees, as confirmed in Fig. 6(a). Furthermore, in order to examine the influence of magnetic saturation, the cogging torque is predicted for different current densities, and the peak values that are due to the two winding connections are compared in Fig", + " However, at a relatively low current density, as in Fig. 5(c), such distribution becomes less balanced; thus, the cogging torque becomes relatively large. Furthermore, the three-phase open-circuit flux linkages, which are seen by the ac coils, are calculated and compared for the two winding connections in Fig. 7. It shows that the flux linkage waveforms of the asymmetric winding connection are nonuniform, as shown in Fig. 7(a), since the winding distributions of such connection are also nonuniform, as shown in Fig. 2(b). On the other hand, the symmetrical winding connection results in uniform flux linkage waveforms, which are larger than their counterparts in the asymmetric connection since such connection has a larger air-gap flux density, as shown in Fig. 3. However, the variations of such waveforms, i.e., from maximum to minimum, are smaller since, for the symmetrical winding connection, the difference between the flux distributions at the maximum flux linkage position of Phase A, i.e., when the rotor poles are fully aligned with the Phase A stator poles, as shown in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001421_iros.2016.7759840-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001421_iros.2016.7759840-Figure3-1.png", + "caption": "Fig. 3: Map generated by continuous mapping.", + "texts": [ + " The new hull boundary is described by only one point per bin; it is the farthest point from the centroid, regardless of whether it comes from the new observation or from the previously simplified map. This process downsamples the boundary point set, which improves the performance of the path planner. It also helps us to handle the corner cases when planes appear and disappear while going in and out of the field of view. An example of resulting map with the continuous mapping process is presented in Fig. 3. The image processing pipeline employed here runs at 25Hz on a single core PC. To provide a sequence of realizable footsteps in the perceived terrain we use a sampling-based step planning method. The approach takes into account the environment, represented as support planes, and efficiently exploits precomputed knowledge about the capabilities of the robot. The step planner relies on a quick validation of a large number of feasible footstep locations. A feasible footstep requires a suitable support surface in the environment, which should be reachable given the kinematic constraints of the robot" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002061_tec.2020.3048442-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002061_tec.2020.3048442-Figure1-1.png", + "caption": "Fig. 1. Cross-sections of doubly salient machines (brown line coil indicates AC armature coil, and green line coil indicates DC field coil). (a) Conventional DSPM machine with stator yoke PM [8]. (b) Proposed DSHE machine.", + "texts": [ + " With advanced control techniques, various PM machine topologies have been presented and the brushless excitation has been the trend to eliminate the mechanical contact of brushes and slip rings [2] [3]. The stator PM machines with doubly salient structure are popular and the flexible location of PM results in diverse machine topologies [4]-[6]. The concept of stationary PM with rotary salient pole to alter flux linkage is firstly introduced in [7] for single phase alternator, and then extended to the three phase doubly salient PM (DSPM) machine in [8], as shown in Fig. 1 (a). The armature coils are non-overlapped wound around the stator teeth, with potential high winding packing factor and compact volume [9]. Since all the excitations are placed in the stationary component, there is no need of brushes/slip rings for field winding excitation, while the heat management is more effective. Moreover, the rotor is the same as that in switched S. Cai and Z. Q. Zhu are with the Electrical Machines and Drives Group, University of Sheffield, Sheffield, U.K. (email: scai14@sheffield", + " Nevertheless, the field winding as well as armature winding are overlapped for the DSHE machines presented in [18]-[27] with the field winding coil pitch of 3 slot pitches. Therefore, the hybridization is obtained as the sacrifice of less compact volume and reduced slot filling factor for the DSHE machine with overlapped field winding compared with original DSPM machine with non-overlapped winding. The purpose of this paper is to present a novel DSHE machine with non-overlapped field winding, which inherits the compact structure of conventional DSPM machine. To achieve hybrid excitation in the DSPM machine of Fig. 1 (a), field winding is introduced to replace part of armature winding in Fig. 1 (b). It is worth emphasizing that iron bridge is attached beside the PM to provide additional flux path for DC excitation, which will be investigated subsequently. By allocating the stator teeth evenly, a novel DSHE machine with non-overlapped field winding is presented. In this paper, the proposed DSHE machine topology as well as the operation principle will be firstly illustrated from the perspective of PM and WF flux paths. Then, the possible stator and rotor pole number combinations are presented and compared under the genetic algorithm embedded global optimization. Furthermore, the effect of iron bridge on the hybridization is investigated with frozen permeability method. Finally, the electromagnetic performance of the proposed DSHE machine is evaluated with finite element (FE) analysis and validated with experiments. The proposed DSHE machine with both non-overlapped armature and field windings is shown in Fig. 1 (b), with 12/8 stator-slots/rotor-poles (Ns/Nr) as an example. The circumferentially magnetized PMs are placed at the stator yoke, with adjacent pieces of opposite directions. Therefore, the flux concentrating effect of the doubly salient PM machine is inherited. The field coils are wound around the PMs sandwiched \u2018one-tooth\u2019, whereas the armature coils are wound around the PMs sandwiched \u2018three-teeth\u2019. The polarity of all the DC coils is identical, and the arrangement of armature coils will be discussed in section III", + " The effect of DC excitation on the phase flux linkages with the assistance of 2D FE is illustrated in Fig. 5. It can be observed that the DC excitation can regulate the phase flux linkages effectively. The fundamental flux linkage can be increased twice with positive DC excitation, and the PM flux can be totally countered by negative DC excitation. Moreover, the flux weakening capability is better than the flux enhancing due to magnetic saturation with positive DC excitation. As illustrated in Fig. 1 (b), the stator tooth of the proposed DSHE machine is composed by field coil wound tooth and armature coil wound tooth separately. The minimum stator element contains 1-field coil wound tooth and m-armature coil wound teeth for an m-phase machine. To achieve symmetrical phase flux linkage, the stator core should include at least m-unit machine. Therefore, the stator tooth and rotor pole number for an m-phase machine can be defined in (2) to obtain the starting torque for the multiphase machine. 1 2 1s r N m m k N k m (2) where k1 and k2 are any integers", + " Downloaded on June 14,2021 at 20:12:10 UTC from IEEE Xplore. Restrictions apply. 0885-8969 (c) 2020 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. > REPLACE THIS LINE WITH YOUR PAPER IDENTIFICATION NUMBER (DOUBLE-CLICK HERE TO EDIT) < 4 number of Ns=12, and other stator tooth number can be extended as well. It should be noted that the unit machine with Ns=12 has three PM poles in Fig. 1 (b), which potentially results in unbalanced magnetic force (UMF). The asymmetric structure and UMF can be eliminated by doubling the stator and rotor poles. The mechanical and electrical angles between two armature coils can be defined in (3) according to the previous analysis, in which the stator tooth number of n=4k (k=1, 2, 3) is wound with field coil. To illustrate the armature winding configuration, the coil EMF phasors and winding connections of 12/7 and 12/8-stator slots/rotor poles HE machines are shown in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000951_j.triboint.2016.06.032-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000951_j.triboint.2016.06.032-Figure1-1.png", + "caption": "Fig. 1. Structure and kinematic diagram of non-clearance ball precision transmission.", + "texts": [ + " Based on analyzing curves of normal meshing force of two models, it is proved that the sliding friction can not be neglected in the establishment of mechanical model and the model of pure rolling friction does not conform to the actual meshing. A coupled mechanical model of rolling and sliding friction is established and the nonlinear finite element method is used to simulate the contacts between balls and grooves. The simulation result shows good matching with the theoretical deduction. 2. Structure and transmission principle Fig. 1 shows the structure of real-time no-clearance steel ball precision transmission, which is consisting of the fixed central plate 1, the combined planetary plate 2, the reduction steel ball system 3, the equal speed steel ball system 4, the output plate (shaft) 5 and the eccentric input shaft 6. The central plate 1 is machined with epitrochoid enclosed groove which has Z1 teeth on the right end face. The combined planetary plate is machined with hypotrochoid enclosed groove which has Z2 teeth on the left end face, and \u2212 =Z Z 22 1 " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000304_j.mechmachtheory.2018.05.006-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000304_j.mechmachtheory.2018.05.006-Figure3-1.png", + "caption": "Fig. 3. Singular robot postures: (a) leg singularity on the first limb, (b) and (c) passive-constraint singularities for two different pointing directions.", + "texts": [ + " This can be readily proven for a robot sub-set of the 3-RPR class of manipulators. For simplicity, we define here the three vectors n i , each perpendicular to the i th leg plane, for i = 1 , 2 , 3 : n i = \u2212d i \u00d7 \u02c6 ai (1) Unactuated kinematic chain singularities take place when a subset of screws of a leg becomes linearly dependent, i.e., when a leg singularity occurs. In particular, for the case under study, two singular postures for each leg are identified that appear when the axes of two of the revolute joints of the same leg become coaxial: Fig. 3 (a) shows the posture at which the revolute pairs are coincident, while it can also occur when they lie on opposite sides of the sphere surface. Moreover, through an appropriate arrangement of the R and P joints composing the planar pair, as the one shown in Fig. 4 , such singular configurations lead also to the linear dependence of the screws of the planar pair sub-chain. This type of singularity occurs when the two vectors a i and d i are parallel, i.e., when n i = 0 . A Passive-Constraint Singularity (PCS) occurs when the dimension of the MP-motion space increases with respect to its global mobility value", + " Each leg develops on the MP a constraint force normal to the plane of the leg and passing through the origin so that the MP motion is a three-dimensional rotation around the origin. When these forces become linearly dependent, e.g., when they are parallel to the same plane, the dimension of the motion space increases from three to four and the mechanism is at a PCS. Unlike the previous case, there is at least a PCS posture for each pointing direction of the MP, i.e. given a pointing vector, it is always possible to find a singular posture by rotating the MP around this axis. Two examples are shown in Fig. 3 : Case (b) shows a PCS posture for the vertical direction while case (c) sketches a different pointing direction. In both cases, the constraints imposed by each leg on the MP, i.e. the three forces perpendicular to the leg planes (shown in the figures as vectors stemming from the center of the spherical motion), become linearly dependent: in fact, the three vectors lie on the same plane and the twist system of MP gains one translational degree of mobility perpendicular to the force plane. This occurs when: det [ n 1 n 2 n 3 ] = 0 (2) The relation imposed by (2) can also be expressed as n 1 \u00d7 n 2 \u00b7 n 3 = 0 . Also, it is worth remarking that if the leg singularity condition is fulfilled, i.e., n i = 0 , a PCS also occurs as can be seen in Fig. 3 (a), namely, n 1 \u00d7 n 2 \u00b7 n 3 = 0 . Considering the above condition, it turns out that the best solution is obtained by actuating the prismatic joints closest to the ground that compose the planar joint of each leg, once they have been brought back to the base platform. Doing that, the wrench acting on the platform is given by a force that lies in the leg plane, whose direction and application point depend of the planar joint sub-chain; within screw theory , force and moment are stored in a six-dimensional array, the wrench " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000302_b978-0-12-813691-1.00006-3-Figure6.20-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000302_b978-0-12-813691-1.00006-3-Figure6.20-1.png", + "caption": "FIGURE 6.20", + "texts": [ + " Despite many good advantages of metallic electrodes, they have some problems, such as alternation of the electrode surface by the adsorption of electrolyte species or the formation of oxide layers. To solve these problems, the electrodes need to be frequently reconditioned through physical, chemical, or even electrochemical polishing. A SPE is a chemically stable substrate on which one to three or even more electrodes (such as the WE, the pseudoreference electrode and the CE) are printed by screen-printing process (Fig. 6.20) (Couto et al., 2016). During the screen-printing process, layer-by-layer deposition of various inks at a solid substrate has been taken place. Screen-printing techniques provides a useful device in order to do in situ analysis of different compounds, especially in the case of pharmaceutical samples (Couto et al., 2016; Tudorache and Bala, 2007). They were reviewed in the reports from Bala (Tudorache and Bala, 2007), Banks (Metters et al., 2011) and Couto (Couto et al., 2016), considering their application on the A schematic representation of metallic disk electrode" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001532_j.wear.2019.01.053-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001532_j.wear.2019.01.053-Figure1-1.png", + "caption": "Fig. 1. Machine tool setting for pinion teeth finishing.", + "texts": [ + " In order to reduce the sensitivity of the gear pair to errors in teeth surfaces and to the mutual position of the mating members appropriately chosen modifications are introduced into the teeth of the pinion. As a result of these modifications theoretically point contact of the meshed teeth surfaces appears instead of linear contact. The modifications are introduced by the variation in machine tool settings and in the profile of the headcutter's cutting edges. The machine tool settings used for pinion tooth finishing are specified in Fig. 1: sliding base setting (c), basic radial (e), basic machine center to back increment (f), basic offset (g), tilt angle ( ), and swivel angle ( ). The other manufacture parameters are the velocity ratio in the kinematic scheme of the machine tool for the generation of the pinion tooth surface (ig1), the radius of the hadd-cutter (rt1), and the radii of the head-cutter profile (r r,prof prof1 2). The tooth surface of the pinion is defined by the following system of equations: = i c f g e r r r r M M M M M r ( ) ( ) ( , , ) ( ) ( , ) ( , , ) p g p p p T T t prof prof (1) 1 1 4 1 3 2 1 1 ( 1) 1 1 2 (24a) =v e 0T T 0 ( 1,1) 0 ( 1) (24b) where r T T( ) 1 1 is the radius vector of head-cutter surface points, matrices M1, Mp1, Mp2, Mp3, and Mp4 provide the coordinate transformations from system KT1 (rigidly connected to the cradle and head-cutter T1) to the stationary coordinate system K ", + " The position vector of the formed gear tooth surface points is obtained by a simple coordinate transformation of vector r T T( ) 2 2 from system KT2 (rigidly connected to head-cutter T2) into the to the stationary coordinate system K, as it follows =r M M r( ) g T T(2) 2 2 1 2 ( 2) (25) The details of the theory of manufacture and meshing of face-milled hypoid gears are presented in Refs. [84] and [86]. Eqs. (23\u201325) represent a system of equations with unknowns uP , P, P , tP , and P , where uP and P are the parameters of head-cutter surface for pinion tooth surface generation, P is the cradle angle (Fig. 1), tP and P are the parameters of the head-cutter surface for the gear teeth formation. The solution is obtained by iterations. The Reynolds and energy equations are coupled through changes in the lubricant rheological state (density and viscosity), dependent on the pressure and temperature. A full numerical solution using finite difference form of the Reynolds and energy equations is required to calculate the pressure and temperature distributions in the oil film. Lubricant film temperature alters three-dimensionally, thus a three-dimension grid mesh is applied in the fluid film" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002583_sapm194625196-Figure4-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002583_sapm194625196-Figure4-1.png", + "caption": "FIG. 4a. MODELS OF BENNETT LINKAGES", + "texts": [ + "2) l a-b l ==k tanP tan R = a + b (a == tan A, b == tan B). Several formulas (apparently different, yet correct) have been published for the value of k. These differences arise from the conventions that have been used for measuring twists and other angles between lines, whether internal or external angles are taken, and whether angles are taken as positive or negative. In this paper, all twists are considered positive and are measured in the same sense, and ranging from zero degrees to all values less than 180 degrees. Fig. 4a shows several paper models of Bennett linkages. Each link is a tetrahedron. A pair of opposite edges of the tetrahedron sene as hinges join ing the link to the adjacent link;;. Fig. 4b is a more practical form of a Bennett linkage. It illustrates how the continuous rotation of a shaft can be transmitted to another non-parallel shaft by means of a hinged connecting rod. , Fw. 4b. ApPLICATION OF B ENNETT LINKAGE 4. The four-cell division of the Bennett linkage (Conoidal case). Consider a Bennett linkage EFGH. Let four links SL, SM, SN and SO having a common hinge at S be joined to the sides of the Bennett linkage by hinges at the points L, M, Nand 0 which lie on the sides. Now let each of the four cells into which the Bennett linkage is divided be itself a Bennett linkage" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000302_b978-0-12-813691-1.00006-3-Figure6.17-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000302_b978-0-12-813691-1.00006-3-Figure6.17-1.png", + "caption": "FIGURE 6.17", + "texts": [ + " The most common electrodes which were used for pharmaceutical analysis are categorized and introduced in this section and the CNM-based modified electrodes in next one. While reviewing electroanalytical techniques, it was found that the earliest voltammetry (especially polarography) is coupled with mercury electrodes, due to their beneficial advantages such as high overpotential toward the proton reduction, wide cathodic potential range, amalgam formation, and renewable surface (Wang, 2006; Harvey, 2000; Heyrovsky\u0301 and Ku\u030ata, 1965). The mercury WEs often consist of a drop suspended from the end of a capillary tube (Fig. 6.17). In the hanging mercury drop electrode (HMDE), a desirably sized Hg drop is formed through pushing it through a narrow capillary tube by a micrometer screw, which makes the suspended drop (Harvey, 2000). In the dropping mercury electrode (DME), successive mercury drops form at the end of a capillary tube resulting from gravity; each drop provides a fresh electrode surface (Harvey, 2000). In the static mercury drop electrode (SMDE), a solenoid-driven mechanical plunger is used to control the flow of mercury and each drop provides a fresh electrode surface (Harvey, 2000)" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001846_tte.2020.3041194-Figure22-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001846_tte.2020.3041194-Figure22-1.png", + "caption": "Fig. 22. Experimental platform. (a) Motor. (b) Platform.", + "texts": [ + " Verification in a three-phase motor In order to show the effectiveness of the proposed hybrid rotor, two three-phase Model 2 and Model 4 with 48 slot/8poles CPM motors are established, and their torque performances are given in Fig. 21. It can be seen that the torque pulsation (e.g. 12th and 24th in three-phase motor) in the Model 2 also can be suppressed greatly by employing the rotor of Mode 4 even in a three-phase configuration. V. EXPERIMENTAL VERIFICATION In order to verify the electromagnetic performances, the proposed Model 4 is built. The photos of the prototype motor and test platform are presented in Fig. 22. Obviously, from Fig. 22(a), it can be observed that the hybrid rotor configuration and shifted spoke-type PM are adopted. The central butterfly cut out in the prototype is used to fix the shaft, and the dynamic balance has been tested to ensure the same polar moment of inertia. Compared with other PM motors, the difficulty of the installation of the magnet of CPM motor is increased and the hybrid configuration increases the manufacturing complexity. Hence, the cost of manufacture becomes a little expensive. The experimental platform of the PM motor drive is also shown in Fig. 22(b). The digital controller is dSPACE 1105, the intelligent power module is PM100CVA120, and the current sensor is a Tek A622 current clamp, in which 0.1 V indicates 1 A. Both back-EMFs and torque measurements are carried out using the same torque sensor which type is Interface T8-20-BA4 with a 1 kHz cutoff frequency, and the range and precision of this torque sensor are 20 Nm and 0.25%, respectively. Moreover, a magnetic brake is used as the load, and the servo motor is used to measure the back-EMFs" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000021_demped.2015.7303703-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000021_demped.2015.7303703-Figure3-1.png", + "caption": "Fig. 3. The layout of the induction m", + "texts": [ + " These faults lead to the d motor parameters; for instance line current, unbalanced air pulsating torque, decrease in efficiency (increase in losses) noise level [12]. The paper involves the dyn induction motor and illustrates flux, the current and the torqu parameters. Additionally, the noise values are presented b conditions are investigated fo eccentricity types. Therefor eccentricity are studied in term magnetic parameters by compar operational conditions. II. METHOD A 2.2 kW, 4 poles, 50 Hz, connected squirrel cage inductio Firstly, the motor geomet parameters of rotor and stato optimal mesh topology is geometry as shown in Fig. 3. motor and examples of the causes of t of the rotor gap of an induction motor process, design features and ns of the asymmetry of the rotor t being at the center of the stator and rotor face are not perfectly bearing wear or mechanical ifferences of characteristics of , non-stable air gap voltage and gap flux density, increase in average torque, reduction in , overheating and increase in amic model of the squirrel cage the deviation in the magnetic e values in terms of electrical deviation in the vibration and y experimental results" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002026_tec.2020.3035258-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002026_tec.2020.3035258-Figure3-1.png", + "caption": "Fig. 3. Mechanical position and connection of armature coils of 6-slot stator with different rotor pole numbers.", + "texts": [ + " For a three-phase HESFM with 6-stator-slot (m=3, k1=2), the potential Nr are 4, 5, 7, and 8, respectively, while for a three-phase HESFM with 12-stator-slot (m=3, k1=4), the preferred Nr are 10, 11, 13, and 14, respectively. Consequently, the three-phase HESFMs having Ns=6, Nr=4, 5, 7, and 8 as well as Ns=12, Nr=10, 11, 13, and 14 are utilized for analysis. The connection of each phase can be obtained by coil back-EMF phasors. The electrical degrees \u03b8e between coil back-EMF phasors can be calculated by \u03b8e=Nr\u00d7\u03b8m (5) where \u03b8m is the mechanical degrees between coil back-EMF phasors. Fig. 3(a) shows the mechanical position of armature coils of the 6-slot stator. For different rotor pole numbers, the electrical position of armature coils can be obtained by (5). In order to obtain symmetrical and higher three-phase backEMFs, the connections between armature coils for 6/4 and 6/5 HESFMs are shown in Figs. 3(b) and 3(c), respectively. It should be noted that the polarities of coils n and n' are opposite due to alternate magnetized PMs in tangential direction. The armature coil connections for 6/7 and 6/8 machines are similar to those of 6/4 and 6/5 machines, respectively, and the only difference is the interchange of phases B and C" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001319_1687814020940072-Figure6-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001319_1687814020940072-Figure6-1.png", + "caption": "Figure 6. The Stewart platform leg construction.", + "texts": [ + " In order to determine the influence of the base platform motion on the system dynamic, a geodetic reference coordinate frame W is established at the antenna orientation center. As shown in Figure 4, the x-axis and the z-axis of frame W coincide with the azimuth axis and the elevation axis of the antenna, respectively. As shown in Figure 5, the rotation matrix RW B from the frame B to W can be expressed as RW B = cfcu sf cfsu sfcu cf sfsu su 0 cu 2 4 3 5 \u00f02\u00de where u and f are the elevation and the azimuth angle of the primary reflector, respectively. Figure 6 represents vectors and coordinate frames used for the kinematic issue of the 6-UPU Stewart platform. For each kinematic chain, a closed vector-loop equation can be written as follows bB i + li1n B i + li2n B i =PB +RB Pp P i \u00f03\u00de where PB is the position coordinate of the point P measured in the frame B, pP i is the coordinate of the point pi measured in the frame P, bB i is the coordinate of the point bi measured in the frame B, li1 is the distance from the point bi to the lower plane of the cylinder, li2 is the distance from the point pi to the lower plane of the cylinder, nB i is the unit direction vector of leg i in the frame B, from bi to pi, and nB i can be obtained as nB i = PB +RB Pp P i bB i li \u00f04\u00de For the ith leg, the leg length li is derived as li = li1 + li2 \u00f05\u00de where li is the length of the leg bipi, and li is derived as li = PB +RB Pp P i bB i \u00f06\u00de As shown in Figure 6, an actuator consists of two bodies: the rotating cylinder and the moving piston. The rotating body, with mass m1i and a constant distance e1, is connected with the universal joint to the base at bi. The moving actuator body with mass m2i and a constant distance e2 is connected with another universal joint to the platform at pi. The length of the actuator is adjustable with a sliding joint between these two bodies. The center of gravity positions of cylinder rB 1i and piston rB 2i are calculated as follows rB 1i = bB i + e1n B i \u00f07\u00de rB 2i = bB i + li e2\u00f0 \u00denB i \u00f08\u00de As shown in Figure 7, each leg connects to the base platform by a universal joint, and the rotation matrix which transfers leg frame li to frame B can be obtained as RB li = cficui sfi cfisui sficui cfi sfisui sfi 0 cui 2 4 3 5 \u00f09\u00de and the unit direction vector of leg i in the frame B is nB i =RB li nli i = cfisui sfisui cui 2 4 3 5 \u00f010\u00de where nli i = \u00bd 0 0 1 T " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002416_tmag.2021.3082326-Figure5-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002416_tmag.2021.3082326-Figure5-1.png", + "caption": "Fig. 5. Coil distribution of phase A of a 16-pole/21-slot machine", + "texts": [ + ", 2/ 2/ \u0399 s s t t ssc dBLRN (47) Authorized licensed use limited to: University of Prince Edward Island. Downloaded on July 03,2021 at 13:15:24 UTC from IEEE Xplore. Restrictions apply. 0018-9464 (c) 2021 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. where Nc is the number of coil turns, \u03b8s is the span angle of coil, and \u03c9 is the rotor angular speed. Considering a 16-pole/21-slot surface-inset machine, the coil distribution of phase A is shown in Fig. 5. The no-load induced EMF in each coil is expressed as ,5,3,1 \u03997 ,5,3,1 \u03992 ,5,3,1 \u03991 \u03c0217 sin2 \u03c02 sin2 sin2 u s duscsA u s duscsA u duscsA N tupKBLNREMF N tupKBLNREMF tupKBLNREMF (48) where B\u03c1sI ' is the amplitude of slotted air-gap flux density, Ns is the stator slot number and Kdu=sin(up\u03c0)/Ns. Hence, the resultant no-load induced phase EMF is EMFA= EMFA1+ EMFA2 + \u2026+EMFA7 (49) The instantaneous electromagnetic torque expression for a three-phase machine can be written as Te= (EMFAiA+ EMFBiB+ EMFCiC)/\u03c9 (50) where iA, iB and iC are three-phase balanced armature currents" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002401_j.mechmachtheory.2021.104371-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002401_j.mechmachtheory.2021.104371-Figure3-1.png", + "caption": "Fig. 3. Meshing direction in the state of tooth back meshing.", + "texts": [ + " \u03b1is the dynamic meshing angle; and \u03b1\u2032 is the theoretical meshing angle, generally 20 \u00b0 The angle between the meshing line and the coordinate axis is used to determine the direction of the meshing force, as shown in Fig. 2 , and the angle between the meshing line and the axis is shown in Eq. (11) : \u03c8 = \u03b1 + \u03b3 (11) The above analysis of the tooth surface meshing state shows that the direction of the meshing line of the tooth back meshing state is different from that of the tooth surface meshing state; as shown in Fig. 3 , = 1 or = \u22121 is used to represent the gear\u2019s meshing state. Similarly, the acute angle between the meshing line and the coordinate axis in the meshing state of the tooth back is analysed as shown in Eq. (12) : \u03c8 = \u03b1 \u2212 \u03b3 (12) The only difference between the helical gear pair and the spur gear pair is that the helical gear system design contains the helix angle parameter \u03b2 . In this paper, the helical gear pair is divided into N equal-width slices in the tooth width direction [21] (as shown in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000093_j.engfailanal.2013.02.003-Figure14-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000093_j.engfailanal.2013.02.003-Figure14-1.png", + "caption": "Fig. 14. Mode shapes and corresponding stress distribution (2nd mode).", + "texts": [], + "surrounding_texts": [ + "An acceleration transducer was fixed near the fractured gear on the bearing housing of the lubricating pump during lifespan tests, and the test signal in time domain was recorded. A fast Fourier transformation was performed to the test data before emergent shutdown, and frequency spectrum was obtained as Fig. 16. From Fig. 16, we could see the frequency component was very complex, so we would give a detailed analysis in the following. Firstly, we calculated the main frequencies that might exist in the lubricating pump. From Table 1 we could get Table 4 by simply multiplication. Then we could divide the big peaks in Fig. 16 into three series named \u2018\u2018A,B,C\u2019\u2019. Here \u2018\u2018A1\u2019\u2019 to \u2018\u2018A6\u2019\u2019 were multiple frequencies of 1500 Hz, and they rooted in exciting force caused by column pistons. For the value of \u2018\u2018A6\u2019\u2019 was much bigger, it was abnormal. And it was because of the disalignment of 9 column pistons according to previous experience. The disalignment might root in the fractured gear on the shaft. \u2018\u2018B1\u2019\u2019 to \u2018\u2018B3\u2019\u2019 were multiple frequencies of 2666.67 Hz, and they rooted in engaging force caused by 16 tooth. For the value of \u2018\u2018B1\u2019\u2019 was much bigger than \u2018\u2018B2\u2019\u2019 and \u2018\u2018B3\u2019\u2019, it was as ordinary vibration characteristics in gear pair. But there were cracks or fractures in the driven gear at that moment as a matter of fact. Inconsistency occurred because the gear pair was too small in the lubricating pump, so it was really hard to test the abnormal characteristics of gears accurately and directly by the acceleration transducer fixed on housing. \u2018\u2018C1\u2019\u2019 to \u2018\u2018C5\u2019\u2019 were multiple frequencies of 1030 Hz, and they rooted in exciting force caused by roller bearings. For the values of \u2018\u2018C2\u2019\u2019 and \u2018\u2018C3\u2019\u2019 were a little bigger, it was abnormal. And it was because of the disalignment of the two shafts, which might also root in the fractured gear on the shaft. Besides, there were a lot of small peaks in Fig. 16. They were caused by the modulation effect of rotation frequency [6]. Therefore, the frequencies were dense and complex in frequency spectrum. From the above test signal analysis, we could come to the following deduction. It was hard to test the abnormal characteristics of gears accurately and directly. But it was still useful to enhance the monitoring, because when we detected the outstanding value of multiple frequencies of the other exciting forces, such as the multiple frequencies of 1500 Hz and of 1030 Hz, we could estimate there may be disalignment in the two shafts, which might root in the abnormal status of gears." + ] + }, + { + "image_filename": "designv10_14_0002061_tec.2020.3048442-Figure10-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002061_tec.2020.3048442-Figure10-1.png", + "caption": "Fig. 10. Open-circuit field distribution segregation of proposed DSHE machine with FPM. (a) PM field without iron bridge. (b) WF without iron bridge. (c) PM field with iron bridge. (d) WF with iron bridge.", + "texts": [], + "surrounding_texts": [ + "The mechanical and electrical angles between two armature coils can be defined in (3) according to the previous analysis, in which the stator tooth number of n=4k (k=1, 2, 3) is wound with field coil. To illustrate the armature winding configuration, the coil EMF phasors and winding connections of 12/7 and 12/8-stator slots/rotor poles HE machines are shown in Fig. 6. Similarly, the armature winding connections of other rotor pole number with Nr=4,5,7,8,10,11\u2026 can be obtained. It is worth emphasizing that the EMF phasors in Fig. 6 indicate the phasor angle of corresponding tooth wound armature coil, whereas the amplitudes of different coils can be different as discussed in Fig. 3. Obviously, the armature winding distribution factor is 1 when the rotor pole number Nr=4k, which can result in higher phase flux linkage as well as back-EMF. 1 1 360 1 , 4 360 1 , 4 n m s n e r s n n k N N n n k N (3)" + ] + }, + { + "image_filename": "designv10_14_0000031_1.4023324-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000031_1.4023324-Figure1-1.png", + "caption": "Fig. 1 Pitch cones of straight bevel gears", + "texts": [ + " 135 / 034504-1Copyright VC 2013 by ASME Downloaded From: http://mechanicaldesign.asmedigitalcollection.asme.org/ on 01/17/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use The basic design parameters of straight bevel gears are the normal module Mn; the number of teeth of pinion and gear zp and zg, respectively; the shaft angle R; and the pressure angle a. The gear ratio is given, in the same manner as that of other types of gears, as follows: i \u00bc zg zp (1) The pitch surfaces of bevel gears are cones as shown in Fig. 1. Therefore, the following equation yields dp0 \u00bc dg0 i (2) where dp0 and dg0 are the pitch circle diameters Pitch Circle Diameter (PCD) of the pinion and gear, respectively. The pitch cone angles of the gear and pinion are represented by dg0 \u00bc tan 1 sin R 1=i\u00fe cos R dp0 \u00bc R dg0 (3) The outer cone distance is determined using Eq. (3) as follows: Re \u00bc dg0 2 sin dg0 (4) Therefore, the mean and inner cone distances are as follows, respectively Rm \u00bc Re b 2 Ri \u00bc Re b (5) where b is the face width. The normal module is defined as that in the center of the tooth surface in this paper, although it is usually chosen as that in the heel side" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000967_03043797.2016.1225002-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000967_03043797.2016.1225002-Figure1-1.png", + "caption": "Figure 1. Rear subframe design for Subaru BRZ; (a) Mg subframe deflection analysis; (b) Cast Al subframe stress analysis; (c) Final subframe CAD model; (d) Designed subframe prototype.", + "texts": [ + " Moreover, during the semester, regular contact was kept among the teacher, students and industrial representatives to report progresses, exchange data, provide feedbacks and discuss concerns through emails or phone calls. Those contacts and communications helped to keep the projects on the right track. Other sponsors such as Airbus Helicopters, CAVS, Wise Seat and SAI heavily participated in their projects in similar manners. It is also worth mentioning that the sponsors\u2019 feedbacks were seriously considered when grading the project outcomes such as oral presentations and final reports. Figures 1 and 2 illustrate design examples acquired from our student teams. Figure 1 shows the design of a rear vehicle subframe. Two materials, magnesium and cast aluminium were tested for the subframe through FEA deflection and stress analyses, and a real subframe model was prototyped based on the recommended design for further experimental validation. Figure 2 shows the design of a tensile test grip, including experimental failure test of original grip and FEA failure analysis of the new design. As for the organisation of this class, two 75-minute sessions per week are devoted to lectures, computer labs and project discussions" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001924_j.jmps.2020.103959-Figure4-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001924_j.jmps.2020.103959-Figure4-1.png", + "caption": "Fig. 4. Experimental setup: the shell sample is mounted on the basal fixture and connected with two hollow rubber tubes, one to the syringe and the other to the pressure sensor. The syringe is pulled by the Bose ElectroForce 3510 tester to extract air at a constant flow rate of 54 ml/s. A microcontroller regulates the operation of the pressure sensor which measures the internal pressure of the shell.", + "texts": [ + " 3 b plots the scaled thickness profile as a function of the normalized arc length from the upper pole to the equator. The maximum thickness of the eight samples is t = 0 . 88 \u00b1 0 . 03 mm . The thickness profile is relatively uniform with t = 0 . 20 mm and minor fluctuations of 0.07 mm along the normalized arc length, except for peaks appearing at s/ s 0 = 0 . 36 and 0.68, which respectively correspond to maximum thickness values of t = 0 . 61 mm and t = 0 . 88 mm . To reduce the volume enclosed by the thin shell and monitor the pressure evolution acting on it, we assembled the experimental setup shown in Fig. 4 . Its main components include a polypropylene syringe to extract the air inside the shell at a controlled flow rate, a Bose ElectroForce 3510 tester (Bose Corporation, Framingham, Massachusetts) used to impart a displacement load on the piston of the syringe, an acrylic fixture to constrain any sample movement, along with a pressure sensor (SM9333, SMI, California) and a microcontroller (Arduino UNO, Arduino, Italy). The thin elastic shell was mounted on an acrylic fixture, which consists of two supporting plates" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000194_s12206-015-0833-3-Figure4-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000194_s12206-015-0833-3-Figure4-1.png", + "caption": "Fig. 4. Sketch of the sample bearing of type FL-HSB2421DFT.", + "texts": [ + " Based on the given maximum contact stress between the element and raceway, the loads applied on the bearing are further determined by the mechanical model mentioned in Fig. 2. At the same time, the contact forces, stresses and deformations between the element and raceway for each sample bearing under the applied load are calculated. Then, the fatigue lives of sample bearings can be further calculated and compared. The effects of size and raceway hardness on the fatigue life of large rolling bearing are studied. The four contact-point ball bearing of type FL-HSB2421 DFT is chosen as a sample bearing. A sketch of bearing FLHSB2421DFT is shown as Fig. 4 based on the standard JBT10705-2007 in China. There are nine sample bearings with different sizes. The structure parameters of every sample bearing are listed in Table 1. D is the bearing outer diameter; d is the bearing inner diameter. T is the bearing width; H is the width of ring with no tooth; and Dpw is the pitch diameter. Fig. 5 shows the variation of bearings\u2019 pitch diameters. The materials of inner and outer rings are 42CrMo4. The ball\u2019s material is GGr15SiMn. The depth of raceway hardening is 5 mm, which consists of surface, middle and base layers" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001626_icra.2019.8793617-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001626_icra.2019.8793617-Figure3-1.png", + "caption": "Fig. 3. Depictions of the mathematical spaces being reasoned about in this work. (a) Pointwise IK considers inversions at a single time. (b) PathwiseIK inverts from an end-effector path function over time (here, the temporal dimension is represented by the circles being extruded to cylinders).", + "texts": [], + "surrounding_texts": [ + "In this section, we review prerequisite terminology and notation used throughout the work and formalize our problem statement, leaving specific algorithmic details for \u00a7IV." + ] + }, + { + "image_filename": "designv10_14_0001725_s11666-020-01039-0-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001725_s11666-020-01039-0-Figure1-1.png", + "caption": "Fig. 1 (a) Experimental schematic diagram illustrating the laser remelting of Inconel 718 substrate under static magnetic field, (b) top view of the detailed magnetic field device, with V and B indicating the directions of laser beam and magnetic field, respectively, (c) the distribution of magnetic field intensity while the distance between the two magnet blocks is 40 mm", + "texts": [ + " The microstructural observation was also conducted to reveal the effects of the magnetic field on the solidification process. At last, the discussion was made on the mechanism of the magnetic field in influencing the laser remelting process. The surface remelting of the Inconel 718 plate was conducted using the AXL-AW700 laser system (Dongguan Aoxin Laser Technology. LTD, China) equipped with a pulsed Nd: YAG laser operating at a 1064 nm wavelength. The laser beam has a Gaussian profile with a 200 lm spot diameter. As illustrated in Fig. 1(a), two cuboid-shaped permanent magnets with a size of 80 mm 9 50 mm 9 40 mm were assembled inside the aluminum frame to create a transverse static magnetic field. The composition of Inconel 718 is shown in Table 2. The Inconel 718 substrate was placed at the center of the static magnetic field to maximize the magnetic intensity during the laser remelting process. As the top view is shown in Fig. 1(b), the magnetic field is perpendicular to the remelting direction of the laser beam on the Inconel 718 substrate. Besides, different magnetic field intensities were conducted to influence the laser remelting process by changing the distance between the two magnet blocks. As shown in Fig. 1(c), the distribution of magnetic field intensity is shown with the maximum value of 0.55 T at the center. Based on the experimental optimization, the processing parameters concerning the laser beam were chosen based on the minimal cracking and pore defects, and the laser remelting parameters are listed in Table 1. In order to account for the influence of substrate geometry, two kinds of specimens were used to measure longitudinal stress (type A) and lateral residual stress (type B). As shown in Fig", + " Similar to the residual deformation results, the x-ray diffraction results show that the compressive residual stress of the remelted Inconel 718 alloy is notably reduced under the magnetic field, and the decrease in stress magnitude is more evident at higher magnetic field intensity. The average values of residual stress for the samples A and B were decreased by 77.05 and 59.76 MPa, respectively, under the static magnetic field of 0.55 T. It is also interesting to notice that the decrease in residual stress is more evident at the center of the remelted region than the surrounding areas, which is corresponding to the magnetic field intensity distribution (see Fig. 1(c)). Similar to the trend of average residual stress at different magnetic fields (Fig. 1(a) and (b)), the distribution of residual stress indicates that higher magnetic field intensity has a more significant influence on the decrease in residual stress. The magnetic force and thermal electromagnetic effect imposed by the static magnetic field within the molten pool can significantly change the flow field as well as the heat transfer pattern, which can determine the solidification behavior even at such a higher cooling rate. To further understand the role of the magnetic field during laser remelting, the microstructural observation was also conducted on the molten pool of Inconel 718 substrate under different conditions" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002158_13506501211010556-Figure8-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002158_13506501211010556-Figure8-1.png", + "caption": "Figure 8. Oil volume fraction contours with different inner race rotating speeds (oil volume flow rate= 3.0 L/min).", + "texts": [ + " The influence of inner race rotating speed will be discussed carefully in the following section. The effect of inner race rotating speed The average oil volume fraction inside the bearing with different inner race rotating speeds is shown in Figure 7. It can be seen that the oil volume fraction becomes lower when the inner race rotating speed increases at different oil volume flow rates. Compared with the lower inner race rotating speed, the variation trend of the oil volume fraction with high rotating speed gradually becomes smaller. Figure 8 shows the simulated oil volume fraction distribution with different inner race rotating speeds. It seems that the outer race is almost covered with lubricant oil at a rotating speed of 5000 r/min. With the increase of rotating speed, the interaction between the bearing elements and the fluid increases, and the centrifugal force becomes larger. The flow velocity of the fluid inside the bearing increases with the higher rotating speed. It results in poor continuity of the oil film on the bearing element surface" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001925_978-3-030-46817-0_30-Figure14-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001925_978-3-030-46817-0_30-Figure14-1.png", + "caption": "Fig. 14. Classic and integrated solar panels.", + "texts": [ + " Renewable Energy Sources in Construction 715 This construction is suitable for roofs with two slopes, north-south orientation, where trapezoidal sheets with solar cells are placed on the south side. The connection of trapezoidal sheets to the roof structure is done with standard screws, shown in the Fig. 12. By installing thermal insulation we get a 3 in 1 system and get an energy efficient object [8] (Fig. 13). Fig. 12. Trapezoidal sheet fastening screw for wooden construction. 716 M. Torlo et al. In the following table we will present a comparison and analysis of classic and integrated roofing from a technical and economical point of view (Fig. 14). The cost of small photovoltaic systems is significantly higher than the cost of conventional technologies but has a downward trend. Also, photovoltaic systems occupy a significant space, so the trend of installing them on roof surfaces is increasingly interesting. The advantage of these systems is that electricity production is at the point Table 1. Comparison of classic and integrated panels Classic Integrated Highly developed and often applicable They are in the development phase Suitable for existing roof structures Suitable for new constructions Additional structural load Less structural load The roof consists of two covering bodies (tile and panels) 3 in 1 roof (panels are also cover, insulation and electricity generator) Simpler technical design More complex technical design They require less training for the assembly workers They require the training and precision of the workers when assembling The aesthetic appearance of the building or the roof is disturbed Aesthetically, it makes a slight difference between cover and panel Harder maintenance Easier maintenance Renewable Energy Sources in Construction 717 of consumption", + " The average velocity of the fluid in the standard channel is from 50 to 55 m/s (Fig. 10). The average velocity of the fluid in the reduced channel is from 26 to 31 m/s (Fig. 11). Fig. 9. Reduced Channel Pressure chart (Input Pressure 10 MPa) 726 O. Onysko et al. Figures 12, 13, 14 show the plot of the transverse velocity distribution in the reduced screw channel. The velocity distribution in the first thread (Fig. 12) is from 27 m/s to 30 m/s, the average section of the channel (Fig. 13) - from 23 m/s to 31 m/s, in the last turn (Fig. 14) - from 22 m/s to 31 m/s. The pattern of concentration of higher velocities in the center of the flow is closer to the end of the channel. Fig. 10. Average velocity of the fluid in the standard channel Fig. 11. Average velocity of the fluid in the reduced channel Computer Studies of the Tightness of the Drill String Connector 727 3. Basing on the points 1.2 we can conclude that the drilling fluid leakage is reduced in the case the partially permeability of tool-joint if the gaps between the roots of the pin and crest of the box of the thread and vice versa reduces" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001925_978-3-030-46817-0_30-Figure30-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001925_978-3-030-46817-0_30-Figure30-1.png", + "caption": "Fig. 30. Max warp angle validation of FEA model.", + "texts": [], + "surrounding_texts": [ + "Figs. 23, 24, 25 and 26. 142 I. Alagi\u0107 In case of ball joint, the results of stress distribution and allowed displacement are presented in Figs. 27, 28, 29, 30, 31 and 32. Finite Element Analysis (FEA) of Automotive Parts Design as Important Issue 143 The displacement achieved as a result of finite element analysis doesn\u2019t much differ from results of the laboratory-test performed by control device MR 96. The maximum displacement appeared into Z direction 0,145 [mm]. On the basis of conducted simulations were possible to affirm that the magnitude of deformations depends on model geometry. The largest concentration of stresses appeared in places near of cover of ball joint. Through change of model geometry it was possible to influence on expansion of stresses and displacement distribution. To analysis of maximum warp angle of ball joint using FEA solver was very low compared to its allowed value 58 \u00b1 6\u00b0." + ] + }, + { + "image_filename": "designv10_14_0001212_tpel.2019.2906431-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001212_tpel.2019.2906431-Figure1-1.png", + "caption": "Fig. 1. Dual 3-phase motor with two isolated neutral points. (a) Assembly drawing. (b) Configuration of the proposed system. (c) Magnetic flux distribution on the FEM model.", + "texts": [ + " Date of publication ***; date of current version March 23, 2019. Recommended for publication by Associate Editor ***. The authors are with the Dept. of Electrical and Computer Engineering and Institute of Electric Power Research, Seoul National University, Seoul, Korea (e-mail: sun.kwon@snu.ac.kr, jungikha@snu.ac.kr). Color versions of one or more of the figures in this paper are available online at http://ieeexplore.ieee.org. Digital Object Identifier *** be taken into account when designing large-capacity machines, as shown in Fig. 1(a). Thus, in a product such as a washer, instead of increasing z-axis depth, the technique of utilizing the created free space in the center of the stator has been applied [5]-[7]. One of the methods is t`o place an additional 3-phase motor on the inside of the outer 3-phase motor to realize high torque as well as to reduce cogging torque. This machine type belongs to a 6-phase motor and is commonly referred to as dual 3-phase motor (D3) connected in parallel, which has two sets of independent 3-phase windings that drive one shaft with isolated neutral points and no electrical phase shift as illustrated in Fig. 1(b) and Fig. 2(a) [8]-[13]. Furthermore, the two 3-phase stators are magnetically separated, and the magnetic flux generated by the currents applied to both motors is completely isolated without interference by the magnetic flux barrier, as shown in Fig 1(c). The dual 3-phase motor driven by a single inverter is mainly adopted in low-cost applications. Therefore, researchers and manufacturers had less interest in the studies of a fault-tolerant operation since it usually increases the cost. However, the open phase fault of the stator windings, which accounts for 21-37% of motor faults except for mechanical parts, is also a considerable problem in the dual 3-phase motor [14], [15]. Therefore, even in applications where motor failure does not cause serious safety problems, it is economically beneficial to have a fault-tolerant operation without additional cost, while the operating area is limited", + " Although initially designed for resolving inverter faults, the additional current path can also be used as a solution to motor faults [37], [38]. These studies proposed novel topologies and control methods, but are not cost-effective and cannot be applied to low-priced products. To achieve the higher power density and greatly enhanced reliability, this paper proposes a fault-tolerant operation of the dual 3-phase motor (FTO-D3), which drives a single shaft and has an isolated flux structure in the parallel-connected machine as shown in Fig. 1(a) and Fig. 2(a). The following should be taken into account when studying fault-tolerant and parallel operation with the single inverter: 1) the torque ripple and the sum of output torque; 2) the output torque of the healthy 3-phase motor; 3) the copper losses and efficiency; 4) the limitation for phase current and thermal rate; 5) no extra hardware with power semiconductors. This paper is structured as follows. First, Section II gives a control method of separating the d- and q-axis current in the stationary reference frame and operating each motor according to its torque" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002583_sapm194625196-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002583_sapm194625196-Figure1-1.png", + "caption": "FIG. 1. PLANE KEMPE LINKAGE (STRAIGHT CASE)", + "texts": [], + "surrounding_texts": [ + "By MICHAEL GOLDBERG* 1. Introduction. The following remarkable theorem is due to A. B. Kempe: Given any plane quadrilateral linkage EFGH, one may, without restricting the deformability of the quadrilateral, join a point S by links to properly selected points L, M, N, 0 on the sides. Extensions and further studies on this theorem were made by Kempe [1] and Darboux [2]. These linkages are illustrated in Figs. 1 and 2. This paper is concerned with an analogous theorem in three dimensions. We propose to investigate under what conditions it is possible for a skew quadrilat eral EFGH composed of four rigid bodies, each of which is hinged to its two adjacent members, to be deformable and to have the further property that four links with a common hinge at S have their other hinges on the sides of the quadri lateral without restricting the deformability of the quadrilateral. The linkage mechanisms presented here are a natural extension of those pre viously described by the author [3]. However, they are not the same as the three-dimensional generalization of Kempe linkages described by Fontene [4]. The linkages of Fontene employ only ball-and-socket joints and have two de grees of freedom, while the linkages here described employ only hinged joints and have only one degree of freedom. 2. Four-bar linkage mechanisms in three dimensions. It was first shown by Delassus [5], and more recently in a: neater demonstration by Dimentberg and Shor [8], that the only movable hinged four-bar mechanisms in three dimen sions are the following:- (a) the prism linkage in which all the hinges are parallel. This corresponds to the plane linkage. (b) the pyramidal or spherical linkage in which all the hinges pass through a point. This is the obvious generalization of the plane linkage; a spherical quadrilateral is used instead of a plane quadrilateral. (c) the skew isogram mechanism due to G. T. Bennett. In this mechanism the opposite links are equal in length, the angle (or twist) between the non intersecting hinges in a link are equal to the corresponding angle of the opposite link, and the lengths of the links are proportional to the sines of their twists. . 3. The Bennett linkage. Consider a four-bar linkage EFGH in which the opposite links are equal in length and all the joints are ball-and-socket joints. This linkage can be made to assume the form of a skew quadrilateral in space. Two typical cases are shown in Figs. 3a and 3b. Let the length of one pair of opposite links be m, and the length of the other opposite pair of links be n. * Bureau of Ordnance, Navy Department. 96" + ] + }, + { + "image_filename": "designv10_14_0001479_tmag.2019.2918190-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001479_tmag.2019.2918190-Figure3-1.png", + "caption": "Fig. 3. Stator core constructions. (a) Six basic modular segments. (b) Proposed modular machine.", + "texts": [ + " However, this side effect will be increasingly weak if the basic modular segment has large slot and pole numbers with the two-pole redundant machine unit. For example, a modular machine with this proposed structure contains six 24S/20P basic modular segments and a 6S/2P redundant machine unit. Under such condition, only 1/61 of the proposed modular machine is wasted, namely, practically negligible. In order to obtain the stator core of the proposed modular machine, the stator cores of the six basic modular segments are equally distributed in the circumference with space left, as shown in Fig. 3(a). Each segment possesses six teeth. Then, the stator core of redundant machine unit with six teeth is cut into six segments in the slot center position, while no coils will be wound on these six teeth for the redundant purpose (named as \u201cRT\u201d for simplicity). Combining the stator core segments of the basic modular segment and those of the redundant machine alternately in the stator yoke, the stator core of the proposed modular machine is obtained, as shown in Fig. 3(b). Similar to the rotor pole number, the stator slot number/tooth number of the proposed modular machine can be obtained as follows: Nsp = 3k Nss + Nsr, k = 1, 2, 3 . . . (2) where Nsp, Nss, and Nsr are the stator slot numbers of the proposed modular machine, the total basic modular segments, and the redundant machine, respectively. Here, they are 42, 6, and 6 when k = 2, respectively. As stated above, the rotor pole number of the proposed modular machine has increased due to the adoption of redundant machine unit", + " (3) where \u03c4sp is the tooth pitch of the proposed modular machine between adjacent effective teeth within one segment and \u03c4ss is the tooth pitch of the basic modular segment. The illustration of two parameters is shown in Figs. 1 and 3, respectively. The width of RT without coils for connecting stator segments is not restricted within its available range (Npr/Nsr pole pitch), since they will not carry coils for electromechanical energy conversion. Thus, it is optimized to minimize the torque ripple underrated operation condition. This value almost equals the half width of the effective teeth here. From the stator core structures shown in Fig. 3, it can also be seen that the repetition number of the proposed modular machine equals the RT number. This is due to the unequal tooth pitches and tooth widths for effective and RT in the proposed modular machine. The influence of repetition number on machine performance will be shown and analyzed later. The winding layout is the last but not least step for constructing the proposed modular machine. The star of slots technique is again used to explain the winding layout of two machines [5]. For the proposed modular machine, there are 36 teeth wound with coils and each of them is numbered to represent a coil, as shown in Fig. 3. The phase shift between two adjacent effective teeth (\u03b1pe) within one segment equals that of the basic modular segment (\u03b1i ) which can be obtained according to the following equation: \u03b1pe = \u03b1i = Nps 180 Nss . (4) The star of slots of the proposed modular machine is shown in the left column of Fig. 4(a). It can be seen that there are 12 uniformly distributed vectors along the circumference, while some vectors only belong to two coils and others are possessed by four coils. This is due to the additional phase shift caused by RT between two segments" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000067_s00170-015-7171-6-Figure8-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000067_s00170-015-7171-6-Figure8-1.png", + "caption": "Fig. 8 Illustration of multi-axis hypoid gear face-milling machine", + "texts": [ + " To calculate the machine settings, a generic postprocessing approach should be established. This work proposes a generic post-processing approach: first, the kinematics chain of this type of CNC face-milling machine tool is formulated, and then the coordinates of the machine axes are calculated to ensure the cutting system locations and orientations with respect to the gear. In the gear industry, several multi-axis CNC face-milling machines are available, and for generosity, the type of Gleason machine is adopted (see Fig. 8). The basic structure of the machine is (1) the machine table moves along the z axis; (2) the carriage assembly moves along the x axis; (3) the workpiece spindle moves along the y axis; (4) the cutter system is set up on the cradle, which rotates along its axis by angle C; and (5) the sliding base rotates about its axis by angle B. In face milling gears, the cutting system rotates in high speed, and the gear blank is indexed with an intermittent rotation after a tooth slot is completely cut. One of the machine features is the cradle with the cutter system, which is set up off the centre" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001964_j.oceaneng.2020.107822-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001964_j.oceaneng.2020.107822-Figure1-1.png", + "caption": "Fig. 1. Videoray Pro 3.", + "texts": [ + " In this paper, an improved SMC algorithm and a customized secondorder FLC have been proposed, and it also reports the comparison results between controllers of different types, through both simulations and experimental testing. Especially, in practical test, the results of these controllers are presented after the dynamic model is changed. In addition, both the simulation and the experimental work have evaluated the controller performances under quantified external disturbances. The VideoRay Pro 3 has been used for both the simulations and practical tests. It is a Remotely Operated Vehicle (ROV) with a camera, and it is powered through a cable, as shown in Fig. 1. It is propelled by two horizontal thrusters and one vertical thruster. The properties and coefficients for VideoRay Pro 3 were obtained from (Wang and Clark, 2007), and are presented in Appendix A. In order to find out the robustness of controllers to variance of the hydrodynamic and dynamic models, the vehicle has been modified with a tube installed on the bottom, as shown at Fig. 2. In this way, the dynamic and hydrodynamic model of the vehicle have been changed, and the controllers have been tested with the same coefficients on the modified ROV" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001847_tro.2020.3038687-Figure10-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001847_tro.2020.3038687-Figure10-1.png", + "caption": "Fig. 10. Design and implementation of 2M2D elbow joint.", + "texts": [ + " 2M2D Wrist Joint As the wrist is the slimmest part and is at the distal end of the arm, the externally actuated type of 2M2D joint was adopted. The design and implementation is shown in Fig. 9. Two motors (B3M-SC-1040-A, Kondo Kagaku, Japan) are located outside the two joints, and accommodated in the forearm in the longitudinal direction. The implemented wrist joint with the forearm link was lighter and slimmer than the forearm and wrist of an average adult. C. 2M2D Elbow Joint The implemented 2M2D elbow joint is shown in Fig. 10. As the elbow joint is between the shoulder joint and the wrist joint, it should have higher torque than the wrist, and save space for the wrist and shoulder joints. The elbow joint was designed with larger motors (B3M-SC-1170-A, Kondo Kagaku, Japan) to provide sufficient torque. The hybrid-actuated type 2M2D joint module was adopted to achieve appropriate size and motor location by coaxially arranging motor 1 and joint 1. The width of the structure was 85 mm, similar in size to that of an adult", + " The results verified that the load was shared by multiple motors in the way determined by the motor-joint routing matrix. The tendon coupling reallocated the outputs of the motors, and thus increased the maximum output torque of a single joint. Section II-B proposes the motor-joint coaxial coupling structure as a new motor location for the coupled tendon-driven mechanism. To validate the characteristics of the motor-joint coaxial coupling structure, we developed an externally actuated elbow joint (see Fig. 16) and compared it with the hybrid-actuated elbow joint. Different from Fig. 10, motor 2 was located outside joint 2. The externally-actuated elbow joint was 79 g heavier because of the additional pulley and bearing and the longer aluminum parts to hold motor 2. Tendons from motor 2 were also longer, 391 and 377 mm for the externally actuated type compared with 187 and 173 mm for the hybrid-actuated type, respectively. The motor-joint coaxial couple structure led to a compact design with shorter tendons and reduced self-weight. However, the center of gravity was shifted toward the distal end, which increased the load on the proximal shoulder joints" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001739_tec.2020.2995880-Figure19-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001739_tec.2020.2995880-Figure19-1.png", + "caption": "Fig. 19 3-D sketch maps of the AFPMM with rotor eccentricity", + "texts": [ + " o i \u03c0 2 \u03c01 6 \u03c0 2 \u03c0A t yg 0 6 1 d d ip R P i R i P N r B r a \u2212 + \u2212 + = = (25) where, a is the number of parallel branches, Nt is the coil number of one stator tooth, Npan is the number of stators and Byg is the axial component of the no-load air gap flux density at y=hm+g. At the rated speed of 4000rpm, the comparison of the back EMF under healthy condition and eccentric condition is illustrated in Fig. 18, in which (a), (b) and (c) correspond to the state of rotor radial deviation(e=0 mm and dr=4 mm), mixed eccentricity (e=1 mm and dr=4 mm) and angular eccentricity (e=1 mm and dr=0 mm) of the AFPMM, respectively. Fig. 19 shows the 3-D sketch maps of the AFPMM with rotor radial deviation and angular eccentricity, and the 3-D sketch map of the AFPMM with mixed eccentricity is shown in Fig. 4(b). The results obtained by the analytical model have a good agreement with the results from the FEA according to Fig. 18. Compared with the healthy condition, the back EMF peak value of the AFPMM with the rotor radial deviation and mixed eccentricity is decreased, while the back EMF peak value of the AFPMM with the angular eccentricity almost remains unchanged" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002097_s10846-020-01279-w-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002097_s10846-020-01279-w-Figure1-1.png", + "caption": "Fig. 1 In the XYZ-dimensional, one of the spheres on the link i and a dynamic obstacle are bounded by spheres whose transforms are T0 Sm and T0 A and radii are rm and rA, respectively", + "texts": [ + " In a micromanipulator dynamic model, all spheres are rigidly attached to the manipulator after motion. Let Ti Sm represents the position and orientation of \u03beSm with respect to \u03be i , the transform T0 Sm is given as follows: T0 Sm = T0 i T i Sm = T0 1 \u00b7 \u00b7 \u00b7Ti\u22121 i Ti Sm . (9) During the iteration process (Algorithms 1 and 2 in Section 4), the matrix T0 Sm transforms into a new matrix T0 Sm\u2032 , which represents the position and orientation of frame \u03be0 with respect to frame \u03beSm\u2032 . The transform T Sm\u2032 Sm is given as follows: T Sm\u2032 Sm = T Sm\u2032 0 T0 Sm = ( T0 Sm\u2032 )\u22121 T0 Sm . (10) As shown in Fig. 1, the distance dAm specifies the Euclidean distance between the obstacle frame \u03beA and the sphere frame \u03beSm . If dAm \u2264 rm + rA, it can be deduced that sphere Sm and sphere A are colliding. The Minkowski sum of these spheres Sm and A can be described by the equation Sm \u2295 A = {sm + a|sm \u2208 CH(Sm), a \u2208 CH(A)}, (11) where CH(Sm) and CH(A) are the convex hulls of sphere Sm and sphere A, respectively. The VO for sphere Sm induced by sphere A for time window \u03c4 is shown in Fig. 2. It can be given as V O\u03c4 Sm|A = {v|\u03bb\u03c4 (pm, v \u2212 vA) \u2229 A \u2295 \u2212Sm = \u2205}, (12) where \u2212Sm = {\u2212sm|sm \u2208 CH(Sm)}, vA is the velocity vector of sphere A, and \u03bb\u03c4 (pm, v \u2212 vA) is a ray starting at the position of T0 Sm with direction v, \u03bb\u03c4 (pm, v \u2212 vA) = {pm + t (v \u2212 vA)|t \u2208 [0, \u03c4 ]}" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002066_012153-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002066_012153-Figure1-1.png", + "caption": "Figure 1. Diagram of the forces acting on the tractor unit when it deviates from the straight-line trajectory", + "texts": [ + " This is especially important for tractors of cotton modification, where the size of the code part is limited by agrotechnical requirements and it is impossible to ensure traction-coupling properties by increasing the tire. We determine the vertical loads acting on the tractor unit in the transport position of the working tools. The chassis absorbs the weight of the tractor and agricultural machine; in motion, the loads are redistributed along the guide and driving wheels. Consider the sum of the moments of all forces acting on the tractor relative to the point OK (Figure 1), we get: (7) Let's compose the equation of moments with respect to the points \u041e\u043d and define Z\u043a ICECAE 2020 IOP Conf. Series: Earth and Environmental Science 614 (2020) 012153 IOP Publishing doi:10.1088/1755-1315/614/1/012153 (8) From equations (7), (8) it can be seen that the rolling resistance force reduces the vertical loads on the steering wheel and the drive wheels are loaded by the same amount due to the redistribution of loads along the axes. In such cases, the most significant feature of the operation of agricultural tractor units is the random nature of external influences, which determines its output indicators" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001758_s13369-020-04689-y-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001758_s13369-020-04689-y-Figure1-1.png", + "caption": "Fig. 1 3D geometric model of SPMSM", + "texts": [ + " Although mechanical magnitudes (bearing friction) were not taken into account after electromechanical conversion, the parameters such as air gap magnetic flux calculation or selection of magnet angle parameter were mainly focused. The order of design equations is obtained in the form of geometric equations, magnetic equations on non-salient structure and d\u2013q electrical equivalent circuit equations [4, 26]. The use of magnetic and electrical equations associated with geometric modeling is common in the design of electric 1 3 motors. The design studies based on the geometric model are simple but provide rapid analysis. The 3D model of the SPMSM has 12 slots and 10 poles shown in Fig.\u00a01. The objective functions are obtained by several assumptions. The facilitating processes are the result of the most appropriate design that underlies these assumptions. The permanent magnet synchronous motor is magnetic and fractional/concentric winding. Slot and pole numbers are chosen as 12 and 10, respectively, and the double layer winding are used. This geometric model can be stated with fewer equations than other buried magnets. There are three important elements: radial diameters of the machine, magnet dimensions and stator slot sizes" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001597_s11012-019-00996-3-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001597_s11012-019-00996-3-Figure3-1.png", + "caption": "Fig. 3 Locations of the ball centers and groove curvature centers of the bearing pairs with a running distance", + "texts": [ + " Accordingly, the new component force FL is derived as: FL \u00bc F0 P n KH \u00f0dL\u00de 3 2 sin aLw \u00f0FL\\F0 P\u00de F0 p \u00f0FL F0 P\u00de ( \u00f021\u00de Similarly, the new normal force QR and deformation dR of each right ball considering the effects of the sliding wear can be obtained as: QR \u00bc F0 P \u00fe FR n sin aRw dR \u00bc 1 KH 2 3 Q 2 3 R 8>< >: \u00f022\u00de where aRw is the new contact angle of each right ball as a result of the horizontal loadFa. Accordingly, the new component force FR is derived as: FR \u00bc KH \u00f0dR\u00de 3 2 n sin aRw F0 P \u00f023\u00de Figure 3 shows locations of the ball centers and groove curvature centers of the bearing pairs after the bearings had their respective normal wear depths. In Fig. 3, hL and hR are the normal wear depths of the left and right bearings, respectively,aPLw and aPRw are the new contact angles of the left and right bearings, respectively, as a result of the residual preload F0 P. For the bearing pairs subjected to the residual preload F0 P, the new distances sPLw and sPRw between curvature centers of the inner and outer races can be determined by: sPLw \u00bc sP hL \u00f024-a\u00de sPRw \u00bc sP hR \u00f024-b\u00de For the bearing pairs subjected to the residual preload F0 P and horizontal load Fa, the new distances sLw and sRw between curvature centers of the inner and outer races can be obtained as: sLw \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi sPLw\u00f0 \u00de2 sO cos aO\u00f0 \u00de2 q x 2 \u00fe\u00f0sO cos aO\u00de2 s \u00f025-a\u00de sRw \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi sPRw\u00f0 \u00de2 sO cos aO\u00f0 \u00de2 q \u00fe x 2 \u00fe\u00f0sO cos aO\u00de2 s \u00f025-b\u00de Then, the new contact angles aLw and aRw can be determined by: sin aLw \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi sPLw\u00f0 \u00de2 sO cos aO\u00f0 \u00de2 q x sLw \u00f026-a\u00de sin aRw \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi sPRw\u00f0 \u00de2 sO cos aO\u00f0 \u00de2 q \u00fe x sRw \u00f026-b\u00de According to the distances sO; sLw and sRw, the normal deformations dL and dR can be rewritten as: dL \u00bc sLw sO \u00f027-a\u00de dR \u00bc sRw sO \u00f027-b\u00de Hence, the new force equilibrium equation can be expressed as: where x0PL is the critical displacement of each left ball as a result of the residual preload F0 P, and can be obtained as Eq" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000525_s12541-016-0036-6-Figure14-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000525_s12541-016-0036-6-Figure14-1.png", + "caption": "Fig. 14 Revised interpolation algorithm of tooth surface", + "texts": [ + " According to the non-geometric-feature segmentation algorithm of discrete data points, grids which are not satisfied with Eq. (24) are extracted from tooth surface as object to interpolation. The control vertex Vi of interpolation curve P(ui) is obtained by interpolation algorithm. The corresponding weighted factor is \u03c9i. \u0394Ei is Ji SiRi 2 = J SiRi 2 i=1 n \u2211= T max \u03b4 1 \u03b4 2 \u2013 \u03b4 2 \u03b4 3 \u2013 \u03b4 1 \u03b4 3 \u2013, ,( ) min \u03b4 1 \u03b4 2 \u2013 \u03b4 2 \u03b4 3 \u2013 \u03b4 1 \u03b4 3 \u2013, ,( )\u2013= the deviation between the control vertex Vi and interpolation curve P(ui), as shown in Fig. 14. (25) Where, Vi is a control vertex. \u0394Vi is the deviation of curve control vertex Vi. \u03c9i is the weighted factor for Vi. \u0394\u03c9i is the weighted factor deviation of \u03c9i. Bi,k is the k order B-spline basis function in u direction. For the compensation of \u0394Ei on NURBS surface, it is necessary to adjust the positions of control vertices. Based on thrice interpolation algorithm of NURBS surface for position compensation of control vertices, the deviations of control vertices can be solved by (\u0394\u03c9i=0) (26) Where, " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000769_978-3-662-43893-0-Figure4.36-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000769_978-3-662-43893-0-Figure4.36-1.png", + "caption": "Fig. 4.36 Determination of boundary deflection using the mirroring method (valid, by analogy, for tooth root stresses)", + "texts": [ + " Different meshing positions, with different bending lever arms, are accordingly accounted for through a corresponding compliance function q(y\u00bc yF). On skewed and spiral bevel gears, the contact lines run diagonally across the tooth face, such that contacting tooth segments show different compliances caused by different bending lever arms and tooth thicknesses. The influence function, which is therefore also determined diagonally across the tooth, is established for each of the tooth segments not lying below the load application point using their compliance (q (y)) in response to loading at point yF (see also Fig. 4.36). 4.4 Stress Analysis 175 The influence of the finite face width b and the tooth boundary on the fading function E1 of the infinitely long tooth is accounted for by mirroring the deflection curve on the imaginary tooth face of the infinite tooth strip. Deflections lying outside the face width b are superimposed to the deflections within it (Fig. 4.35) [JARA50]. E \u00bc E1 \u00fe \u0394E1 \u00f04:315\u00de The result can be further improved by introducing a corrective function Wf . E \u00bc E1 x \u00bc x V x R \u00fe E1 x \u00bc x V \u00fe x R Wf x V ; x R \u00f04:316\u00de where x V \u00bc xVj j mmn ; x Vmax \u00bc 6; x R \u00bc xRj j mmn ; xV , xR according to Fig:4:35 \u00f04:317\u00de The influence of face canting on skewed and spiral gear teeth is allowed for by means of approximation Wf", + " q \u00bc qB \u00fe qS \u00fe qR \u00f04:325\u00de Bending compliance qB is determined on the basis of the equation for the deflection curve \u03b7 of a trapezoidal-shaped beam of variable cross-section (cf. [WEBE53]). In the case of hypoid gears, the tooth profile becomes asymmetrical and an imaginary mid-line of the tooth is tilted by the limit pressure angle \u03b1lim (see Sect. 2.2.5.3). In calculating the torque at a section given by distance y, it is necessary to add the bending component arising from the non-centric compressive force component of the normal force on the tooth (Fig. 4.36). 178 4 Load Capacity and Efficiency Determining Hertzian deformation With bending deflection, shear deflection and deflection of the part of the gear body adjacent to the tooth, a strong mutual influence exists between the individual tooth segments. In the case of Hertzian contact, it is assumed that only the particular segment of the tooth subjected to the load, will experience a deformation. Consequently, there is no need to allow for mutual influences and in terms of Hertzian contact the tooth is composed of decoupled segments of finite width", + " Form of the tooth faces W\u03c3 \u03b2St\u00f0 \u00de \u00bc 1 0:72 \u03b2Stj j \u00fe for the acute angled side of the tooth face for the obtuse angled side of the tooth face \u03b2 in radians \u00f04:332\u00de Boundary stress reduction R \u03be \u00f0 \u00de \u00bc tanh \u03be R \u00fe 1 \u00f04:333\u00de Influence function S\u00bc S1 \u03be \u00bc \u03be V \u03be R \u00fe S1 \u03be \u00bc \u03be V \u00fe \u03be R W\u03c3 R \u00f04:334\u00de where: \u03beV \u00bc \u03beV=mmn and \u03beR \u00bc \u03beR=mmn (see Fig. 4.35) Influence numbers sij may be determined using the stress reference value N and function value S: Influence numbers sij \u00bc S Nij sij describes the influence of the force Fnj exerted at point i on the stress at point i \u00f04:335\u00de Stress reference value Nij \u00bc 6yFj cos\u03b10 j s2Fni \u0394b Nij stress reference value for a tooth segment of width\u0394bat point i as a result of loading at point j \u00f04:336\u00de The stress reference value N includes the working point, lever arm h, force application angle \u03b1\u2019 (see Fig. 4.36) of the single force applied at point j, and the tooth thickness sFn in the normal section of the tooth root at each considered point i. This value is calculated for a tooth segment modeled as a cantilevered beam. 184 4 Load Capacity and Efficiency Once the nominal stress curve has been calculated in three dimensions, the local stress (notch stress) on a plane model is determined in a second step. The stress concentration due to the notch effect in the tooth root of a gear is dependent on the fillet notch parameter 2\u03c1Fn/sFn and on the lever arm of the load yF/sFn, and is thus different for each tooth segment of the numerical model" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001304_j.mechmachtheory.2019.103697-Figure23-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001304_j.mechmachtheory.2019.103697-Figure23-1.png", + "caption": "Fig. 23. The fabricated prototype of the six-axis force/torque sensor.", + "texts": [ + " 03 2 . 6 \u22127 . 99 1 . 92 \u22124 . 37 5 . 65 \u22122 . 11 3 . 82 8 . 12 1 . 92 \u22122 . 27 \u22126 . 01 \u22124 . 67 \u23a4 \u23a5 \u23a5 \u23a5 \u23a5 \u23a6 (41) The Norm corresponding to each column of matrix D fs is [ 9 . 91 11 . 51 4 . 55 8 . 57 9 . 03 6 . 57 ] and the Standard Deviation corresponding to each column of matrix D fs is [ 4 . 42 5 . 15 2 . 03 3 . 83 4 . 04 2 . 39 ] . The errors of the Norm and the Standard Deviation corresponding to each column of matrices D fs are within a reasonable range through comparing the calculation results. Fig. 23 is the fabricated prototype of the six-axis force/torque sensor and Fig. 24 is the six-axis force/torque sensor calibration equipment. The paper adopts the least squares fitting principle to calibrate the sensor. In order to provide large loading force and torque, a deadweight force standard machine which can provide precise loading force and loading direc- tion is used. Due to the deadweight force standard machine can only provide tension and pressure, the sensor can be loaded forces \u00b1 F x , \u00b1 F y , \u00b1 F z directly and cannot be loaded pure torques \u00b1 M x , \u00b1 M y , \u00b1 M z " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000367_978-3-642-53742-4-Figure3.23-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000367_978-3-642-53742-4-Figure3.23-1.png", + "caption": "Fig. 3.23 Model of the two-link manipulator with uncertain link lengths", + "texts": [ + " Note that for each evaluation of J\u0302d(ud), (M1 + 1) \u00b7 \u00b7 \u00b7(Ms + 1) simulations of the differential equation (3.12) are required. As a result, a set of Pareto optimal control sequences ud is obtained. By applying these control sequences to the technical system described by the differential equation (3.6), the system behaves Pareto optimally with respect to the prescribed quantity of interest (e.g. control effort) and performance in the presence of uncertain parameters. In this section, a robot arm modeled as a two-link mechanism (see Fig. 3.23) is considered for which the lengths L1 and L2 of the two links are assumed to be unknown or not measurable exactly. The mechanism consists of two coupled planar rigid bodies, where \u03b8i, i = 1,2, denote the orientation of the ith link measured counterclockwise from the positive horizontal axis. The system is controlled via two control torques denoted with u(t) = (\u03c41(t),\u03c42(t)), acting in both joints of the two time-dependent links (see [107] for a detailed model description). The goal is to determine a control sequence u(t) which steers the robot arm tip (modeled as the end point of the two-link mechanism) from a prescribed initial state x0 to a 82 S" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000451_1.4026589-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000451_1.4026589-Figure1-1.png", + "caption": "Fig. 1 Schematic of the equivalent-surface contact model (the mean height of the peaks is the reference plane with which the heights of the asperities are defined)", + "texts": [ + " The fluid traction per unit length of the line contact is obtained by integrating the fluid shear stresses over the Hertz area to yield Ff \u00bc \u00f0a a sf \u00f0x\u00dedx (10) An average temperature rise in the contact may be calculated DT \u00bc 1 2ah \u00f0a a \u00f0h 0 T\u00f0x; z\u00dedzdx To (11) This temperature rise includes the heat generated by the asperitycontact friction and, thus, may be taken as the macro flash temperature of the line contact. 2.3 The Asperity-Contact Part. As in Johnson [14], the asperity contacts in the parallel gap of the Hertzian region are modeled in the framework of the Greenwood\u2013Williamson equivalent rough surface model [29]. Figure 1 shows such a setting with the equivalent rough surface in contact with a rigid flat. Each asperity contact is modeled as a spherical frictional contact. Chang and Zhang [19] developed a basic model for such a contact of an elastic-perfectly plastic material and Chang and Zhang [20] further developed it to include the asperity-contact temperature and fatty-acids boundary lubrication. The model is further developed in this paper with a more complete description of the boundary lubrication by including a surface reaction film", + " Following the same derivation as in the asperity pressure, the modification is to replace d by d\u00fe dm throughout Eq. (29) since the junction growth would come sooner and become more severe with the normal approach in the asperity contact. 2.3.6 Collection of Asperity-Contact Variables. The asperitylevel variables are statistically integrated to obtain a number of mixed-lubrication variables contributed by the asperity contacts. Given a macro EHL film thickness h, the normal approach in an individual asperity contact, referring to Fig. 1, is given by d \u00bc za d (32) where d \u00bc h zs. The total asperity-contact area, load, and friction are calculated by Ata\u00f0d\u00de \u00bc gaAn \u00f01 d Aa/\u00f0za\u00dedza (33) wa\u00f0d\u00de \u00bc gaAn \u00f01 d paAa/\u00f0za\u00dedza (34) Fa\u00f0d\u00de \u00bc gaAn \u00f01 d saAa/\u00f0za\u00dedza (35) where An \u00bc 2a is the Hertz area per unit length of the line contact, ga is the density of the asperities, and /\u00f0za\u00de is the probability distribution of the asperity heights. For a gross fundamental study, the surface roughness may be assumed to be isotropic and the asperity heights of the equivalent surface may be taken to be of a Gaussian distribution /\u00f0za\u00de \u00bc 1ffiffiffiffiffi 2p p ra exp 0:5 za ra 2 \" # (36) where ra is the standard deviation of the asperity heights", + " Nomenclature a \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 8wRe=\u00f0pE0e\u00de p , Hertz half width of the line contact Aa \u00bc area of an asperity contact Aap \u00bc area of all asperity contacts undergoing plastic flow A% ap \u00bc percentage ratio of Aap=Ata An \u00bc Hertz area per unit length of the line contact Ata \u00bc total area of asperity contacts cf \u00bc lubricant specific heat c1;2 \u00bc specific heat of solid 1 or 2 Cf \u00bc fraction of the load carried by fluid d \u00bc distance between the rigid flat to the plane of the mean height of the asperities (see Fig. 1) E0e \u00bc 2E0, equivalent modulus defined for the EHL variables E0 \u00bc \u00f0\u00f01 2 1=E1\u00de \u00fe \u00f01 2 2=E2\u00de\u00de 1 , equivalent Young\u2019s modulus E1;2 \u00bc Young\u2019s modulus of solid 1 or 2 F \u00bc total friction force per unit length of the line contact Fa \u00bc friction force of asperity contacts Ff \u00bc fluid traction of the line contact FPI \u00bc Fus=\u00f02a\u00de, friction power intensity of the line contact G \u00bc aE0e, EHL material parameter h \u00bc bulk EHL film thickness H \u00bc surface hardness of the softer solid k \u00bc shear strength of the softer solid Kf \u00bc lubricant thermal conductivity K1;2 \u00bc thermal conductivity of solid 1 or 2 pa \u00bc asperity pressure pf \u00bc fluid pressure distribution pF \u00bc asperity pressure at the onset of plastic flow ph \u00bc 2w=\u00f0pa\u00de, Hertz peak pressure pm \u00bc w=\u00f02a\u00de, mean pressure of the line contact pY \u00bc asperity pressure at initial yielding ra \u00bc ffiffiffiffiffiffiffiffiffiffi Aa=p p , radius of the asperity contact area R \u00bc asperity radius Re \u00bc equivalent radius of the line contact T \u00bc temperature in the lubricant Ta \u00bc temperature in an asperity contact Taf \u00bc flash temperature of an asperity contact Tad \u00bc desorption temperature of an adsorbed boundary film To \u00bc ambient temperature Tr \u00bc softening point of a chemical boundary film u \u00bc entraining velocity of the contact us \u00bc sliding velocity of the contact u1;2 \u00bc velocity of surface 1 or 2 U \u00bc gou=\u00f0E0eRe\u00de, EHL speed parameter w \u00bc normal load per unit length of the line contact wa \u00bc load carried by asperity contacts W \u00bc w=\u00f0E0eRe\u00de, EHL load parameter x \u00bc coordinate along the line contact Y \u00bc yield strength of the softer solid z \u00bc coordinate along the film thickness za \u00bc height of an asperity zs \u00bc distance from the mean height of asperities to the mean level of surface (see Fig. 1) ac \u00bc Tabor constant used in Eqs. (23) and (31) _c \u00bc shear strain rate in the lubricant d \u00bc normal approach of an asperity contact (see Fig. 1) d1 \u00bc first critical normal approach at which plastic yielding occurs in an asperity d2 \u00bc second critical normal approach beyond which asperity becomes fully plastic d10 \u00bc first critical normal approach of a frictionless asperity contact d20 \u00bc second critical normal approach of a frictionless asperity contact dm \u00bc a normal approach defined to account for the EHL pressure preloading (see Eq. (24)) DT \u00bc flash temperature of the overall contact g \u00bc lubricant viscosity ga \u00bc number of asperities per unit surface area go \u00bc lubricant ambient viscosity j1 \u00bc K1=q1c1, thermal diffusivity of solid 1 j2 \u00bc K2=q2c2, thermal diffusivity of solid 2 K \u00bc h=r, specific film thickness or k-ratio of the contact la \u00bc sa=pa, friction coefficient of an asperity contact lad \u00bc friction coefficient of an asperity contact with an adsorbed boundary film lr \u00bc friction coefficient of an asperity contact with a chemical boundary film 1;2 \u00bc Poisson\u2019s ratio of solid 1 or 2 n \u00bc us=u, slide-to-roll ratio of the contact / \u00bc probability distribution of the asperity heights q1;2 \u00bc density of solid 1 or 2 qf \u00bc lubricant density r \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi r2 1 \u00fe r2 2 p , RMS roughness of the equivalent surface ra \u00bc standard deviation of asperity heights r1;2 \u00bc RMS roughness of surface 1 or 2 sa \u00bc asperity interfacial shear stress sf \u00bc shear stress in the lubricant so \u00bc Eyring stress of the lubricant, so 5:0 MPa [1] Bell, J" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001478_j.acme.2019.06.005-Figure10-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001478_j.acme.2019.06.005-Figure10-1.png", + "caption": "Fig. 10 \u2013 Presentation of calculation results for individual models.", + "texts": [ + " Due to the fact that the models representing the carriages did not differ from each other, it was decided to use one model four times (aggregation of the stiffness matrix according to Eq. (14)), which saved time needed to build other models. Next, the degrees of freedom of the presented substructures were reduced in accordance with the previously presented procedure, thus reducing the dimensionality of the guide system model from 156,600 to 48 degrees of freedom (9999%), and from 18,024 to 96 degrees of freedom (9946%) for the table. The next step was synthesis of the substructures to obtain the global model response. The results are presented in Fig. 10. As can be seen in Fig. 10, displacement areas obtained (using MSC Nastran \u2013 Linear Static SOL 101) on the basis of individual models differ slightly (less than 1% difference in maximum displacement value). The difference can be noticed when using the Ohta and Tanaka model due to the fact that this model only takes into account vertical stiffness. The use of the substructuring method to calculate the same models does not affect the accuracy of the results in any way. It does, however, significantly shorten the calculation time" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002004_1077546320964287-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002004_1077546320964287-Figure1-1.png", + "caption": "Figure 1. Schematic of 2DoF RP manipulator system. Note: 2DoF: 2 degree-of-freedom; RP: revolute\u2013prismatic.", + "texts": [ + " Defining the parameter estimation error by ~\u03b4 \u00bc \u03b4\u0302 \u03b4 (32) In the light of equations (31) and (32), the term _~\u03b4 is obtained as _~\u03b4 \u00bc _\u0302\u03b4 \u00bc q 1jsj (33) Constructing the Lyapunov functional by V2 \u00bc 0:5 s2 \u00fe q~\u03b42 (34) Using equations (9) and (33), the differential of the above Lyapunov function becomes _V 2 \u00bc s _s\u00feq~\u03b4 _\u0302\u03b4 \u00bc~\u03b4jsj\u00fe s f \u00f0\u03be\u00f0t\u00de; t\u00de _\u03be2d\u00f0t\u00de\u00fe \u03bb\u00f0\u03be2\u00f0t\u00de \u03be2d\u00f0t\u00de\u00de\u00fe\u03bce\u00f0t\u00deq=p\u00fed\u00f0\u03be\u00f0t\u00de; t\u00de\u00feg\u00f0\u03be\u00f0t\u00de; t\u00deu\u00f0t\u00de (35) Substituting equation (8) in the above equation, we can attain _V 2 \u00bc s _s\u00feq~\u03b4 _\u0302\u03b4 \u00bc~\u03b4jsj\u00fe s d\u00f0\u03be\u00f0t\u00de; t\u00de \u03b4\u0302sgn\u00f0s\u00de m1 \u03b2jsj 1 sgn\u00f0s\u00de m2jsj\u03b1sgn\u00f0s\u00de (36) Because sd \u2264 jsjjdj, equation (36) can be expressed as _V 2 \u00bc ~\u03b4jsj\u00fe sd\u00f0\u03be\u00f0t\u00de; t\u00de \u03b4\u0302jsj m1 \u03b2jsj 1 jsj m2jsj\u03b1\u00fe1 \u2264 ~\u03b4jsj\u00fe jd\u00f0\u03be\u00f0t\u00de; t\u00dej \u03b4\u0302 jsj m1 \u03b2jsj 1 jsj m2jsj\u03b1\u00fe1 (37) By adding and subtracting \u03b4jsj to and from the right-hand side term of equation (37), one obtains _V 2#~\u03b4jsj\u00fe jd\u00f0\u03be\u00f0t\u00de; t\u00dej \u03b4\u0302 jsj m1 \u03b2jsj 1 jsj m2jsj\u03b1\u00fe1 \u03b4jsj\u00fe\u03b4jsj \u00bc \u00f0\u03b4 jd\u00f0\u03be\u00f0t\u00de; t\u00dej\u00dejsj m1 \u03b2jsj 1 jsj m2jsj\u03b1\u00fe1 \u2264 m1 \u03b2jsj 1 jsj m2jsj\u03b1\u00fe1<0 (38) Therefore, relying on the adaptive terminal sliding mode tracker (30), it is concluded that the Lyapunov function (34) is decreased gradually. This completed the proof. In this section, the proposed control method is tested for the trajectory tracking of a uncertain 2DoF RP industrial manipulator as shown in Figure 1. The manipulator dynamics, with an assumption that the center of mass of each link is located at the center (Feng et al., 2002), can be represented mathematically as M0\u00f0q\u00de\u20acq\u00fe C0\u00f0q; _q\u00de _q\u00fe G0\u00f0q\u00de \u00bc \u03c4 \u00fe \u03c4d \u00fe Fd\u00f0q; _q; \u20acq\u00de (39) where M0 \u00bc \" M11\u00f0q\u00de M12\u00f0q\u00de M21\u00f0q\u00de M22\u00f0q\u00de # with M11 \u00bc Lm1l21 4 \u00feLm2q 2 2\u00feLm2l2l1\u00feLm2l21 4 /\u00fe I1\u00fe I2 M12 \u00bcM21 \u00bc 0; M22 \u00bc Lm2 C0\u00f0q; _q\u00de\u00bc 2 666664 Lm2q2\u00feLm2l1 2 _q2 Lm2q2\u00feLm2l1 2 Lm2q2 Lm2l1 2 _q1 0 3 777775 G0\u00f0q\u00de\u00bc 2 664 Lm1gl1cosq1 2 \u00feLm2g l1 2 \u00feq2 cosq1 Lm2g sinq1 3 775 The system state vector q\u00f0t\u00de \u00bc \u00bdq1\u00f0t\u00de; q2\u00f0t\u00de T 2R2 consists of angular position (q1(t)) and distance between links 1 and 2 (q2(t))" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000037_751476-Figure10-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000037_751476-Figure10-1.png", + "caption": "Figure 10: Comparison of proposed tooth contact stress and that from [7]. (a) The result of proposed method. (b) The result of benchmark FEM from [7].", + "texts": [ + " But using the FEM by ANSYS software, the computation time is about 12 hours for 80 analysis points in the mesh period. 0 5 10 15 20 3.5 3.0 2.5 2.0 1.5 \u00d7108 Roll angle \ud835\udf03 (\u2218) M es h sti ffn es sK (N /m ) Proposed method FE method Figure 8: Mesh stiffness from the proposed and FE method. 3.2. Case 2. The second example is a gear pair with tooth profile modification from [7]. The gear parameters are presented in [7].The tip relief amplitude is 10 \ud835\udf07mstarting at 20.9\u2218 and the lead crown relief is 5 \ud835\udf07m on both pinion and gear. The applied torque is 340N\u22c5m. Figure 10 compares the tooth contact stress of the gear pair from the proposed analytic model with the model of [7] which is calculated by the benchmark FEM. The proposed model compares well with the model of [7]. The overall shape of the tooth contact stress matches that of [7] rather well. Therefore, the gear mesh stiffness model is effective not only for a gear pair with tooth lead crown relief but also for a gear pair with tip relief. This also ensures the assumptions are acceptable for gear mesh stiffness model" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000943_1350650115611155-Figure17-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000943_1350650115611155-Figure17-1.png", + "caption": "Figure 17. Equivalent transformation of nonuniform preload.", + "texts": [ + " If only applying the axial preload in traditional way, the nonuniform distribution of axial and radial clearance caused by external loads cannot be eliminated and the contact status may not be improved. Therefore, the nonuniform preload should be considered which can be realized by applying three forces (F1, F2, and F3) on the outer ring of ball bearing. The coordinate system of bearing under nonuniform preload and external loads is shown in Figure 16. The equivalent transformation process for nonuniform preload is shown in Figure 17. In Figure 17, the action points of component preloads F1, F2, and F3 are uniformly distributing on bearing outer ring and their directions are along the increasing X-axis, which can be equivalent to an axial resultant preload Ftotal and an equivalent moment Meq. The specific calculation is shown as follows. The axial resultant preload Ftotal \u00bc F1 \u00fe F2 \u00fe F3 \u00f031\u00de The moment along with Z-axis and along with Y-axis Mz \u00bc F1 Rp\u00feF2 sin\u00f0 =6\u00de Rp\u00feF3 sin\u00f0 =6\u00de Rp \u00f032\u00de My \u00bc F2 cos\u00f0 =6\u00de Rp F3 cos\u00f0 =6\u00de Rp \u00f033\u00de where, Rp is the arm of force, its length is equal to the distance from action point of preload to bearing center" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000021_demped.2015.7303703-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000021_demped.2015.7303703-Figure1-1.png", + "caption": "Fig. 1. Cross-section of the eccentric misalignmen", + "texts": [ + " These studies cou accordance with the following topics; i measurement of unbalanced flux caused by a [3], calculation of the air gap flux density eccentricity and definition of the main factor [4], magnetic and mechanical problems re fault, diagnosis of eccentricity fault in ord system [5-8]. More recently, most of th eccentricity are about the modeling of the m reasons that cause eccentricity and differen and dynamic eccentricity [2, 9-11]. A. Eccentricity in Induction Motor Eccentricity is the case of non-uniform between the stator and rotor. Static, dyn eccentricity that is combination of both stati types of eccentricity. The stator is centered on first axis and th on second axis on the Fig.1. In a healthy second axes must be coincident. If the secon of rotation, the static eccentricity occurs. Dy appears to be a problem, when the first rotation. In addition, center of rotation between first and second axes, the mixed ecc this time. rel cage induction ethod (FEM) for ccentricity (static, e studied with the eccentricity on the x density, current ted. The torque entric cases was ibration and noise ccentric conditions , static eccentricity, harmonics, torque trial and daily life maintenance and could be worked ical, mechanical, failure analysis of years in terms of itutes a significant ty and unbalanced have been studied ld be classified in dentification and symmetric air gap as a function of s that cause UMP sulting from this er to prevent the e studies on the otors for different t degrees of static air gap that exist amic and mixed c and dynamic are e rotor is centered machine, first and d axis is the center namic eccentricity axis is center of can be anywhere entricity occurs at The inequality of the air connected with manufacturing operating conditions" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001161_tia.2018.2874350-Figure7-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001161_tia.2018.2874350-Figure7-1.png", + "caption": "Fig. 7. 3D magnet eddy current loss distribution under rated working condition (IphaseA=0 A).", + "texts": [ + " > REPLACE THIS LINE WITH YOUR PAPER IDENTIFICATION NUMBER (DOUBLE-CLICK HERE TO EDIT) < 4 The magnet eddy current loss is compared in Fig. 6 which can be divided into two different regions. The loss in the region A is lower in the conventional machine, while the opposite trend can be observed in the region B. As can be seen, the magnet eddy current loss under rated working condition, i.e. the current equals to 10A, is reduced by 81.5% in the proposed machine, which shows the great effectiveness of this method. In order to validate the proposed method further, the 3D FE results are carried out and compared in Fig. 7 and Table III. As shown, the 3D results also confirm the effectiveness of the proposed method. The slight reduction of magnet eddy current loss in 3D FE results is mainly due to the flux leakage in the end part of machine. In order to investigate in details, the spatial harmonics are analyzed and shown in Fig. 8, in which the representative harmonic, i.e. the 2nd spatial harmonic, is selected. In the figure, the first six legends indicate the amplitudes of the 2nd order flux density harmonics from different sources which are corresponded with the left y-axis, and the last two legends indicates the phase difference of the 2nd flux density harmonics between PM and armature fields in conventional and proposed machines which are corresponded with the right y-axis" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000702_1.5040603-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000702_1.5040603-Figure2-1.png", + "caption": "FIG. 2. (a) Model of a moving Gaussian beam to calculate the melt shape \u0393m(y,z) of a single track, (b) model to tilt this cross section by \u0394z = zL, to add a circular upper track shape SUT, to superimpose shapes dislocated by steps of wH, zL, (c) highlighted areas of the track cross section: remelting vertically AV, horizontally AH, combined AVH, generated track AT, (d) case of cavity formation AC.", + "texts": [ + " The losses can be divided into beam reflection losses and into heating of the material, except the above considered melting of the generated track AT. During the SLM process, the laser beam melts a wider total track area, Atot, than the generated one, AT. This melting cross section is an integral part of the here applied mathematical model to estimate the loss contributions originating from remelting of tracks and to analyze the shape of the tracks, as valuable information. The melting cross section shape \u0393m is estimated in a simplified manner by the model of a moving Gaussian source of heat,27 see Fig. 2(a), for which the temperature field can be calculated: T(x, y, z) \u00bc aPL(cpkwL) 1f (x, y, z, v): (3) (Cartesian coordinates x, y, z, absorptivity \u03b1, thermal diffusivity \u03ba, laser beam radius wL.) The distribution function f is f \u00bc \u00f01 exp ( H) (2p3)1=2(1\u00fe m2) dm, (4) with H \u00bc (X \u00fe r=2m2) 2 2(1\u00fe m2) \u00fe Z2 2m2 , (5) and with (time t) m2 \u00bc 2kt WL ; r \u00bc R kv ; X \u00bc x wL ; Y \u00bc y WL ; Z \u00bc z WL : The resulting melting cross section has a circular-like shape \u0393m, of width 2ym and depth zm. Although calculated for a semi-infinite sheet, the heat losses are sufficiently representative to qualitatively, and to some extent quantitatively, describe the conditions for overlapping tracks. As a next step, the cross section shape \u0393m is tilted, see Fig. 2(b), to an extent (angle \u03b1t) that the height of the left end of the melt pool, at \u2212ym, is one layer depth, zL, higher than at the right end, +ym. This new, tilted shape of the lower part of the track is denoted SLT(y,z). The upper shape of the track, SUT(y,z), is approximated by a circle. The radius RU [and center coordinates OU(x,y)] of the circle is chosen such that it sufficiently matches the mass balance. The calculations turned out to have little sensitivity to this radius. Overlapping of the tracks is modeled by duplication of this track shape, by displacing the shape by multiples of the hatch distance wH and layer depth zL, respectively. The model was programmed in MATLAB software. The main purpose of the model is to analyze the SLM-process through these overlapping shapes. While the basic cross section of a single track is simple, overlapping becomes rapidly complex, as visualized in Fig. 2(c). The newest track generated has an area AT, which is equal to the reference area, AT=wH\u00b7zL. The much larger area from the melt pool shape SLT causes remelting of previous layers, which can be subdivided into three categories, namely, vertical, AV, horizontal, AH, and combined remelting (losses), AVH. Beyond a limit of increasing layer depth or, as in Fig. 2(d), hatch width, cavities of growing area AC form between the tracks. This model enables manifold analysis because the overlapping cross sections can cause many different segments that experience multiple remelting in a complex combination. In the study, wH and zL were varied to identify and explain parameter trends by the aid of the model. One purpose of the model is to interpret resulting track cross section shapes that were obtained from experiments, such as in Fig. 1(a). The additive manufacturing experiments were carried out by a Renishaw SLM system at Fraunhofer IWS Dresden, Germany", + " For some of them, the adjacent tracks that caused this shape are indicated, V0 and V\u2013 representing remelting from a vertical track ( j = +1) in the same column, i = 0, or from a previous column, i =\u22121. H+ was formed by a horizontal subsequent track, i = +1 ( j = 0). SLT is the original shape from melting, which governs the lower shape of the element. For multiple remelting, these are the final, determining remelting tracks that shape the fingerprint. For large wH and zL, cavities occur. The detailed cavity shapes might be affected by surface tension driven flow. Figure 9 shows the percentage of vertical and horizontal remelting areas relative to the whole melted track, see also Fig. 2(c), for increasing deposition area (multiplying wH\u00b7zL, for the combinations studied). Increasing hatch distance reduces horizontal remelting AH, see, e.g., for zL = 40 \u03bcm, while for increasing layer depth the vertical remelting AV is lowered, both contributing to less melting losses and instead a growing track deposition At. Only for large wH and zL does cavity formation AC start, as also explained through Fig. 7. From the experiments, Fig. 3, a selection of typical resulting remelting cross sections AR can be seen in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000393_tmag.2015.2443137-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000393_tmag.2015.2443137-Figure1-1.png", + "caption": "Fig. 1. Configuration of the proposed DSFC motor.", + "texts": [ + " Furthermore, in fault conditions, the DSFC motor is able to operate in the switched reluctance motor (SRM) mode where the AlNiCo PMs are totally demagnetized, effectively enhancing the fault-tolerant capability. For fair performance evaluation, the output torque and speed range of the DSFC motor are analyzed and compared with that of a rare-earth doubly salient PM (DSPM) motor. The fault-tolerant capability is also verified by the theory analysis and the simulation results. Moreover, the experimental results of flux weakening and fault-tolerant operation are presented to examine the performance of the DSFC motor. Fig. 1 shows the configuration of the DSFC motor, which has several distinct features. It illustrates that there are no windings or PMs in the outer rotor, offering a high mechanical robustness. Furthermore, the concentrated armature windings are located in the outer layer of the stator, while the magnetizing windings and AlNiCo PMs are located in the inner layer of the stator. This double-layer stator topology is effective to prevent accidental demagnetization of AlNiCo PMs caused by the armature reaction" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000314_acs.2921-Figure6-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000314_acs.2921-Figure6-1.png", + "caption": "FIGURE 6 Switched RLC circuit", + "texts": [ + " On the other hand, the BLF-based controller in the work of Liu and Tong67 only can guarantee that the output constraint is not violated (Figure 2); however, the prescribed transient and steady-state tracking performance cannot be retained (Figure 3). The adaptive controller without any BLF and PPC can track the given command (Figure 2) by selecting appropriate parameters, while both the output constraint and the prescribed performance indexes cannot be guaranteed (Figures 2 and 3). The adaptive laws \ud835\udf121, \ud835\udf122, ?\u0302?1, and ?\u0302?2 in Figure 4 are shown to be bounded. Figure 5 shows the evolution of switching signals. This section applies the proposed control scheme to a switched RLC circuit as depicted in Figure 6. It has been widely employed in order to perform low-frequency signal processing in integrated circuits.2,24 The dynamic equations are given by \u23a7\u23aa\u23a8\u23aa\u23a9 x\u03071 = 1 L x2 x\u03072 = u \u2212 1 C\ud835\udf0e(t) x1 \u2212 R L x2 \ud835\udc66 = x1, (92) where \ud835\udf0e(t) \u2236 [0,+\u221e) \u2192 \u0304 = {1, 2, \u2026 ,}. For each s, s = 1, \u2026 , , Cs denotes the sth capacitor, x = [qc, \ud835\udf19L]T denotes the charge in the capacitor and the flux in the inductance, and u denotes the voltage. To handle unknown nonlinearities of RLCs, the universal approximators have been employed in the work of Long and Zhao,2 which, unfortunately, increase the complexity and obstacles in the application" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001144_iet-cta.2018.5380-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001144_iet-cta.2018.5380-Figure3-1.png", + "caption": "Fig. 3 Illustration of the industrial emulator setup with flexible belt described by (43)", + "texts": [ + " It is consisted of a drive disk and a payload disk and modelled as the following LTI system: x\u03071 x\u03072 x\u03073 x\u03074 = 0 1 0 0 \u2212209.6 \u22122 838.4 1.7 0 0 0 1 77.9 0.15 \u2212311.8 \u22122.47 A x1 x2 x3 x4 + 0 2306 0 0 B u + D\u03c6, (43) IET Control Theory Appl., 2018, Vol. 12 Iss. 17, pp. 2347-2356 \u00a9 The Institution of Engineering and Technology 2018 2353 where x := [x1, x2, x3, x4]T is defined as the state, x1 and x3 are the angular positions of the drive disk and load disk, respectively, and x2 and x4 are the angular velocity of the drive disk and load disk, respectively, as shown in Fig. 3. For simplicity, it is assumed that x is fully measurable and this assumption is reasonable because x1 and x3 can be obtained by using encoders while x2 and x4 can be estimated by designing two separate velocity observers [1]. The control input u is to drive x := [x1, x2, x3, x4]T to zero in the presence of twice differentiable perturbation \u03c6 \u2208 \u211d with a constant perturbation matrix D \u2208 \u211d4 \u00d7 1. Here, a simple sliding surface s = 1 1/2306 1 1 G x (44) is designed and its derivative is s\u0307 = Gx\u0307 = G[Ax + Bu + D\u03c6] " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000894_tro.2015.2489499-Figure4-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000894_tro.2015.2489499-Figure4-1.png", + "caption": "Fig. 4. Parameterization of a unit vector ui with respect to a given frame x, y, and z.", + "texts": [ + " In what follows, we only consider that we observe the leg direction ui , and not the leg edges in the image space, as the leg edges are only used as a measure of ui . Therefore, the problem is the same, except in the fact that we must consider the singularity of the mapping between the edges and ui , but this problem is well handled: these singularities appear when n1 i and n2 i are collinear, i.e., the cylinders are at infinity [14]. In the general case, the unit vector ui can obviously be parameterized by two independent coordinates, which can be two angles, for example, the angles \u03b1 and \u03b2 of Fig. 4 defined such that cos \u03b1 = x \u00b7 v = y \u00b7 w (where v and w are defined such that z \u00b7 v = z \u00b7 w = 0) and cos \u03b2 = u \u00b7 x. Thus, \u03b1 is the angle of the first rotation of the link direction ui around z, and \u03b2 is the angle of the second rotation around v. It is well known that a U joint is able to orient a link around two orthogonal axes of rotation, such as z and v. Thus, U joints can be the virtual actuators we are looking for, with generalized coordinates \u03b1 and \u03b2. Of course, other solutions can exist, but U joints are the simplest ones", + " If a U joint is the virtual actuator that makes the vector ui move, it is obvious that: 1) if the value of ui is fixed, the U joint coordinates \u03b1 and \u03b2 must be constant, i.e., the actuator must be blocked; 2) if the value of ui is changing, the U joint coordinates \u03b1 and \u03b2 must also vary. As a result, to ensure the aforementioned properties for \u03b1 and \u03b2 if ui is expressed in the base or camera frame (but the problem is identical as the camera is considered fixed on the ground), vectors x, y, and z of Fig. 4 must be the vectors defining the base or camera frame. Thus, in terms of properties for the virtual actuator, this implies that the first U joint axis must be constant w.r.t. the base frame, i.e., the U joint must be attached to a link performing a translation w.r.t. the base frame.4 However, in most cases, the real leg architecture is not composed of U joints attached to links performing a translation w.r.t. the base frame. Thus, the architecture of the hidden robot leg must be modified w.r.t. the real leg such as depicted in Fig" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000769_978-3-662-43893-0-Figure6.3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000769_978-3-662-43893-0-Figure6.3-1.png", + "caption": "Fig. 6.3 Principle of a relief ground profile blade", + "texts": [ + " The relief surface behind the primary cutting edge is set back against the direction of primary motion by the side relief angle (6), and the relief surface behind the secondary cutting edge is similarly set back by the side relief angle (7). The tip relief is inclined against the direction of primary motion by the tip relief angle (8) and the front face is additionally inclined to the plane normal to the direction of primary motion by the hook angle (9). These technological angles always relate to the shape of the cutting wedge irrespective of the geometry of the cutting edge holder. Relief ground profile blades are reground on the front face only. The amount of regrinding is towards the back of the blade (see Fig. 6.3), such that the blade profile remains unchanged. The sharpening or regrinding can be repeated until the remaining cross-section of the blade becomes too weak. With most types of tools, the relief ground blades remain in the cutter head for regrinding purposes (see Sect. 6.2.3.5). The stock amount SA to be removed must be bigger than the crater depth in the front face and bigger than the width of wear on the relief surfaces. The blade displaying the largest amount of wear determines the amount of regrinding for all the blades on a cutter head, since all profile blades must have the same tip height after re-sharpening" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000769_978-3-662-43893-0-Figure4.31-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000769_978-3-662-43893-0-Figure4.31-1.png", + "caption": "Fig. 4.31 Spring model of a bevel gear pair according to [NEUP83]", + "texts": [ + " Even in conjugate gears, this condition produces an uneven load distribution along the contact line. At the beginning or end of the contact, or if the meshing gears have a large overlap ratio, there are times when there is no load on certain segments, causing a significant stiffening effect in the immediate vicinity of the neighboring loaded tooth segments. Since several tooth pairs are simultaneously meshing in a gear set, they contribute to the transmission of load in the form of parallelconnected springs (Fig. 4.31). Each individual tooth pair takes over a part of the total load in proportion to its spring stiffness. The load distribution may thus be calculated with the aid of deflections; there is no explicit exact solution. General approach to calculate load distribution Approaches for the determination of load distribution must reflect the complex stress\u2013deflection conditions in tooth contact. 168 4 Load Capacity and Efficiency A model which has proved useful stems from cylindrical gear analysis, its basic concept being known from a series of publications [ZIEG71, SCHM73, OEHM75, SABL77, HOSE78, PLAC88]" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000527_6.2016-2102-Figure4-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000527_6.2016-2102-Figure4-1.png", + "caption": "Figure 4. Continuous vortex sheet", + "texts": [ + " The induced side force is11 \u2206FY = \u03b7q\u0304Svtavt u\u0304 V\u221e (14) where \u03b7 is the aerodynamic efficiency factor at the tail, Svt is the vertical tail area, avt is the lift curve slope of the vertical tail, q\u0304 = \u03c1\u221eV 2 \u221e 2 is the dynamic pressure, and u\u0304 = 1 hz \u222b hz 0 u(z)dz. Remark 2 The proposed formation controller mainly focuses on the position control, working as an outerloop controller. Hence, the rolling moment effects are not considered. In the continuous vortex sheet model, the lift distribution is assumed to be varying in the spanwise direction. The elliptical distribution is assumed for the lift distribution in our analysis. According to Lemma 2, the vortex strength \u0393 has the similar elliptical distribution as shown in Figure 4. The elliptical distribution equation is \u0393(y) = \u03930 \u221a 1\u2212 ( 2y b )2 ,\u2212 b 2 \u2264 y \u2264 b 2 (15) where \u03930 = 4L \u03c1\u221eV\u221eb\u03c0 . Based on this model, the induced upwash velocity at a certain point on the trailing aircraft wing in Figure 2 is w(s) = 1 4\u03c0 \u222b b 2 \u2212 b2 d\u0393/dy ( \u2206y + s+ b 2 \u2212 y ) \u2206z2 + ( \u2206y + s+ b 2 \u2212 y )2 1 + \u2206x\u221a \u2206x2 + \u2206z2 + (\u2206y + s+ b 2 \u2212 y)2 dy (16) 4 of 15 American Institute of Aeronautics and Astronautics D ow nl oa de d by K U N G L IG A T E K N IS K A H O G SK O L E N K T H o n Ja nu ar y 7, 2 01 6 | h ttp :// ar c" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001571_physreve.99.012614-Figure5-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001571_physreve.99.012614-Figure5-1.png", + "caption": "FIG. 5. Schematic illustration of the binary collision simulation. b means the impact parameter, which is defined as the minimum distance between the paths of two objects if their interaction is absent. This definition is identical with the distance indicated in the figure. r means the range of interaction. \u03b8in means the incoming angle of two objects.", + "texts": [ + " 4(c) inset], unlike the transition at small \u03b2, where the broad tail reduces the global R [green triangles in Fig. 4(c) inset]. In Fig. 4(d), where we fix the strength of volume exclusion \u03b2 = 1 and vary that of scattering exclusion \u03b1, only the monotonic profile is obtained. Order-disorder transition occurs at around \u03b1 = 0.38, and above this, R increases monotonically for increasing \u03b1 and converges to R = 1. 012614-7 Here we investigate what mechanism causes the effective alignment in this system with only exclusion effects. For this, we perform the binary collision simulation, illustrated in Fig. 5. In this subsection, we assume that the noise \u03bej is negligible. First, we put an intuitive consideration to get the intuition to clarify the origin of alignment. For simplicity, we restrict ourselves to the symmetric collision when the impact parameter b = 0 and the collision angle \u03b8in is much smaller than \u03c0 in this paragraph. See Fig. 5 for schematic illustration of the definitions of parameters. (The collision angle \u03b8in should be much smaller than \u03c0 for the discussion below, because, e.g., for \u03b8in = \u03c0 , the effect of \u03b1 becomes irrelevant due to the symmetry.) Figures 6(a), 6(b) and 6(c) explain what can occur when these exclusions are included. In the absence of mechanical exclusion, i.e., \u03b2 = 0, whereas \u03b1 > 0, Eqs. (5) and (6) for |qj | = 1 are symmetric under the time-reversal operation together with the inversion of the sign of each polarity qj , as long as S and the parameters in Jq j do not depend on time t and Jq j is the even function of the polarities qj " + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001240_tmag.2019.2894700-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001240_tmag.2019.2894700-Figure2-1.png", + "caption": "Fig. 2. (a) Flux distribution and (b) air gap flux density distribution with CeFeB forward magnetized.", + "texts": [ + " To solve this problem, magnetic barriers are added, which are circular-arc air slots in rotor core, and they make it difficult for NdFeB flux to pass through CeFeB. Moreover, the PM flux path of half pole is divided into three parts, as numbered in Fig. 1(b), which will improve the flux distribution of the rotor by adjusting the placement of barriers. With the d-axis current pulse provided by stator windings, the magnetization state of CeFeB can be changed, which will further change the air gap flux density of the machine. Thus, the speed range of the machine can be expanded by adjusting the magnetization state of CeFeB. Fig. 2 shows the flux and air gap flux density distributions of the machine when CeFeB is forward magnetized. It is indicated that most of the flux provided by NdFeB and CeFeB enters the air gap, and the flux distribution in the rotor core is uneven since the rotor core of half pole is divided into three parts by magnetic barriers, as shown in Fig. 1(b). The flux densities of Parts 1 and 2 are higher than Part 3, because the fluxes of Parts 1 and 2 are mainly provided by NdFeB, and the flux from CeFeB mostly goes through Part 3" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000080_j.engfailanal.2013.03.008-Figure13-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000080_j.engfailanal.2013.03.008-Figure13-1.png", + "caption": "Fig. 13. Shape of cracked gear (2nd order, a = 0.8, root crack).", + "texts": [], + "surrounding_texts": [ + "appears. The natural frequency is also affected by the position of the crack. From Table 1 it is clearly seen that the natural frequency is decrescent when the crack moves from the top to the root of the gear, but that the decrescent velocity of the natural frequency caused by crack position is faster than the drop in velocity of the natural frequency caused by crack size. When the crack is at the reference circle, the natural frequency of the gear is distinctly bigger than when the crack is at the root. The reason for this is that the damage to the structure\u2019s stiffness is bigger when the crack is at the root of the tooth than when the crack is near the top of the tooth. In addition, Table 1 shows the analytical result of two-dimension and threedimension modeling when the crack exists at the tooth root. As the two-dimension analysis neglects the coupling effect of the shaft direction for the gear, while the gear is quite thick actually; so, the analytical result of the two-dimension model is bigger than that of the three-dimension model.\nFigs. 13\u201315 show the second, fifth and seventh order vibration shapes of the cracked gear when the crack is located at the root of the tooth and crack size is a = 0.8. These clearly show the three-dimension amplitude at every point and the amplitude near the crack is much bigger than at an area without a crack; this is very important for identifying cracks. Figs. 16\u201318 show the second, fifth and seventh order vibration shapes of the cracked gear when the crack is located in the reference circle and the crack size is a = 0.4. It can be seen that the vibration shape is obviously different from that of the crack which is located at the root of the tooth, and that the amplitude is also different near the crack. Therefore, these should be taken into consideration when doing dynamic design.\nThe model of a single-tooth involute gear was established by apdl language in ANSYS, the transition curve was drawn using the precise transition curve equation. Crack locations were set at the pitch circle and tooth root, and the model of a single tooth is shown in Fig. 19. Crack form is a combination of triangular and rectangular forms, whose shape is shown in Fig. 20, where a is crack length and point 1\u20132\u20133\u20134\u20135 is the path defined in ANSYS to solve SIF. The load acts at the top of the tooth and the pitch circle. The position of the load at the top of the tooth is denoted by point A, the position of the load on top of the pitch circle is denoted by point E, and the position of the load under the crack in the pitch circle is denoted by point F. Then, based on this model, KI, KII can be determined using the displacement extrapolation method provided by ANSYS.\nThe main parameters for the gear were: module of gear m = 5 mm, number of teeth z = 20, modification coefficient x = 0, addendum coefficient ha = 1.00, bottom clearance coefficient c = 0.25, Young\u2019s modulus E = 2.06 1011 Pa, Poisson\u2019s ratio m = 0.3, crack length at pitch circle a = 0.8 (mm), and unit normal load w = 40 N/mm. The analyzed results of SIF under different locations of load are shown in Figs. 21 and 22. From Fig. 21, it can be seen that KI of point F is lower, and gradually approaches zero as the crack length increases, but KI of point A increases as a/m increases, KI of point E reaches a minimum", + "when a/m equals 0.45 approximately. While it is shown that KII of point E is bigger than the other points in Fig. 22. Overall, the trends in KI, KII are parabola, but the results at the top and bottom of the crack are different; that is, KI, KII at the top of the crack are bigger than when at the bottom of the crack. Based on analysis of these results the final fitted formulas for KI for point A and KII for point E can be found and are presented at the end of the paper.\nAssuming a crack length of a = 0.8 mm at the pitch circle, when the other parameters are fixed, and only the magnitude of the load is changed, the relationship between load and SIF can be obtained by analysis as shown in Figs. 23 and 24. From these figures, it can be seen that the relationship between load and SIF is linear, that is to say: KI = A1w, KII = A2w, where A1 and A2 are coefficients that are relevant to the main parameters of the gear, position of load, crack length, and so on. From the figures, it can also be seen that the rate of changed for KI, KII is fastest when the load acts at point E.\nSimilarly, assuming a crack length of a = 0.8 mm at the pitch circle and the other parameters are fixed, and only the module of the gear is changed, the relationship between the module and SIF can be obtained by analysis as shown in Figs. 25 and 26. From these figures, it can be seen that the relationship between the module and SIF is roughly conic; that is, KI = B1, KII = - B2, where B1 and B2 are coefficients that are relevant to the main parameters of the gear, position of load, crack length, and so" + ] + }, + { + "image_filename": "designv10_14_0001524_j.wear.2019.203106-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001524_j.wear.2019.203106-Figure2-1.png", + "caption": "Fig. 2. Technical drawing of the tested gear (All dimension are in mm).", + "texts": [ + " The connection to the output torque meter is established with a flexible coupling at the end of the gearbox. The brake mechanism, which represents the load, is connected to the output torque meter again with a flexible coupling. The desired movement is provided to the motor via the servomotor driver and the desired load level is adjusted using the brake mechanism. The device designed and manufactured in this study contains a gearbox in which gears with m \u00bc 1 mm module and z \u00bc 20 teeth could be inserted. A technical drawing of the gear tested in this study is presented in Fig. 2. The gears are fixed on a shaft with a tolerance of \u00d86h6 using a type A feather key in the radial direction and a shaft ring in the axial direction. A fixed amount of the same lubricant was added to the gearbox before each test. T. Tezel et al. Wear 440\u2013441 (2019) 203106 The gears were manufactured from 316L steel, AlSi10Mg, and Ti6Al4V materials using both the additive manufacturing and traditional method. Details of production techniques are summarized in Table 1. The cold-drawn 316L and Ti6Al4V materials were obtained as cylindrical bars with a diameter of 25 mm", + " However, since AlSi10Mg is a cast alloy, it was obtained by casting in a 25 mm ingot to ensure operability. After this stage, all materials were turned to the desired diameter and the teeth were cut thoroughly in a special machine that cuts teeth using the rolling method (a hobbing machine). Later, gears with the desired dimensions were processed, keyways were cut, and the process was finalized with grinding. The term \u201ctraditional manufacturing\u201d is used in reference to manufacturing on machines without computer aid Looking at Fig. 2, we can see that the tooth is cut using a conventional hobbing machine, whereas operations are performed on a conventional lathe. The situation in conventionally expressed casting and heat treatment is similar. AlSi10Mg is an important cast aluminium alloy that was developed based on the composition of the eutectic point in the Al\u2013Si phase diagram. The first step of the thermal process for the aluminiummagnesium-silicon system is homogenization treatment at 530 \ufffdC. The homogenization treatment took place over 24 h in order to ensure that the coarse Mg2Si deposits caused by the internal structure of the casting had dissolved completely in the solid solution" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001712_j.ymssp.2020.106823-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001712_j.ymssp.2020.106823-Figure1-1.png", + "caption": "Fig 1. Dynamic model of a pair of spur gears.", + "texts": [ + " Then the influence of the constant and modulated external excitations are systematically investigated, respectively. And several nonlinear motions in this gear system such as periodic motion, sub/sup-harmonic motion, bifurcation and jump phenomenon are demonstrated in more details through numerical studies. For the purpose of illustration, the auxiliary components of the gearbox are ignored, and a generalized model for the spur gear pair system is developed by assuming the gear main body as a rigid with the elastic active teeth. Fig. 1 shows the two gears are represented by their base, pitch and addendum circles. Ip;g and hp;g are the mass moment of inertia and the absolution angular displacement for the gears. The subscripts p and g refer to the pinion and wheel, respectively. Rop;og , Rpp;pg and Rbp;bg denote the addendum, pitch and base radius of the pinion and the wheel, respectively. The gears rotate with the driving rotational speed of the pinion _hp and the drag torque Tg acting on the wheel (a list of symbols is given in the Appendix)", + " To further establish its influence on the dynamical characteristics, the maximum values for the dimensionless DTE are captured in Fig. 24 over a wide range of rotational speed. For simplicity, the drag torque is constant and rotation speed assign ratio increases from 0.0 to 1.5%. As can be seen from Fig. 24, the main resonance locations near 4696 RPM, and the second and fourth harmonics of the mesh frequency located near 2510 RPM and 1336 RPM are predicted. The nature frequency for the gear pair in Fig. 1 happen at 4852 RPM from the FE/CM model, which is less than 5% bigger compared with the main resonance frequency [41]. The resonance could excite the high amplitude vibration and finally tooth separation occurs. Therefore, strong softening nonlinearity could happen so that these curves bend sharply to the left and then the double-valued region appears which means a wider range of contact loss occurs. The dynamic response remains on the lower branch for increasing speed sweeps starting from the low speed region, and then suddenly \u2018\u2018jump\u201d to the upper branch near resonance, as shown near 2529 RPM" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001646_s11771-019-4207-3-Figure11-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001646_s11771-019-4207-3-Figure11-1.png", + "caption": "Figure 11 Spherical hob error source", + "texts": [], + "surrounding_texts": [ + "Assembly error is inevitable during assembly process of spherical hob. Since the rotation of the spherical hob is the rotation of the analog spur gear, the error of the spherical hob can be equivalent to the installation error of the spur gear, such as the similar offset error \u2206E, the inclined slot angle error \u03b1, the tool radial error \u0394r and front and back displacement error \u03b4. Figures 12(a) and (b) show these equivalent installation errors. The coordinate systems including equivalent errors \u2206E, \u03b1, \u2206r and \u03b4 are shown in Figure 12(c). Here, the coordinate system Sh is rigidly connected to the theoretical position of the spherical hob, and Sa, Sb, Sc are the auxiliary coordinate system. The coordinate system Se is the actual position of the spherical hob. Table 1 lists the basic parameters of the face gear. The tooth surface equation of face gear including the equivalent installation errors in coordinate system S2 is expressed as: 2 w 2, ,( , , , , , , , ) ( , )h h h h al E r l E r M M , , , w w( ) ( ) ( ) ( , )a b b c c eM M r M r (14) where , 1 0 0 0 1 0 0 ; 0 0 1 0 0 0 0 1 h a E M , cos 0 sin 0 0 1 0 0 sin 0 cos 0 0 0 0 1 a b M ; , 1 0 0 0 0 1 0 r 0 0 1 0 0 0 0 1 b c M ; , 1 0 0 0 0 cos sin 0 0 sin cos 0 0 0 0 1 c e M ; J. Cent. South Univ. (2019) 26: 2704\u22122716 2711 For any point H in theoretical tooth surface of face gear, the unit normal vector 2 2 2 2( , , )H H H H x y zn n n n and position vector 2 2 2 2( , , )H H H Hr x y z are known. The position vector 2 2 2 2( , , )H H H H e e e er x y z of H' points on the actual tooth surface can be measured by means of coordinate measuring machine. The normal direction of the tooth surface in the extension theory is \u03beH between H and H'. The relations among 2 2 2 2( , , ),H H H Hr x y z 2 2 2 2( , , ),H H H H e e e er x y z 2 2 2 2( , , )H H H H x y zn n n n and \u03beH are presented as: 2 2 2 2 2 2 2 2 2 H H H H x e H H H H y e H H H H z e x n x y n y z n z (15) By solving Eq. (15), the tooth profile error caused by installation error \u2206E, \u03b1, \u2206r and \u03b4 can be calculated. Similarly, if a tooth profile error is known, the equivalent installation error can be determined." + ] + }, + { + "image_filename": "designv10_14_0001276_jas.2019.1911660-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001276_jas.2019.1911660-Figure1-1.png", + "caption": "Fig. 1. Kinematic Model of passive flexible needle.", + "texts": [ + " Section V includes the discussion and finally, concluding remarks are made in section VI. A. Kinematic Model of Bevel Tip Needle i1 i2 \u03b8 \u00d7 The kinematic model of the needle considered in our study is the generalized nonholonomic unicycle model presented by Webster et al. [24]. The insertion velocity and the rotational velocity are the two inputs of the passive needle kinematic model where it is assumed that the torsional compliance of the needle shaft is neglected. Thus, the needle has two DOF including both independent insertion and rotation. As shown in Fig. 1, frame P is the inertial world reference frame, and frame Q is the needle tip. With reference to universal frame P, parameter is the front wheel angle depending on insertion length. The position and orientation of frame Q, relative to frame P can be described precisely by a 4 4 homogeneous transformation matrix. gPQ = [ RPQ pPQ 0 1 ] \u2208 S E(3) (1) v = p\u03071i1+ p\u03072i2 (2) where p\u03071 = [0,0,1,0,k,0]T (which relates with insertion velocity) p\u03072 = [0,0,0,0,0,1]T (which relates with rotation velocity) k = tan\u03b8 l1 i1 Here, the curvature attained by the needle during insertion is given as . Due to insertion input , the needle moves forward along the z-axis direction and also curves its path around the y-axis of the body reference frame, which represents the orientation of the needle tip. The combined insertion and the passive curvature of the needle are depicted in Fig. 1. In our study, the spinning of the needle is the primary input which causes it to follow the straight path whereas, the insertion is considered as the secondary input and is assumed to be constant. In order to define the orientation between the frames P and Q, we have utilized the Z-X-Y fixed angles as generalized RPQ \u03b3 \u03b1 \u03b2 x,y,z,\u03b1,\u03b2 \u03b3 {P} {Q} pPQ = [ x y z ]T \u2208 R3 [ x y z \u03b1 \u03b2 \u03b3 ]T \u2208 U \u2282 R6 coordinates for parameterizing the rotation matrix . Let , and be the roll, pitch, and yaw of the needle in the plane, respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001683_tia.2020.2968036-Figure35-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001683_tia.2020.2968036-Figure35-1.png", + "caption": "Fig. 35. Test rig for 3PSC under GLT load.", + "texts": [ + " In order to verify the theoretical analysis by FE method coupled with the motion equation, experiments of 3PSC with a GLT load are conducted on a 12-slot/10-pole prototype SPMSM. The details of the prototype SPMSM are shown in Fig. 33 and the parameters are listed Table III. In order to conduct the experiments in a controllable manner, Sm2Co17 is selected as the PM material, which has a strong demagnetization withstand capability. The back-EMF is shown in Fig. 34 and the measured results matches well with the FE results. The experiments are simplified as a weight driving the shortcircuited prototype SPMSM from the static state. The test rig is shown in Fig. 35. When the 3PSC starts, the weight is released and drives the short-circuited SPMSM through a pully and string. Since the weight provides a constant load torque, this process can be seen as the 3PSC under the GLT load from initial speed 0 r/min and initial current 0 A. The load torque is 1 Nm and Authorized licensed use limited to: McMaster University. Downloaded on May 03,2020 at 02:38:07 UTC from IEEE Xplore. Restrictions apply. the total inertia is 0.004947 kg\u00b7m2. An oscilloscope is used to display the speed and current waveforms" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001679_j.mechmachtheory.2019.103769-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001679_j.mechmachtheory.2019.103769-Figure1-1.png", + "caption": "Fig. 1. Schematic comparison of the centered and dislocated positions of a radially loaded rolling bearing. (For interpretation of the references to colour in this figure legend, the reader is referred to the web version of this article.)", + "texts": [ + " Load distribution describes how the external load is distributed among the individual rolling elements, i.e. how many rolling elements are loaded and to what extend (rolling element load). This information can be used to determine the contact stress and thereafter the fatigue life of a rolling bearing. The radial displacement of a rolling bearing comprises half the internal radial clearance and the elastic deformation (alias radial deflection) of the bearing races on the inner and outer ring contact, i.e. how much a shaft is dislocated compared to its (idealized) centered position (cf. Fig. 1 ). \u2217 Corresponding author. E-mail address: schleich@mfk.fau.de (B. Schleich). https://doi.org/10.1016/j.mechmachtheory.2019.103769 0094-114X/\u00a9 2019 Elsevier Ltd. All rights reserved. Nomenclature A area of a cell mm \u00b2 a left limit of a uniform distribution [ X ] a o actual amplitude of the wave with wave order o \u03bcm a o , max maximum allowable amplitude of the wave with wave order o \u03bcm B width of a rolling bearing mm b right limit of a uniform distribution [ X ] C center point of a contact body - D diameter of a rolling element mm d shortest vector between two opposing surfaces - d IR bore diameter of the inner ring of a roller bearing mm d OR outer Bore diameter of the outer ring of a roller bearing mm E diameter of the raceway of an outer ring mm F diameter of the raceway of an inner ring mm F R radial load of a roller bearing N F resid residual of the equilibrium of forces N g remaining gap between two opposing surface points after contact \u03bcm g remaining gap between two opposing surfaces evaluated in discrete surface points - G R internal radial bearing clearance \u03bcm h initial gap between two opposing surface points before contact \u03bcm h initial gaps between contact bodies evaluated in discrete surface points - K influence coefficient matrix - k kernel function to model linear elastic material behavior mm/N k P factor to modify the profiling of a rolling element - L eff effective length of a roller mm N number of cells - n number of samples - n normalized direction vector - N ( \u03bc, \u03c3 ) normal probability distribution with mean value \u03bc and standard deviation \u03c3 - O considered harmonic wave orders - o order of a single wave - P profiling of a rolling element \u03bcm P ( X ) probability of event X - P ( Y | X ) conditional probability of output variable Y given input variable X - p contact pressure N/mm \u00b2 p contact pressure evaluated in discrete surface points - Q force on a rolling element N S surface point on the second contact body - s mathematical description of a surface (with geometric deviations) - s 2 X empirical variance of random variable X [ X 2 ] T X tolerance value of variable X [ X ] T surface point on the first contact body - t mathematical description of a surface (with geometric deviations) - T c value of a circularity tolerance \u03bcm u elastic deformation of contact bodies \u03bcm u elastic deformations of discrete surface points - U( a, b ) uniform probability distribution with left limit a and right limit b - V \u2217 complementary energy Nmm W waviness \u03bcm X random input variable [ X ] x position vector within the contact plane respectively point within the contact zone - X\u0304 empirical mean value of random variable X [ X ] x lb lower boundary limit of a random variable [ X ] x ub upper boundary limit of a random variable [ X ] X C cluster of random input variables [ X ] Y output variable [ Y ] Z number of rolling elements - \u03b3 X empirical skewness of random variable X - rigid body approach of two contacting bodies \u03bcm \u03b4 deflection of a roller bearing \u03bcm \u03b4 rigid body approach of two contacting bodies evaluated in discrete surface points - \u03b4Borg density-based sensitivity index - \u03b4R radial displacement of a roller bearing \u03bcm \u025b load distribution factor - \u025b maximum allowable residual of equilibrium of forces N \u03b8 threshold of a branch of a k-d tree [ X ] \u03d1o phase shift of the wave with wave order o \u03bcX mean value of the probability density function of random variable X [ X ] \u03be point within the contact zone - \u03c1 pearson correlation coefficient - \u03c3 X standard deviation of the probability density function of the normally distributed variable X [ X ] angular coordinate \u03c6 separation angle between rolling elements \u03d5 angular coordinate \u03c8 angular position of a rolling element contact zone - \u0304 surfaces outside the contact zone - \u03c9 cyc rotational speed s \u22121 Note: Bold symbols are higher dimensional variables", + " Since manufacturers of roller bearings have their own internal specification for the geometric dimensioning and tolerancing of roller bearing components and these specifications are usually confidential, assumptions must be made to determine the tolerance values. The tolerance values of the diameters of the bearing components ( T E , T F and T D ) are estimated based on the values of the internal radial clearance. Assuming that the diameters are symmetrically distributed, a simple approximation of the internal radial clearance (cf. Fig. 1 ) is used to generate an under-determined system of equations: I) \u03bcE \u2212 \u03bcF \u2212 2 \u03bcD = \u03bcG R II ) ( \u03bcE + T E 2 ) \u2212 ( \u03bcF \u2212 T F 2 ) \u2212 2 ( \u03bcD \u2212 T D 2 ) = G R , max (10) III ) ( \u03bcE \u2212 T E 2 ) \u2212 ( \u03bcF + T F 2 ) \u2212 2 ( \u03bcD + T D 2 ) = G R , min II \u2212 III ) T E + T F + T D = G R , max \u2212 G R , min (11) Publicly accessible data such as CAD-models or catalogue data provided by bearing manufacturers give orientation for the mean values of the bearing components. According to such information, the nominal values of the diameters of a NJ2306 are E = 62 ", + " The diametric deviations of the rolling elements, which coincide with the direction of the external load (D1 lying in the direction of the external load respectively D7 lying in the opposing direction), as well as the deviations of the raceway diameter of the inner ring (F) also have a perceivable influence on the internal radial clearance. All other input parameters seem to have marginal contribution, which, to some extent, are caused by numerical inaccuracies as well as some other issues arising when using k-d trees. Nevertheless, the clustering does not seem to dilute the information about the contribution of the diametric deviations of the rolling elements. The three main contributors also correspond to the input variables used in the simplified analytical formula for the calculation of the internal radial clearance (cf. Fig. 1 ). The slightly larger contribution of the deviations of the rolling element diameters (D) compared to the deviations of the raceway diameter of the inner ring (F) is caused by the assumed probability distributions of the input parameters (cf. Table 1 ). The uniform distribution of the rolling element diameters represents a pessimistic assumption about the manufacturing process of the rolling elements. Samples are generated within the entire tolerance range. The normal distribution of the raceway diameter of the inner rings, on the other hand, concentrates the samples around their mean value" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000098_cca.2014.6981467-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000098_cca.2014.6981467-Figure2-1.png", + "caption": "Fig. 2. Wingcopter side view, with propellers tilted by an angle \u03c7 = \u03c0 2", + "texts": [ + " Nice Sophia Antipolis, CNRS, I3S, UMR 7271, 06900 Sophia Antipolis, France. ducard@i3s.unice.fr 2Minh-Duc Hua is with ISIR, CNRS-UPMC, Paris, France. hua@isir.upmc.fr II. VEHICLE DESCRIPTION AND DYNAMICS The vehicle is sketched in Figs. 1 and 2. This hybrid flying machine is designed on the basis of a fixed-wing aircraft. On each wing, a pair of propellers is mounted with a mechanism that enables to tilt simultaneously the propellers of one pair around the yb axis by an angle \u03c7 \u2208 [0, \u03c0/2], as shown in Fig. 2. The angle corresponding to the right (resp. left) pair of propellers is denoted \u03c7R (resp. \u03c7L). Therefore, it is possible to control the direction of the total thrust generated by the 4 propellers about the yb axis. In addition, alike classical quadrotor UAVs, it is also possible to monitor the control torque vector by controlling the 4 propellers\u2019 speed, in addition to aerodynamic torques from control surfaces. The vehicle is equipped with 4 control surfaces, namely an aileron on each wing with a deflection \u03b4a1 and \u03b4a2, respectively, an elevator with a deflection \u03b4e, and two rudders with deflection angles \u03b4r1 and \u03b4r2, respectively" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000750_j.mechmachtheory.2018.11.002-Figure18-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000750_j.mechmachtheory.2018.11.002-Figure18-1.png", + "caption": "Fig. 18. Distribution of von Mises stresses for the face gear of: (a) Case 4, and (b) Case 5.", + "texts": [ + " An area of high contact stresses reaching a value of 1824 MPa is obtained for Case 2 due to the appearance of edge contact on the top edge of the face gear tooth surface. The maximum contact stress in Case 3 has been reduced drastically due to the application of a smaller effective longitudinal crowning at the top because of the application of a larger diameter for the circular cutter. Case 4 considers parabolic profile crowning for the reference shaper, which relieves the high stress on the tooth surface near the tooth root compared with Case 1, but there is still a large peak in stresses for the contact position 4, which is of 1454 MPa (shown in Fig. 18 (a)). The von Mises contact stress evolution for Case 5 is smooth all over the cycle of meshing, and yields a maximum value of 1145 MPa as shown in Fig. 15 . Figs. 19 and 20 show the evolution of bending stresses for the pinion and the face gear. The maximum bending stress for the face gear of Case 0 is the lowest of all the cases considered because of the load distribution along the contact lines. However, the variation of the maximum bending stress for all the cases of design is small. For Case 5, the parabolic profile modification for the reference shaper makes the contact pattern to be more localized and therefore the bending stress at the pinion root surfaces becomes larger than for the other cases" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000089_j.ijheatmasstransfer.2014.01.035-Figure15-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000089_j.ijheatmasstransfer.2014.01.035-Figure15-1.png", + "caption": "Fig. 15. Schematic of the analytical domain for heat conduction in the presence of moving heat source.", + "texts": [ + "3 provides the zero-th order temperature field as a representation of the long-term temperature distribution of the gearbox which by definition is time-independent. It should be noted, however, that this solution merely serves as an initial estimate of the temperature filed for the transient step of the calculation (Eq. (23)) in which the stationary-state temperature field, Ht\u00f0x; t\u00de is resolved. The system of rotating gears with a localized heat source at the point of contact resembles the problem of moving source on a semi-infinite slab (Fig. 15) for which, a closed solution of the following form can be established for the temperature distribution, H \u00bc _QK0 Pe2r exp Pe2f \u00f0C:1\u00de where K0\u00f0f \u00de is the modified Bessel function of the second kind and order zero and f \u00bc x t and r \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi f2 \u00fe y2 q . The streamwise deriva- tive of Eq. (C.1) (i.e. in the direction of motion) takes the following form, dH df \u00bc _QPe2 2f r K1 Pe2r \u00fe K0 Pe2r exp Pe2f \u00f0C:2\u00de Fig. 16 illustrates the variation of dimensionless temperature gradient along the dimensionless depth for different Pe and at two different locations with respect to the moving heat source" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001375_1350650116648058-Figure5-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001375_1350650116648058-Figure5-1.png", + "caption": "Figure 5. (a) Main casing in which are mounted the discs with specific set-ups to (1) impose an electric current through the contact and (2) regulate the jet oil flow Qjet by varying the relative pipe losses in an opened deviation; (b) the twin-disc contact; (c) various cylindrical and crowned disc profiles and roughness configurations.", + "texts": [ + " In order to generalize the understanding of WECs, the main drivers previously pinpointed15,52 will be further discussed and progressively transposed on a twin-disc machine (TDM) in attempts to better understand how tribomechanical (sliding kinematics) and tribochemical drivers (water contamination, electrical potentials, and lubricant additives) may affect the formation of WECs and premature failures in general. Considering the authors\u2019 previous WEC reproductions in ACBBs,13,15 the aim is first to transpose the respective contact parameters to reproduce similar surface microcracks and, second, to achieve WEClike embrittled propagation with the help of consistent tribochemical drivers. Twin-disc machine. In this study, RCF experiments are led on a TDM designed to simulate rolling contact fatigue between two discs with an enhanced control of the tribological parameters (Figure 5).75,76 The operating condition ranges explored in the 62 tests led in this study are detailed in Table 1. One disc is cylindrical and the counter one crowned to avoid edge-effects. The normal load N, the oil jet flow rate Qjet, and temperature Tjet are kept constant throughout each test. Let the slide to roll ratio be defined as SRR\u00bc 2*(V1 V2)/(V1\u00feV2); V being the respective surface speed referring to the driver disc 1 and follower disc 2 when SRR> 0. The discs are driven at a fixed SRR adjusted by specific combinations of pulleys", + " Those constant conditions are consistent with WEC reproductions that did not require any major transient events.2,13,15,16,32,33 Specimens. The discs are 35mm in radius and made out of the same martensitic through-hardened 100Cr6 (AISI52100) steel batch as the aforementioned WEC-affected ACBBs, thus presenting a top quality cleanliness and a hardness of 780 Hv (63 HRc).13,15 The crowned discs either have a 17.5mm crown radius to form circular contacts, or a 200mm crown radius to form elliptical contacts of ratio k\u00bc 5 (Figure 5(c)). The raceways are generally superfinished to a standard root mean square roughness Rq,std 0.02\u20130.03 mm, alike bearing raceways, or ground to Rq,rough 0.12mm, the grinding marks being always circumferential as on bearing raceways. Lubricants. The large majority of the TDM tests have been performed with lubricant A, a fully formulated ISO VG46 mineral gearbox oil previously used to reproduce WECs on ACBBs.13,15 Additional TDM tests have been conducted with lubricants B and C. at University College London on May 24, 2016pij", + "comDownloaded from They are respectively two similar fully formulated ISO VG46 and VG35 semi-synthetic gearbox oils also previously associated with WECs52 and known to contain high amounts of sulfur/phosphorus EP/AW and detergent additives. The chemical contents of lubricants A, B, and C determined by infrared emission spectrometry are detailed in Table 2. TDM tests previously performed with various in-house PAO8 base oil additive blends referenced D1, D2, D3, and D4 in Table 3, have also been cross sectioned and analyzed. Analyses. Each TDM test is either interrupted willingly after a given number of cycles or stopped automatically when a surface defect is detected by the electromagnetic sensors (Figure 5(a)). The discs are then successively rinsed in ethyl acetate, isopropanol and heptane ultrasonic baths. Surface analyses are led with a digital microscope and an optical 3D profilometer to assess the surface microcrack evolution, wear rates, and tribofilm formation. For tribofilm assessment, EDTA wiping77 and scanning electron microscopic (SEM)/energy-dispersive spectroscopy (EDS) analyses are also employed. Eventually, axial and circumferential cross sections, as well as fractographs are systematically performed looking for possible microstructural alterations and in particular WEC-like features", + " To account for the severe lubrication regime of the aforementioned CRTBs despite the standard superfinished raceways of the TDM discs, as well as for the evidence of the cage locally scrapping the lubricant off rolling elements in WECaffected ACBBs,15,52 some TDM tests have been performed under quasi-starved lubrication regimes. To do so, the rotational speed was reversed with respect to the oil jet and a deviation has been installed in the oil circuit to flush part of the oil directly back to the thermostatic bath, thus reducing the oil jet flow rate Qjet as schemed in (Figure 5(a. 2)).75 The ratio between Qjet and Qdev is set by progressively adjusting the pipe lengths. For 13 tests running under specific contact conditions (2GPa, reversed 11m/s, 7% SRR, lubricant A at 50 C), Qjet has been reduced to 0.13 L/ min, which is just sufficient to evacuate the generated heat and avoid scuffing. Transposition. By reducing the oil flow reduction, surface microcracks like those observed on WECaffected ACBB raceway borders have been reproduced with superfinished raceways on follower raceway in less than 2 106 cycles and for PH< 2GPa (Figure 7)", + "comDownloaded from with high concentrations of polar compounds will favor insidious discharges and an electro-chemical decomposition of the lubricant at nascent surfaces.71 It is thus supposed that tribocharging mainly catalyzes the lubricant decomposition and hydrogen generation, rather than the electro-thermal crack initiation concentrated at carbides as suggested by some authors.85 Transposition. An electrical circuit has been installed on the TDM to impose an electric current IC through the contact (Figure 5(a. 1)). The housing of the top and bottom shaft-disc assemblies is insulated with PTFE. The power supply is set to 15V DC 50Hz and the electrical potential UC between the rotating discs is applied by stripped copper wires brushing directly on the respective shafts. A variable resistor RV is added to the circuit to impose currents from 0.1 to 500mA. Preliminary results have revealed that the contact resistivity RC, thus supposedly the lubricant resistivity (since tests were led in full-film conditions), drops drastically when IC exceeds 1 mA (Figure 9), in agreement with measurements on full bearing apparatuses" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001515_j.addma.2019.100892-Figure20-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001515_j.addma.2019.100892-Figure20-1.png", + "caption": "Fig. 20: The final milling solution.", + "texts": [ + " Openings are designed based on the original part, and the struts are designed with a constant thickness of 20mm. Junctions are rounded off with a radius of 10mm. The front edge and the spindle attachment are designed according to the original C-frame. The design of the spindle attachment as an octagonal prism is advantageous for milling since both outer surfaces correspond to the thickness of the semi-finished part, and the process steps for milling of the remaining octagonal surfaces can be combined with other production steps. Chamferings are also designed according to the original C-frame. Fig. 20 shows the final milling solution. Fig. 21 shows the finite element model of the final SLM solution. The simulations use the material data for stainless steel alloy 316L (E = 190GPa, = 0.3) wherein the employed value for the modulus of elasticity constitutes the mean value of powder based values found in the open literature (Renishaw (2017), Concept Laser (2017), SLM (2018), Zhang et al. (2013), Marbury (2017)). The deviations of the literature values stem from different powder manufacturers, SLM process parameters, machinery and postprocessing strategies" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001925_978-3-030-46817-0_30-Figure31-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001925_978-3-030-46817-0_30-Figure31-1.png", + "caption": "Fig. 31. Plane warp validation of FEA model. Fig. 32. Stress von misses precision validation of FEA model.", + "texts": [], + "surrounding_texts": [ + "It can be seen from the obtained results that analysis with FEM and the software Algor gave good predictions of stress distribution during actual exploitations conditions of wheel hub. The simulation process has been carried out using MES module of Algor software. The numerical predictions of automobile parts behaviour can give us a lot of information and help with defining recent load increase in hub wheels with insurance of their high exploitation characteristics in the increasingly complex conditions of durable exploitation. Application of Algor software decisively supports numeric stress analysis and strength evaluation of wheel hubs since it effectively allows the simulation of the complex exploitation conditions under actual service loads by having installed the data and methods used and approved by experimental stress analysis. It therefore is a strong tool to improve the reliability of numeric strength evaluation. Modeling of exploitation characteristics of wheel hub and other automobile parts can be optimized by using the FEA. The finite element analysis method is becoming a common tool in automobile parts development. 144 I. Alagi\u0107 Based on presented experimental research and results analysis, following can be concluded: \u2013 Models of ball joint were designed in three-dimensional software Mechanical Desktop. Designed parts of model were transformed into FEA software Algor, where solid mesh was generated on the basis of surface mesh. \u2013 With FEM it has become possible to predict both the magnitude and the distribution of the stresses and displacement in ball joint assembly due to the tensile test of ball journal from housing of ball joint part No. 181 16 001 0 (pull force, magnitude 25 kN). \u2013 The quality of the mesh proved to be essential factor in performing successfully FEM analysis. The number, size and shape of the elements are of importance for the solution accuracy. The number of nodes influences especially regarding the simulation of tensile force. \u2013 The results of laboratory test performed by control device MR 96 were similar to the results obtained as result of finite element analysis. There are the effects of fact that laboratory conditions of testing were known during the FEA modelling and design of ball joint, especially designing of assembly technology for steering system of vehicle. The results of tests were allowed application of ball joint as proper part of suspension and steering system for passenger vehicle [10]. \u2013 The improvement of design and test of ball joint as well as its application can be achieved by using FEA combined with actual testing procedures as useful information for designing od assembly technology of the whole steering system of passenger vehicle [12]. \u2013 The conducted research has begun with creation of 3D-CAD solid approximate model in the form of a multi-body system, after that solid mesh was generated where all meshed elements assumed to be perfectly rigid, and in final stage of testing finite element analysis was performed using Algor software package. \u2013 From the presented results we can conclude that the distribution of deformation and stress do not exceed the upper limit value and that there are neither damages nor surface defects after performed tensile test. \u2013 The results of tensile test performed by control device MR 96 were closer to the results of FEM simulation. \u2013 Using FEM made possible to predict the whole tensile test of tie rod assembly. \u2013 The correctness and accuracy of computed results is still dependent on the selection related to various modelling parameters. Some of the most important aspects, such as boundary conditions or correct mesh and type of elements are performing a decisive role in achieving of correct results. \u2013 The mentioned conclusion is only valid for above defined working conditions and incorrigible estimated value of tensile force (F = 30000 N). Finite Element Analysis (FEA) of Automotive Parts Design as Important Issue 145" + ] + }, + { + "image_filename": "designv10_14_0000093_j.engfailanal.2013.02.003-Figure15-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000093_j.engfailanal.2013.02.003-Figure15-1.png", + "caption": "Fig. 15. Mode shapes and corresponding stress distribution (3rd mode).", + "texts": [], + "surrounding_texts": [ + "An acceleration transducer was fixed near the fractured gear on the bearing housing of the lubricating pump during lifespan tests, and the test signal in time domain was recorded. A fast Fourier transformation was performed to the test data before emergent shutdown, and frequency spectrum was obtained as Fig. 16. From Fig. 16, we could see the frequency component was very complex, so we would give a detailed analysis in the following. Firstly, we calculated the main frequencies that might exist in the lubricating pump. From Table 1 we could get Table 4 by simply multiplication. Then we could divide the big peaks in Fig. 16 into three series named \u2018\u2018A,B,C\u2019\u2019. Here \u2018\u2018A1\u2019\u2019 to \u2018\u2018A6\u2019\u2019 were multiple frequencies of 1500 Hz, and they rooted in exciting force caused by column pistons. For the value of \u2018\u2018A6\u2019\u2019 was much bigger, it was abnormal. And it was because of the disalignment of 9 column pistons according to previous experience. The disalignment might root in the fractured gear on the shaft. \u2018\u2018B1\u2019\u2019 to \u2018\u2018B3\u2019\u2019 were multiple frequencies of 2666.67 Hz, and they rooted in engaging force caused by 16 tooth. For the value of \u2018\u2018B1\u2019\u2019 was much bigger than \u2018\u2018B2\u2019\u2019 and \u2018\u2018B3\u2019\u2019, it was as ordinary vibration characteristics in gear pair. But there were cracks or fractures in the driven gear at that moment as a matter of fact. Inconsistency occurred because the gear pair was too small in the lubricating pump, so it was really hard to test the abnormal characteristics of gears accurately and directly by the acceleration transducer fixed on housing. \u2018\u2018C1\u2019\u2019 to \u2018\u2018C5\u2019\u2019 were multiple frequencies of 1030 Hz, and they rooted in exciting force caused by roller bearings. For the values of \u2018\u2018C2\u2019\u2019 and \u2018\u2018C3\u2019\u2019 were a little bigger, it was abnormal. And it was because of the disalignment of the two shafts, which might also root in the fractured gear on the shaft. Besides, there were a lot of small peaks in Fig. 16. They were caused by the modulation effect of rotation frequency [6]. Therefore, the frequencies were dense and complex in frequency spectrum. From the above test signal analysis, we could come to the following deduction. It was hard to test the abnormal characteristics of gears accurately and directly. But it was still useful to enhance the monitoring, because when we detected the outstanding value of multiple frequencies of the other exciting forces, such as the multiple frequencies of 1500 Hz and of 1030 Hz, we could estimate there may be disalignment in the two shafts, which might root in the abnormal status of gears." + ] + }, + { + "image_filename": "designv10_14_0000794_978-1-4939-2065-5_7-Figure7.16-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000794_978-1-4939-2065-5_7-Figure7.16-1.png", + "caption": "Fig. 7.16 (a) Snapshot of the polar dimer. An epoxy bead is attached to a steel wire to form polar (asymmetric) dimers. (b) Top view of the experiment. Two aligning electromagnets at constant current are orthogonal to two driving magnets, controlled by a computer. (c) The dimer with magnetization m experiences torque \u03c4mag to align with the magnetic field. Dimer orientation a\u0302 oscillates around \u3008a\u0302\u3009, which is parallel to Balign", + "texts": [ + " Propulsion enabled by fluid elasticity has been predicted for the three special cases of reciprocal motion: a flapping surface extending from a plane [130, 131]; a sphere which generates small-amplitude sinusoidal motion of fluid along its surface [8]; and a \u201cwriggling\u201d cylinder with reciprocal forward and backward strokes at different rates [88]. However, there remains little experimental demonstration, and such propulsion of free, finite-amplitude swimmers has been seldom studied. Here, we briefly discuss recent experiments [92,129] in which a single rigid object, in this case a dumbbell particle or dimer, is actuated in a reciprocal manner in very viscous fluids. In the experiments, the dimer such as the one shown in Fig. 7.16 is immersed in a fluid and repeatedly reoriented by a magnetic field. The effects of inertia are absent due to the high fluid viscosity (\u223c 10Pa \u00b7 s), resulting in Re 0.1 comparable to that of a swimming microorganism. By applying only magnetic torques, the apparatus reciprocally actuates just one degree of freedom in the system, the dimer\u2019s orientation a\u0302. For a purely viscous Newtonian fluid at low Reynolds numbers, the authors found no net motion because a\u0302(t) is cyclic; this is as expected. Experimental Setup Before diving into the discussion, let us briefly describe the experimental setup. More details can be found elsewhere [92, 129]. The artificial swimmer is a polar (asymmetric) dimer (Fig. 7.16a); symmetric dimers are also used for control but no net motion is expected. The polar dimer consists of a piece of carbon steel wire of length 2Rdimer = 2.5\u20133mm and diameter 230\u03bcm, with an epoxy bead of diameter 2Rdimer \u223c 500\u03bcm at one end. The dimer is then immersed in a fluid bath that is surrounded by four electromagnets; a schematic of the apparatus is shown in Fig. 7.16b. The dimer has orientation a\u0302 and is magnetized with moment m = a\u0302m, so that a uniform magnetic field B reorients it with torque \u00f8mag = m\u00d7B, as depicted in Fig. 7.16c. Working Fluids The dimer is immersed in a container (50 mm tall, 30 mm in diameter) of either Newtonian or viscoelastic fluid (Fig. 7.16b). The Newtonian fluid is a 96 %-corn syrup aqueous solution (by mass) with a kinematic viscosity \u03bc/\u03c1 of approximately 4\u00d7104 cSt. Two viscoelastic solutions are prepared: a dilute polymeric solution and a WLM solutions. The polymeric solution is made by adding 0.17 % (by mass) of high-molecular-weight polyacrylamide (PAA, MW = 1\u00d7 106) to a viscous Newtonian solvent (93 %-corn syrup aqueous solution). The solution has nearly constant shear viscosity of approximately 50Pa \u00b7 s and a relaxation time \u03bb of approximately 2 s" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002049_1464419320972870-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002049_1464419320972870-Figure3-1.png", + "caption": "Figure 3. Initial angle of defects on inner and outer raceway.", + "texts": [ + " The oil film damping coefficient is given as Cfilm \u00bc @F @u (8) where u is the velocity coefficient. The damping coefficients for inner and outer raceway can be shown in equations (9) and (10), respectively. Co \u00bc Cfo 1\u00feCso 1 1 (9) Ci \u00bc Cfi 1\u00feCsi 1 1 (10) where Cfo, Cso, Cfi and Csi are oil film damping coefficient and structural damping coefficient for outer and inner raceway, respectively. The comprehensive damping coefficient38 Cb could be thrived from equations (9) and (10), Cb \u00bc 1 Ci 1\u00feCo 1\u00f0 \u00de (11) The schematic of initial angular position of rolling element bearing is shown in Figure 3. The angle posi- tion of the i-th ball in inner and outer raceway relative to the Y axis is: hi\u00bc 2p N i 1\u00f0 \u00de \u00fe xs xc\u00f0 \u00de t inner raceway 2p N i 1\u00f0 \u00de \u00fe xc t outer raceway 8>< >: (12) where Z denotes the number of rolling elements, xc and xs refer to rotating speed of the cage and the shaft. Figure 3 shows clearly that the initial angular position of localized fault on inner and outer raceway are hi0 and ho0, respectively. The angle hIAD and hOAD between two localized faults on inner raceway and outer raceway are shown in Figure 4. The inner race is fixed with the shaft. When the shaft rotates with xs, the angular position of the ith defect on inner raceway changes as equation (13), hi di \u00bc xst\u00fe hi0 \u00fe i 1\u00f0 \u00dehIAD i \u00bc 1 : NDI (13) where NDI is the number of defects on inner raceway. When the outer ring is fixed on bearing pedestal, the angular position of the i-th defect on outer raceway remains constant, which is shown as equation (14), hi do \u00bc ho0 \u00fe i 1\u00f0 \u00dehOAD i \u00bc 1 : NDO (14) where NDO is the number of defects on outer raceway" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002413_tie.2021.3078388-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002413_tie.2021.3078388-Figure2-1.png", + "caption": "Fig. 2. Structure schematic diagram a) Exploded view of the PMSM [12] b) Resolver", + "texts": [ + " The effect of different origins on the frequency features of noise peaks in electric motors is summarized in Table III. Here, fHN is the frequency of the harmonic noise related to the fundamental and (6k\u00b11)th current harmonics. fSN is the frequency of the switching noise related to the high-frequency current near the switching frequency. The values of fHN and fSN are the same as those of the electromagnetic forces listed in Table II. An axial-flux external-rotor in-wheel PMSM with 30 poles and 27 slots is selected to validate the theoretical analysis. The exploded view of the PMSM is shown in Fig. 2a. The in-wheel motor adopts the single-stator-single-rotor topology. The PMs are bonded on the front cover and form the outer rotor system with the shell and end cover. The stator is fixed on the shaft through a stator tray. Fractional slot concentrated winding is adopted, with the advantage of high efficiency, high torque density, and high copper slot fill factor [29]. The position sensor is a variable reluctance resolver with 3 pole pairs, as shown in Fig. 2b. Table IV lists the main parameters of the axial-flux PMSM. Authorized licensed use limited to: California State University Fresno. Downloaded on June 23,2021 at 23:11:19 UTC from IEEE Xplore. Restrictions apply. 0278-0046 (c) 2021 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. TABLE IV MAIN PARAMETERS OF PMSM Description Value Description Value Pole/Slot number 30/27 Remnant flux density of PM 1" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002046_tia.2020.3040205-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002046_tia.2020.3040205-Figure3-1.png", + "caption": "Fig. 3. Schematic diagram of operational principle under only PM excitation. (a) Positive maximal armature coil flux. (b) Negative maximal armature coil flux.", + "texts": [ + " In essence, the operation principles of the proposed machines show similarity to that of other magnetically-geared machines, e.g., flux-switching machines and doubly salient machines. Namely, both armature and PM (or DC excitation) magnetic fields are modulated by the salient secondary poles in the air-gap. As a result, multitudes of synchronized field harmonic pairs are generated to produce stable thrust force, which is generally termed as the \u2018air-gap field modulation effect\u2019. Alternatively, its operation principles can be also expounded from the perspective of flux-switching principle, which is illustrated in Fig. 3. From this figure, PM flux in the armature coil clearly switches its direction with the movement of the secondary. In other words, the armature coil PM flux-linkage changes periodically along with the relative movement between the primary and the secondary due to the variation of air-gap permeance. As can be seen from Fig. 4, sinusoidal and symmetrical three-phase flux-linkage waveforms can be observed for the proposed machine, which means that if sinusoidal and symmetrical armature currents are fed, stable thrust force can be generated" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000894_tro.2015.2489499-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000894_tro.2015.2489499-Figure3-1.png", + "caption": "Fig. 3. (a) General robot leg and (b) its corresponding hidden robot leg when the vector ui is observed.", + "texts": [ + " The concept of the hidden robot model comes from the following observation: In the classical control approach, the encoders measure the motion of the actuator; in the previously described control approach (see Section II), the leg directions or leg edges are observed. Therefore, in a reciprocal manner, one could wonder to what kind of virtual actuators such observations correspond. The main objective of this section is to give a general answer to this question. Let us consider a general leg for a parallel robot in which the direction ui of a segment is observed (see Fig. 3(a)\u2014in this figure, the last segment is considered observed, but the following explanations can be generalized to any segment located in the leg chain). In what follows, we only consider that we observe the leg direction ui , and not the leg edges in the image space, as the leg edges are only used as a measure of ui . Therefore, the problem is the same, except in the fact that we must consider the singularity of the mapping between the edges and ui , but this problem is well handled: these singularities appear when n1 i and n2 i are collinear, i", + " 4 must be the vectors defining the base or camera frame. Thus, in terms of properties for the virtual actuator, this implies that the first U joint axis must be constant w.r.t. the base frame, i.e., the U joint must be attached to a link performing a translation w.r.t. the base frame.4 However, in most cases, the real leg architecture is not composed of U joints attached to links performing a translation w.r.t. the base frame. Thus, the architecture of the hidden robot leg must be modified w.r.t. the real leg such as depicted in Fig. 3(b). The U joint must be mounted on a passive kinematic chain composed of at most three orthogonal passive P joints that ensures that the link to which it is attached performs a translation w.r.t. the base frame. This passive chain is also linked to the segments before the observed links so that they do not change their kinematic properties in terms of motion. Note that: 4In the case where the camera is not mounted on the frame but on a moving link, the virtual U joint must be attached on a link performing a translation w" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002015_j.asr.2020.09.040-Figure27-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002015_j.asr.2020.09.040-Figure27-1.png", + "caption": "Fig. 27. Basic concept of thickness t and t and fold with holes.", + "texts": [ + " In order to evaluate the deployment characteristics of the proposed folding approach, we performed experiments considering the folding and unfolding of planar and curved membranes. In this section we describe our experiment setup, analysis and obtained results. The basic idea of our experimental setup is portrayed by Fig. 26. Basically, we evaluated the following types of spiral folding methods: the planar and curved spiral folding with and without multiple spirals under thickness considerations (Fig. 27). For deployment, a method that is less affected by the deployment intrinsics is desirable in order to contrast the deployment performance with respect to differences in folding methods. As such, inspired by the tensile deployment methods, we adopt a method of expanding the membrane by applying tension to the edges of the membrane model. We used threads with clips attached to the edges of the membranes, and loads inserted through a container below the membrane. Basically, we performed our experiments as follows: Fold a new membrane model", + " By inserting weights (with 9 g mass) into the container consecutively, tension is generated in the membranes, thus the deployment characteristics is obtained from the relationship between the applied tension (load), membrane expansion and the actual deployment radius. Also, the influence of the film thickness during folding is larger than the actual film thickness t when compounded over the spiral pattern. In other words, the thickness of compounding multiple layers is larger than the expected integer multiples of t. We rather consider a virtual thickness t > t (Fig. 27(a)) to accomodate the kind of buckling effect observed due to the winding of several layers of thick membranes (Fig. 27(b)). Since it is difficult to predict where the buckling effect is to occur, for simplicity we use reasonably small holes at the corners of creases (Fig. 27(c)) to avoid the kind of buckling effect due to winding, thus enabling to make a compact folding while ensuring the virtual thickness to be within the upper bound t (Fig. 27(d)). In line with the above-mentioned concepts, we evaluated the following types of planar layouts, as shown by Fig. 28: Model 1. Spiral fold with t \u00bc 0:1 mm.; this model is inspired by the conventional formulation of the spiral folding with planar membrane as shown by Fig. 2\u20136. Model 2. Multiple spiral folding, with t \u00bc 0:1 mm.; this model takes into account the multi-spiral folding pattern of the above layout, in which superimposition of the clockwise and counterclockwise orientations is realized by two concentric spirals" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002138_lra.2021.3068648-Figure2-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002138_lra.2021.3068648-Figure2-1.png", + "caption": "Fig. 2. Kinematics relationship of the continuum manipulator: (a) bending in YZ plane, (b) bending in XZ plane.", + "texts": [ + " For a general manipulator with contact blocks, the height profiles can be defined in Eq. (3): G4\u00d7n 2 = \u23a1 \u23a2\u23a2\u23a2\u23a3 h1 c1 h2 c1 \u00b7 \u00b7 \u00b7 hn/2 c1 h1 c2 h2 c2 \u00b7 \u00b7 \u00b7 hn/2 c2 h1 c3 h2 c3 \u00b7 \u00b7 \u00b7 hn/2 c3 h1 c4 h2 c4 \u00b7 \u00b7 \u00b7 hn/2 c4 \u23a4 \u23a5\u23a5\u23a5\u23a6 (3) where hi c1, hi c2 ,hi c3, hi c4 donate the height profiles of the i-th segment in the coordinate axes +X, +Y, \u2212X, \u2212Y direction. The deflection will be a composed bending when the continuum manipulator performs spatial deformation. Therefore, the different bending angles are used to represent the bending in different directions. As shown in Fig. 2(a), \u03b8i denotes the bending angle of the i-th segment in YZ plane, and\u03b2j represents the bending angle of the j-th segment in XZ plane. The relationship between the maximum bending angle \u03b8i\u2212max resulted from the self-collision and hi c is calculated: hi c = ( hb \u03b8i\u2212max \u2212 D1 2 ) tan \u03b8i\u2212max (4) The deflection angle will keep its maximal value when the self-collision occurs, even though the tendon is actuated continuously. The deflection angles of non-colliding segments are uniform. The relationship between the bending angle \u03b8i and the tendon length of i-th segment li without self-collision (as shown in Fig. 2(a)) is defined: li = 2 ( hb \u03b8i \u2212 rc cos \u03b4 ) sin \u03b8i 2 (5) The bending angles of the manipulator in YZ plane need to be calculated when the tendon length of the manipulator is given considering the contact blocks. The details are summarized in Algorithm 1. The basic idea of the algorithm is that Authorized licensed use limited to: Univ of Calif Santa Barbara. Downloaded on June 23,2021 at 02:43:41 UTC from IEEE Xplore. Restrictions apply. Algorithm I: Solve the Bending Angles of the Manipulator With CCMs", + "(5)); end if end for break else for i:=1 to n for j := 1 to length(column) if i == column(j) then \u03b8i\u2190\u03b8i\u2212max(i); li\u21902(hb/\u03b8i \u2212 rc/ \u221a 2) sin(\u03b8i/2); ll\u2190ll \u2212 li , \u03b8i\u2212max(i)\u21901; end if end for end for end if end the segment bending angle is equal to the maximum limited bending angle while checking if a contact event has occurred. For other segments which contact event has not occurred, an equal portion of the bending angle to each of the flexure segments is allocated. The bending angles of the manipulator in XZ plane are determined in a similar way, as shown in Fig. 2(b). The tip pose can be calculated through the formula: [ Rtip Ptip 0 1 ] = Tb n/2\u220f i=j=1 (Tx(\u03b8i)TdTy(\u03b2j)Td) (6) where, Rtip and Ptip are the tip rotation matrix and position with respect to the base frame. Tb denotes the transformation matrix from the base frame {Ob} to the first segment frame. Tddonates the translation along the disk. Tx(\u03b8i) represents the transformation from {Oi} to {Oi+1} for the i-th backbone. Ty(\u03b2j) represents the transformation from {Oj} to {Oj+1} for the j-th backbone, the transformation can be calculated: Tx(\u03b8i) = \u23a1 \u23a2\u23a3 1 0 0 0 0 c\u03b8i \u2212s\u03b8i (hb/\u03b8i)(c\u03b8i \u2212 1) 0 s\u03b8i c\u03b8i (hb/\u03b8i)s\u03b8i 0 0 0 1 \u23a4 \u23a5\u23a6 (7) Ty(\u03b2j) = \u23a1 \u23a2\u23a3 c\u03b2j 0 s\u03b2j (hb/\u03b2j)(1\u2212 c\u03b2j ) 0 1 0 0 \u2212s\u03b2j 0 c\u03b2j (hb/\u03b2j)s\u03b2j 0 0 0 1 \u23a4 \u23a5\u23a6 (8) Tb = [ I3\u00d73 [ 0 0 l ] T 0 1 ] ,Td = [ I3x3 [0 0 hd] T 0 1 ] (9) where, s\u03b8 denotes sin \u03b8, c\u03b8 denotes cos \u03b8" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0002026_tec.2020.3035258-Figure5-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0002026_tec.2020.3035258-Figure5-1.png", + "caption": "Fig. 5. Winding layout and configuration of HESFM with different slot/pole number combinations.", + "texts": [ + " The armature coil connections for 6/7 and 6/8 machines are similar to those of 6/4 and 6/5 machines, respectively, and the only difference is the interchange of phases B and C. The mechanical position of armature coils of the 12-slot stator is shown in Fig. 4(a). The armature coil connections for 12/10 and 12/11 machines are shown in Figs. 4(b) and 4(c), respectively. Similarly, the armature coil connections for 12/13 and 12/14 machines can be obtained by interchanging phase B with phase C in 12/10 and 12/11 machines, respectively. The winding layout and configuration of the HESFM with different slot and pole number combinations are illustrated in Fig. 5. It can be found that the connection between field coils remains the same for different slot and pole number combinations, whereas the connection between armature coils changes. The HESFM with different Ns/Nr are firstly optimized by the genetic algorithm embedded in the FEM software (ANSYS/Maxwell), and the optimization goal is the highest average electromagnetic torque. During optimization, the stator outer diameter, the stack length, the air-gap length, the length of the iron bridge, the rotor shaft diameter, the PM volume, the AW copper loss, the current density in FW slot (Jdc), and the slot filling factor for AW and FW slots remain the same and constant" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001923_j.mtcomm.2020.101115-Figure22-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001923_j.mtcomm.2020.101115-Figure22-1.png", + "caption": "Fig. 22. Dimension in the letter \u201cC\u201d (dimensions in inches).", + "texts": [ + " Relation between the flat lay-up and the final configuration It can be seen that the method of 4D printing of composites (4DPC) can produce composite structures of complex geometries such as those of the few letters of the alphabet. However, the correspondence between the original flat lay-up and the final configuration depends on the nuances in the manufacturing procedure, which is presented below. The case of the letter C is used to illustrate these aspects. The radii of curvature and other dimensions were measured using Fiji [17] image processing software. An example of the measurements is presented on Fig. 22. The dimensions in the original flat lay-up in Fig. 3 are used for comparison. \u2022 Starting from the bottom, there is a flat portion of 2 inch in length. Both the flat lay up and the final configuration show this feature. \u2022 There is a curved portion (about one quarter of a circle) with radius of R = 3.8 inch. This has a length of \u03c0R/2 = 6.0 inch. Fig. 3 shows the length of the un-symmetric laminate [0/903] of 6.5 inch. The difference is due to the uncertainty in distinguishing the transition between the curved part and the straight part. \u2022 In Fig. 22, there is a fairly straight part in the middle of length 3 inch. This corresponds to the [04] laminate in Fig. 3. Theoretically this is supposed to be a straight laminate. However observation of Fig. 22 shows that this exhibits a degree of twist. The reason for this is due to the fact this laminate has to support the upper portion of the letter C, which imposes a bending moment. Since this laminate is very thin (0.6 mm), it deforms under the bending moment. \u2022 The upper portion of the letter shows a curved part of radius 3.75 inch in Fig. 22. Fig. 3 shows a laminate with [0/903] lay up, 6.5 inch long. \u2022 In Fig. 22, the actual configuration is shown in comparison with the ideal configuration. The ideal configuration does not take into account the effect of the change from one type of lay up sequence to another. For example, in Fig. 3, there is a change from [04] lay up sequence to [0/903] and then back to [04]. This change can have an effect on the final configuration. This is the subject of a further investigation. Using the following values for the properties of a layer (E1 = 128.9 GPa, E2 = 6.3 GPa, G12 = 4.4 GPa, \u03bd12 = 0.33, \u03b11 = - 0.018 \u00d7 10\u22126/oC, \u03b12 = 29.37 \u00d7 10\u22126/oC), and laminate theory (equations 1\u20137), the radius of curvature for laminate of [0/903] stacking sequence is calculated to be R = 3.7 inch (9.4 cm). This agrees with the experimental values as shown in Fig. 22. The radius of curvature obtained here is about 3.7 inch (94 mm). This is for a laminate with stacking sequence [0/903]. For a laminate with stacking sequence [0/90], the radius of curvature is slightly higher. Thicker laminates with proportional lay up sequence will have higher radius of curvature. These radii of curvature can be too high if small structures of sharper radii of curvature are desired. For larger structures this may not be a problem. The letters presented in this paper only utilize laminates made up of 0\u00b0 and 90\u00b0 layers" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000750_j.mechmachtheory.2018.11.002-Figure4-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000750_j.mechmachtheory.2018.11.002-Figure4-1.png", + "caption": "Fig. 4. Coordinate systems used for determination of the generating surfaces of the circular cutters.", + "texts": [ + " The angle of extension of the edge profile is obtained as \u03be = cos \u22121 ( n E 1 s \u00b7 \u2212r E 0 s | r E 0 s | ) (11) and with it, the equations of a generic point Q in coordinate system S s of the left side edge profile can be obtained as follows: { x Q = x E 0 \u2212 \u03c1e cos (90 \u2212 \u03be + \u03bb) y Q = y E 0 + \u03c1e sin (90 \u2212 \u03be + \u03bb) (12) Here, \u03bb is the profile parameter of circular arc E 1 E 2 and \u03be the angle of extension of the edge profile as given by Eq. (11) . Coordinates of point E 0 in Eq. (12) correspond to the coordinates of vector r E 0 s (u, \u03d5 s ) represented in coordinate system S s . The equations for the right side edge profile can be obtained following a similar derivation. Two main design parameters will be considered to get the geometry of the circular cutters: (i) the mean cutter radius, R c , and (ii) the cutter profile tilt angle, \u03b4. Fig. 4 shows the definition of the circular cutter design parameters as well as the coordinate systems used for the determination of the generating surfaces. The profile of the shaper for both the active part and the edge radius were defined previously in coordinate system S s , rigidly connected to the shaper. We recall that the equation of the active part of the profile in S s is given by Eq. (8) and for the edge radius by Eq. (12) . Auxiliary coordinate systems S e , S f , and S g ( Fig. 4 ) are used to transform the profile of the shaper from coordinate system S s to coordinate system S , in which the profile of the circular cutter is defined. Based on them, the generating profile of h the circular cutter as a function of its profile parameter u is obtained in coordinate system S h as r h (u ) = M hg M g f M f e M es r s (u ) (13) Here, M es , M fe , M gf , and M hg represent auxiliary transformation matrices that allow performing successive algebraic operations to transform the definition of the generating profile from the coordinate system S s , fixed to the shaper, to coordinate system S h , fixed to the section of the circular cutter in which its generating profile is defined", + " Matrices M fe , and M gf and M hg are given by M f e = \u23a1 \u23a2 \u23a3 cos \u03b4 \u2212 sin \u03b4 0 0 sin \u03b4 cos \u03b4 0 0 0 0 1 0 0 0 0 1 \u23a4 \u23a5 \u23a6 (15) M g f = \u23a1 \u23a2 \u23a3 1 0 0 0 0 1 0 R c 0 0 1 0 0 0 0 1 \u23a4 \u23a5 \u23a6 (16) M hg = \u23a1 \u23a2 \u23a3 0 0 1 0 0 1 0 0 \u22121 0 0 0 0 0 0 1 \u23a4 \u23a5 \u23a6 (17) Matrix of transformation M hg allows axis z h to be considered the axis of rotation of the circular cutter. The section of the circular cutter was defined in coordinate system S h by Eq. (13) as a function of profile parameter u . Let us now consider a fixed coordinate system S d coinciding with coordinate system S h in Fig. 4 . The section of the circular cutter defined in coordinate system S h is rotated around axis z d an angle \u03b8 to form the generating surfaces of the circular cutter in fixed coordinate system S d as follows: r (c) d (u, \u03b8 ) = M dh (\u03b8 ) r h (u ) (18) where M dh (\u03b8 ) = \u23a1 \u23a2 \u23a3 cos \u03b8 sin \u03b8 0 0 \u2212 sin \u03b8 cos \u03b8 0 0 0 0 1 0 0 0 0 1 \u23a4 \u23a5 \u23a6 (19) The transformation from coordinate system S h to coordinate system S d is illustrated in Fig. 5 . Fig. 6 shows the relative position of the left side generating circular cutter and the virtual shaper that it mimics", + " For that, we will now consider the transformation from coordinate system S d to coordinate system S s \u2032 by considering the matrices M g \u2032 d , M f \u2032 g \u2032 , M e \u2032 f \u2032 , and M s \u2032 e \u2032 as follows r (c) s \u2032 (u, \u03b8 ) = M s \u2032 e \u2032 M e \u2032 f \u2032 M f \u2032 g \u2032 M g \u2032 d r d (u, \u03b8 ) (20) The transformation matrices needed to represent the generating surface of the circular cutter in the coordinate system S s are similar to those used to represent the cross section of the circular cutter and represented using the coordinate systems shown in Fig. 4 . Therefore, M s \u2032 e \u2032 = M \u22121 es , M e \u2032 f \u2032 = M \u22121 f e , M f \u2032 g \u2032 = M \u22121 g f , M g \u2032 d = M \u22121 hg . We recall that the shaper represented in Fig. 4 and the shaper represented in Fig. 6 have different purposes. The former is only for generating the circular cutter surface, however the latter is for describing the generating relationship between the circular cutter and the face gear. Although the cross section of the circular cutter coincides with that of the reference shaper, when the section is rotated about the axis of the circular cutter considering the cutter profile tilt angle \u03b4, the generating surface of the circular cutter might be below the surface of the reference shaper if the cutter profile tilt angle is not sufficient" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000209_j.procs.2017.01.204-Figure1-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000209_j.procs.2017.01.204-Figure1-1.png", + "caption": "Fig. 1. Three link manipulator model resembling human lower limb (After12)", + "texts": [ + " a hybrid proportional-derivative (PD) particle swarm optimised active force control (PSOAFC) of a three DOF lower limb exoskeleton system subjected to a form of disturbance. The system is aimed to rehabilitate the flexion/extension of the hip, knee, and ankle, respectively. The performance of the proposed control architecture shall then be compared to a standard PD controller by considering the same operating conditions of the former. 2. Lower Limb Dynamics The lower limb dynamics of the human limb and exoskeleton are modelled as rigid links joined by joints (bones) as depicted in Fig. 1. The model is restricted to the sagittal plane whilst the human-machine interaction is assumed to be seamless and free from frictional elements. In Figure 1 the subscripts 1, 2 and 3 illustrates the parameters of the first link (thigh), second link (crus/shank) and third link (foot) respectively, as well as the position of the hip, knee and ankle joints respectively. L is the length segments of the limb; Lc is the length segments of the limb about its centroidal axis and is the angular position of the links. The Euler-Lagrange formulation is used to derive the governing equations for the nonlinear dynamic system as below13 ( ) ( , )= + + ( ) + dD C G (0) where is the actuated torque vector, D is the 3 \u00d7 3 inertia matrix of the system, C is the centripetal and Coriolis torque vector, G is the gravitational torque vector, whilst d is the external disturbance torque vector" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0000527_6.2016-2102-Figure3-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0000527_6.2016-2102-Figure3-1.png", + "caption": "Figure 3. Rotation of the aerodynamic forces", + "texts": [ + " Similarly, the induced upwash velocity due to the left vortex filament is calculated by wl(s) = \u0393 4\u03c0 \u2206y + b+ s (\u2206y + b+ s)2 + \u2206z2 + r2 c [ 1 + \u2206x\u221a \u2206x2 + \u2206z2 + (\u2206y + b+ s)2 ] (8) The total induced upwash velocity distribution along the trailing aircraft wing span is w(s) = wl(s) + wr(s) (9) The total sidewash velocity at s = b 2 is computed by u(z) = \u0393 4\u03c0 \u2206z + z (\u2206y + b 2 )2 + (\u2206z + z)2 + r2 c 1 + \u2206x\u221a \u2206x2 + (\u2206z + z)2 + (\u2206y + b 2 )2 + \u0393 4\u03c0 \u2206z + z (\u2206y + 3b 2 )2 + (\u2206z + z)2 + r2 c 1 + \u2206x\u221a \u2206x2 + (\u2206z + z)2 + (\u2206y + 3b 2 )2 (10) 3 of 15 American Institute of Aeronautics and Astronautics D ow nl oa de d by K U N G L IG A T E K N IS K A H O G SK O L E N K T H o n Ja nu ar y 7, 2 01 6 | h ttp :// ar c. ai aa .o rg | D O I: 1 0. 25 14 /6 .2 01 6- 21 02 where z is from 0 to the top of the vertical tail at \u2212hz, denoting the distance from a certain point at the vertical tail to the tail bottom. The induced upwash will increase the angle of attack of the trailing aircraft by \u2206\u03b1. The aerodynamic forces of the trailing aircraft would be rotated as illustrated in Figure 3. The induced lift at a certain point on the trailing aircraft wing is d\u2206L(s) = 1 2 \u03c1\u221eV 2 \u221ec(s)Cl\u03b1(s)\u2206\u03b1(s)ds =\u21d2 \u2206L(s) = \u222b b 0 d\u2206L(s)ds (11) where c(s) denotes the chord at this point; Cl\u03b1(s) is the local section lift curve slope; \u2206\u03b1(s) is the induced angle of attack, calculated by \u2206\u03b1(s) = tan\u22121 ( w(s) V\u221e ) \u2248 w(s) V\u221e (12) The reduction in induced drag is due to the forward tilting of the lift as shown in Figure 3. So the drag reduction is calculated by \u2206D = L\u2032sin(\u2206\u03b1) \u2248 L\u2032 w\u0304 V\u221e (13) where w\u0304 = 1 b \u222b b 0 w(s)ds represents the averaged upwash velocity along the trailing aircraft wing. The induced side force is11 \u2206FY = \u03b7q\u0304Svtavt u\u0304 V\u221e (14) where \u03b7 is the aerodynamic efficiency factor at the tail, Svt is the vertical tail area, avt is the lift curve slope of the vertical tail, q\u0304 = \u03c1\u221eV 2 \u221e 2 is the dynamic pressure, and u\u0304 = 1 hz \u222b hz 0 u(z)dz. Remark 2 The proposed formation controller mainly focuses on the position control, working as an outerloop controller" + ], + "surrounding_texts": [] + }, + { + "image_filename": "designv10_14_0001114_978-3-319-79099-2_18-Figure7-1.png", + "original_path": "designv10-14/openalex_figure/designv10_14_0001114_978-3-319-79099-2_18-Figure7-1.png", + "caption": "Fig. 7 a Drop weight test rig used for testing samples of pomelo peel and giant sequoia bark. b Impact pendulum used for testing coconut endocarp samples. a Adapted from Thielen et al. (2015) and b adapted from Schmier et al. (2016) with permission", + "texts": [ + " While we analyzed the overall ability of pomelos to withstand drops from several meters height onto a concrete floor and on a platform equippedwith force transducers by performing free fall testswithwhole fruits (Fischer et al. 2010;Thielen et al. 2012), a more detailed analysis of the deformation behavior and energy dissipation capacity of fresh and freeze-dried peel tissue was only feasible with smaller peel samples. Therefore, we cut out cylindrical peel samples and tested them by performing drop weight tests (Thielen et al. 2015) using a custom-built test rig (Fig. 7a). This test rig allowed to record the force exerted by the impactor and transmitted through the sample to the force sensor over time. By tracking the velocity of the impactor, the coefficient of restitution was determined. The same experimental setup was utilized to analyze cuboid samples of the bark of an approximately 90 years old and 42 m high giant sequoia. The impact resistance of coconut endocarp samples was measured using an impact pendulumwith an instrumented impact hammer of 7.5 J work capacity (Fig. 7b). The hammer had an impact velocity v0 of 3.85 m/s when hitting the sample. With this test setup, it was possible to record the force F at the contact point of hammer and sample over the time t. By integrating twice the force with regard to time using Eq. 25, the displacement of the impact hammer while fracturing the sample was calculated (e.g., Thielen et al. 2015): s (t) t\u222b t1 (v0 \u2212 y\u222b t1 F (x) m dx) dy (25) with t1