[ { "image_filename": "designv11_101_0001016_s1052618812010025-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001016_s1052618812010025-Figure1-1.png", "caption": "Fig. 1. Structure of the hydro mount: (1) rubber shell ring; (2) body frame; (3) bearing plate; (4) upper plate; (5) lower plate; (6) membrane; (7) metal ring; (8) diaphragm; (9) tray; (10) base; (11) air space; (12) hole for the air space; (13) working chamber (piston space); (14) additional chamber (compensating space); (15) hole for the plate; (16) orifice channel.", "texts": [ " This hydro mount is designed for application in the vibration insulation systems of machines and construction engineering structures. The hydro mount consists of the static load bearing elastic element, a tapered shell ring under the dome of which the working chamber (head end) is located. There is another chamber (compensating space) behind the partition. Both chambers are filled with hydraulic fluid. The working and additional chambers are separated by the partition in which a membrane with orifice holes is located. The structure of the hydro mount is represented in Fig. 1. The elastic element, a tapered shell ring is cured to the body frame and bearing plate. Under the dome of the tapered shell ring, the working chamber (head end) is located, and behind the partition there is an additional chamber (compensating space). The partition consists of the upper and lower plates, between which the metal membrane is mounted. Both the upper and lower plates have a center hole. The membrane includes four orifice channels (holes \u22051.5) located uniformly by the periphery \u220520 and four orifice channels (holes \u22052", " The diagram of the partition is rep resented in Fig. 6. The reduced mass of the reviewed system will make mred = , where A is the area of the piston action, \u03c1, the liquid density, h, the height of the membrane channel, R, the outer radius of the partition, and r0, the radius of the central hole of the upper and lower plate. The frequency of setting of this system is defined by the formula fn = 1/2\u03c0(c/mred)1/2. Using this expression, let us write down it for the vibration isolator represented in Fig. 1 set to the fre quency f = 150 Hz: 150 = 1/2\u03c0(c/mred)1/2, where c is the static stiffness of the hydro mount, and mred, the reduced mass of the hydro mount. By analogy we obtain the expression of the vibration isolator at the frequency f = 45 Hz: 45 = 1/2\u03c0(c/700)1/2. Setting such a vibration isolator to the frequency 45 Hz is possible by increasing the number of the intermediate chambers of the vibration insulator three times, and increasing the width of the gap adjusted by the transfer coefficient" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003368_b978-0-08-098332-5.00006-1-Figure6.14-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003368_b978-0-08-098332-5.00006-1-Figure6.14-1.png", "caption": "Figure 6.14 Influence of stator supply voltage on torque\u2013speed curves.", "texts": [ " We established earlier that at any given slip, the air-gap flux density is proportional to the applied voltage, and the induced current in the rotor is proportional to the flux density. The torque \u2013 which depends on the product of the flux and the rotor current \u2013 therefore depends on the square of the applied voltage. This means that a comparatively modest fall in the voltage will result in a much larger reduction in torque capability, with adverse effects which may not be apparent to the unwary until too late. To illustrate the problem, consider the torque\u2013speed curves for a cage motor shown in Figure 6.14. The curves (which have been expanded to focus attention on the low-slip region) are drawn for full voltage (100%), and for a modestly reduced voltage of 90%. With full voltage and full-load torque the motor will run at point X, with a slip of say 5%. Since this is the normal full-load condition, the rotor and stator currents will be at their rated values. Now suppose that the voltage falls to 90%. The load torque is assumed to be constant so the new operating point will be at Y. Since the air-gap flux density is now only 0" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000031_amm.162.11-Figure4-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000031_amm.162.11-Figure4-1.png", "caption": "Figure 4 A current position of the mechanism and the velocities\u2019 triangle", "texts": [ " (22) As can be seen the angular increment jj ,1\u2212\u2206\u03b1 is monotonous decreasing with each step in the considered domain. The transmission\u2018s input-output ratio is: ' 1 / 1 / / 4 2 \u03c8\u03d5\u03c8\u03c8 \u03d5 \u03c8 \u03d5 \u03c9 \u03c9 ==== ddd d dtd dtd , (23) with: 1 1 ,1 ,1' \u2212 \u2212 \u2212 \u2212 \u2212 \u2212 = \u2206 \u2206 = jj jj jj jj \u03d5\u03d5 \u03c8\u03c8 \u03d5 \u03c8 \u03c8 , (24) 24 '\u03c9\u03c8\u03c9 \u22c5= (23\u00b4) and (25) 2 2 2 2 24 '\"' ' \u03b5\u03c8\u03c9\u03c8 \u03c9 \u03c8\u03c9 \u03c8\u03c9 \u03b5 +=+== dt d dt d dt d A . The kinematic analysis of the band mechanism may also be done by means of vectorial equations written for velocities and accelerations distributions [9]. For a current position of the mechanism (Fig. 4), the velocities of a point belonging to the imaginary cam ( CA ) in respect with the velocity of the point A of the input crank 2, can be expressed as: AAAA CC VVV += , (26) where: , , 4 2 CA A DAV C \u00d7= \u00d7= \u03c9 \u03c9 ./ ,/ 2 CA A DAdtdV ldtdV C \u22c5= \u22c5= \u03c8 \u03d5 (27) AAC V is the relative velocity between the two instantaneous superposed points CA and A . If the above relationships are divided with input angular velocity ( 2\u03c9 ) and the equation (26) is represented 090 rotated counterclockwise it results the triangle COAa , i" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001683_iros.2011.6094520-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001683_iros.2011.6094520-Figure1-1.png", "caption": "Fig. 1. Developed experimental walking robot", "texts": [ " But parametric excitation walking by knee-joint actuation has not been realized by an experimental robot. The purpose of this paper is to realize sustainable walking of an experimental kneed biped robot based on parametric excitation principle. For parametric excitation walking, a kneed biped robot may require more energy than the robot with telescopic legs. One advantage over telescopic leg is that a knee joint is simple and is realized by using a directdrive motor. The developed experimental biped robot is shown in Fig. 1, which has a servomotor at each knee joint and has no actuator at a hip joint. In parametric excitation walking proposed by Harata et al. [8], [9], not only walking performance but also walking ability strongly depends on the design of reference trajectories for knee joint angles. Furthermore, the reference trajectories are assumed to start from the instance of heel strike, but the developed robot does not have a ground sensor. Hence, we first show by numerical simulation that a robot can walk sustainably with a simple periodic reference trajectory for a swing leg knee", " Intuitively, gymnasts on a horizontal bar bend their bodies in inverse direction to swing up, which is a similar to the swing leg motion. Another example of inverse bending is an acrobot [13], which has similar mechanism to a swing-leg of our biped robot. The acrobot has an actuator on the knee and the robot can be controlled to swing up, which needs to increase mechanical energy. Harata et al. [9] also shows the advantage of the inverse bending. In this section, we introduce the experimental robot for parametric excitation walking that we have developed. The developed experimental biped robot is shown in Fig. 1. The robot has four parallel legs. Each leg has a semicircular foot and a knee joint with a servomotor. To restrict the robot movement in sagittal plane, the inside two legs and the outside two legs are synchronized, respectively. Upper legs are synchronized by structural constraint, such that the inside two upper legs and the outside two upper legs are connected by bars, and lower legs are synchronized by control input. A hip joint is free and has no actuators. Battery and controller are implemented on the inside two upper legs" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000204_amm.157-158.300-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000204_amm.157-158.300-Figure1-1.png", "caption": "Fig. 1 Structure Diagram of The Press Mounting Machine", "texts": [ " The press should also meet the following technical parameters, besides the above technical index. \u25cfThe overall dimension 1860\u00d71865\u00d73152 mm; \u25cfThe maximum press mounting force: 100KN; \u25cfRange 500 mm. Overall scheme design. The press mounting machine is composed of Dragon Gate frame, press mounting cylinder, the movable crossbeam, the center distance adjusting device, pressure head, guide pin, self-powered platform, lifting cylinder, guide rail, feeding dolly, the press controlled and tested system, which are shown in Fig. 1. 1-The press mounting cylinder.2-Dragon Gate frame.3-The movable crossbeam .4-The center distance adjusting device.5-The Error correction mechanism .6- The press head.7-The bearing outer ring.8-The guide pillar.9-The gearbox shell.10-The control cabinet.11-The feeding tray.12-The feeding dolly.13-The guide rail.14-The self-powered platform 15-The lifting cylinder. The Dragon Gate frame equipped with guide pin 13 and feeding dolly 14 is the body part of the equipment,through whose feeding automatic line the work-piece is delivered to the Feeding tray 12 of the press mounting device; under the instructions of control system, lift cylinder 15 carrying self-powered platform 14 moves download and locates feeding dolly on the frame workbench; After system, the press cylinder 1 drives the press head fixed on movable crossbeam 3 download movement, before this the bearing outer ring 7 has already been installed on the press head" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000747_amm.437.152-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000747_amm.437.152-Figure1-1.png", "caption": "Fig. 1(b) Schematic drawing of discrete micro-slip mode", "texts": [ " Having dry friction in the system complicated the dynamic analysis of the structure due to its nonlinear nature. Establishment of dry friction damping model has been extensively studied and two kinds of friction damping mechanisms have been widely used in general. basically, these models attribute to Coulomb which idealizes slip as taking place at a point. Macro-slip, which assumes entire contact surface as either slipping or stuck , models a friction damper as a spring and a Coulomb contact point which slips when friction force is greater than a certain value (Fig. 1a). This model was used to analyze the dynamic response considering dry friction damper by Griffin[3], Oden [4][5] which fits the experiment data. In micro-slip model, which is more realistic, the entire friction surface allow to slip at local area without gross slip. Iwan [6] proposed two micro discrete models that are parallel-series model(Fig. 1b) and series-parallel model which simulate contact surface as several individual Coulomb points and were proved to be effective to simulate the local slip on contact surface [7]. And then, Menq [8] developed a new continuous micro-slip model. By using this model, it is found that incorporating the effects of partial slip of the friction interface can result in significant reductions of the resonant response. Goodman and Krumpp use this model in their analysis of root All rights reserved. No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of Trans Tech Publications, www", " Slop of AB equals shear stiffness kt,; A1 is the maximum slip displacement of node A under normal force N1; \u00b5N1 is the maximum friction force and \u00b5 is poisson ratio. A,B,C,D are the critical point that contact status transform between stick and slip. Formula describing Fig. 4 is shown in equation (1). * 1 1 1 * 1 1 * 1 1 , 1* 1 1 1 * 1 1 ( ) ,0 , 2 cos (1 ) ( ) , , 2 d N f dd k x A N N N F k Ak A x N N \u00b5 \u03b8 \u03b8 \u00b5 \u03b8 \u03b8 \u03c0 \u00b5 \u03b8 \u00b5 \u03c0 \u03b8 \u03b8 \u03c0 \u00b5 \u03b8 \u03c0 \u03b8 \u03c0 \u2212 \u2212 + \u2264 \u2264 \u2212 \u2264 \u2264 = = \u2212 + \u2212 \u2264 \u2264 + + \u2264 \u2264 . (1) Dynamic equation of complex structures such as blade and disk(Fig. 1) can be wrote as: [ ] ( ){ } [ ] ( ){ } [ ] ( ){ } { } { } [ ] ( ){ } [ ] ( ){ } [ ] ( ){ } { } .. . 0 , .. . , sinBB B B B B N f DD D D D D N f M x t C x t K x t F t F M x t C x t K x t F \u03c9 + + = \u2212 + + = . (2) Where subscript B refers to the blade and D refers to the disk. The nonlinearity of the system resides in the dry friction force {FN, f} , acting at the contact surface. {F0}sin\u03c9t is the exciting force whose frequency is \u03c9. {N1}is the normal force at contact surface. For any individual pair of corresponding points, FN, f has piecewise linearity as Eq" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003240_978-0-8176-8370-2-Figure8.19-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003240_978-0-8176-8370-2-Figure8.19-1.png", "caption": "Figure 8.19: As in Figure 8.18 but in the G-case.", "texts": [], "surrounding_texts": [ "Lastly, we consider the full problem: the magnetic and electric fields are both present and neither parallel nor orthogonal. The intersection parabolas with the \u03be1\u03be3 plane are generic: their symmetry axis is always vertical but translated along \u03be1, and the concavity can assume positive or negative values. In Figures 8.18\u20138.23 three different cases are considered, which show that the analytical results (bottom-left subpictures) agreewith the numerical ones (bottom-right subpictures) very well. In Figure 8.24 we show how the FMI distribution changes when the angle between the two fields varies from zero (top-left subpicture) to \u03c0 (bottomright subpicture), with regular step \u03c0/8." ] }, { "image_filename": "designv11_101_0001652_ecj.10392-Figure6-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001652_ecj.10392-Figure6-1.png", "caption": "Fig. 6. Composite manipulator consists of M manipulators.", "texts": [ " According to several simulation results, this method successfully removes self-collisions even in cases where manipulators must pass through narrow holes, as discussed in Section 6. In this section the BFA is extended so that the paths of multiple manipulators that cooperatively achieve complicated tasks can be generated efficiently. First, let the composite manipulator M be a set of manipulators {M1, M2, . . . , MM} that cooperatively convey a workpiece while holding M different points P1, P2, . . . , PM on the workpiece, that is, each Mm holds position Pm on the workpiece by its top link as shown in Fig. 6. Here, each Mm consists of Nm-links, and its base link is fixed at position Bm. Also, to calculate the paths of the individual manipulators in parallel, as discussed later, link numbers are assigned so that the link numbers of each Mm increase from the top to the base links and the link numbers of the Mk become smaller than those of the Mm when k < m. For example, link numbers 1 and N1 are assigned to the top and base links of M1, and N1 + 1 and N1 + N2 are assigned to the top and the base links of M2, as shown in the figure" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003334_b11646-403-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003334_b11646-403-Figure3-1.png", "caption": "Figure 3. Physical modeling of flexural toppling failure in rock slopes with tilting machine before and after failure.", "texts": [ " In the process of this modeling the dimensions of the models are of paramount importance. If the dimensions are too small, then the failure will not occur. On the other hand, as was mentioned earlier in this paper, the samples are so brittle, hence, construction and handling of large samples are so difficult. Thus, the maximum height of slope models, in this study, was determined to be as of 42 centimeters. For this limitation, most of our models were stable during the test and only one model was unsuccessful (Figure 3). To assess the results of the existing theoretical approaches, the stabilities of these models have been analyzed with Majdi and Amini, and Amini et al methods. The comparative results are presented in Figure 4. As it can be seen from the figure, the results of the experimental and the theoretical models are fairly close to and are in a good agreement with each other. 774 \u00a9 2011 by Taylor & Francis Group, LLC D ow nl oa de d by [ C or ne ll U ni ve rs ity ] at 1 4: 46 3 1 O ct ob er 2 01 6" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002962_gt2013-94951-Figure6-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002962_gt2013-94951-Figure6-1.png", "caption": "Fig. 6. Overall Model of AFB", "texts": [ " The shape functions are expressed with respect to the element coordinate system ,, . The element stiffness matrix is defined by equation (9), abK=K 1212 (9) Where abK is defined by equation (10), 1 1 1 1 ~ jd\u03b6d\u03b6BDB=K b T aab (10) In equation (10), j is the determinant of the Jacobian matrix describing the element\u2019s connection between the natural and the physical space [11] Finally, in order to calculate the equivalent nodal forces, which are the result of the surface forces, equation (11) is used. d\u03b7jhN=f ss T asurf (11) Figure 6 illustrates the structural and CFD model geometry and mesh grid. As seen in figure 6, top foil\u2019s geometry is modeled as a cylinder\u2019s section with an angle of 355p . The one end of Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 05/04/2014 Terms of Use: http://asme.org/terms 5 Copyright \u00a9 2013 by ASME the top foil at 0 is fixed at yx \u03b8,\u03b8w,v,u, and z\u03b8 degrees of freedom in order to model the fixed end of the structure. Bump foil is modeled using 2-node spring finite elements. Each element has 2 degrees of freedom which are linear translations along the x and y axis of the global coordinate system for each of the 2 nodes" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000577_amm.157-158.53-Figure6-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000577_amm.157-158.53-Figure6-1.png", "caption": "Fig. 6 The plane model of Fig. 7 The schematic diagram Fig .8 The distribution map", "texts": [ " Partial discharge can be approximately treated as point source of ultrasonic signal wave. Considering the diffuse attenuation and the attenuation of sound after spreading some distance in the medium, ultrasonic sound pressure can be calculated as: x M N pp p e t cos\u2212\u03b1= \u03b2 (2) Where, cos\u03b2 is the cosine of p-wave refraction angle; tp is p-wave refractive index. In order to describe ultrasound propagation accurately, the generator stator model will be refined. Generator stator model is obtained based on Fig 4, as shown in Fig 6. generator stator of angle calculation of sensor distribution Two discharge points are set in the model: one is placed between stator slot and winding. According to Snell Law, the ultrasonic directly incident angle is within 29 degrees; the other discharge point is located between wingding and inner air gap, incident angle of which is within 43 degrees. In the model, when distance from discharge point to stator ektexine is different, rang of direct incident angle would also be different. When discharge is close to the stator ektexine, the direct incident wave range is smaller" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000480_s1052618811050153-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000480_s1052618811050153-Figure3-1.png", "caption": "Fig. 3.", "texts": [ " Vr 1/2a/r 1/4D3 2 r 2 \u2013( ), Vz az, a V/ h D3 2 /D2 2 1\u2013( )[ ].= = = The third stage ends upon contact of the material with the crown cavity bottom at the boundary r = 1/2D3 for the cloth thickness h3 = h1 \u2013 H3 + H2 and the stamp displacement s3 = H0 + H2 \u2013 (H1 + h3). Stage 4. The crown corners bounded by the cavity bottom and the fillet boundary are filled upon the plastic flow at the contact boundary with the cavity bottom. At the beginning of the crown corner filling, the contact length is small (Fig. 3), and in the middle of the crown cavity, there exists a plastic region with velocity field (28). The crown corners are filled without material extrusion into the fin. In the case of a linear approximation of the inclined boundary AB, we obtain the equations for the length of the free boundary of the cavity b and the contact length L from the condition of conservation of the blank volume, (29) where h is the current cloth thickness and h2 is the cloth thickness at the end of the second stage. The plastic region with the contact boundary L separated from the forge symmetry axis is simulated by the centered field of slip lines in the case of plane deformation with the singular point B and the free boundary with the inclination angle \u03b2 = " ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000215_amm.86.838-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000215_amm.86.838-Figure1-1.png", "caption": "Fig. 1 Front view of gear measuring machine", "texts": [ " The goal of this research is to develop a GMM which enables to measure not only working flanks, but also gear tooth root and bottom profiles in various patterns by installing a 3D scanning probe to it. Moreover, we try to realize higher measuring speed with scanning measurement and more precision measurement with the GMM developed than conventional GMMs and lower price measuring machine than conventional ones. GMM to be developed consists of conventional GMM (CLP35DDS, Osaka Seimitsu Kikai) and a 3D scanning probe (SP600Q, RENISHAW). The front view of the GMM is shown in Fig.1. It has one rotational axis and three axes perpendicular to each other. Radial axis consists of roller guide, ball screw and servo motor and the movement resolution is 0.1\u00b5m. Direct drive system it applied to tangential axis, which consists of roller guide and linear motor, and the movement resolution is 0.05\u00b5m. Axial axis consists of scraped casting guide, ball screw and servo motor and the movement All rights reserved. No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of TTP, www" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001134_amr.479-481.752-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001134_amr.479-481.752-Figure1-1.png", "caption": "Figure 1. Non-linear torsional vibration model", "texts": [ " Therefore, the related studies on this type of spatial gearing, beveloid gear, are less extensive and through. In the present study, a general nonlinear torsional time-varying vibration model of a beveloid gear pair is proposed. Also, numerical simulation results applying the proposed method are examined to gain a better understanding of the mesh damping and backlash nonlinearity on dynamics of crossed beveloid gears. The generalized lumped parameter torsional vibration model of a beveloid gear pair with small shaft angle is shown in Fig.1. The pinion and gear are modeled as rigid bodies with torsional displacements and their coordinates. Gear mesh is simulated by a pair of stiffness and damping elements along the line of action direction. Backlash and transmission error are also included. The equations of motion of the two degree-of-freedom torsional vibration model can be derived as ( ) ( ) pNLdmpNLdmppp TeefkeegcI =\u2212+\u2212+ \u03bb\u03bb\u03b8 (1.a) ( ) ( ) gNLdmgNLdmggg TeefkeegcI \u2212=\u2212\u2212\u2212\u2212 \u03bb\u03bb\u03b8 (1.b) Where pI and gI are mass momoents of intertias of pinion and gear, p\u03b8 and g\u03b8 are the torsional displacements of pinion and gear, pT and gT are mean load torques on pinion and gear, mk is mesh stiffness, mc is mesh damping, de is dynamic transmission error, NLe is unloaded transmission error, Le is the loaded transmission error, ( )NLd eeg \u2212 is a non-linear velocity function, ( )NLd eef \u2212 is a non-linear displacement function" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001704_appeec.2012.6307730-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001704_appeec.2012.6307730-Figure1-1.png", "caption": "Figure 1\uff0eElectronically controlled powertrain test bench", "texts": [], "surrounding_texts": [ "technical indicator of mechanical gearbox, which\ndirectly reflects the capacity of gearbox\ntransmission power. In this paper, concerning the\nhybrid electric bus transmission with synthetic box\nassembly, electronically controlled powertrain test\nbench is used to respectively test the transmission\nefficiencies of the gearbox assembly under different\ninput speeds, torques and oil temperatures. These\ntransmission efficiencies reflect the compared\nefficiencies of transmission assembly with different\ngears, thus to judge the transmission structure\ndeformation, motion intervention, abnormal wear\nand many other properties under different\nconditions. This test not only provides a reference\nfor the transmission efficiency test method, but also\nprovides experimental support for quality\nassessment of hybrid electric bus transmission with\nsynthetic box assembly.\nKeywords\u2014transmission efficiency; hybrid electric bus;\ntransmission; test method\nI. INTRODUCTION\nEnergy conservation and environmental protection is the eternal theme of automotive development. Improving the transmission efficiency of the auto parts is one major way to\nachieve energy saving. As the key component of the automotive powertrain system, transmission plays a very significant role in power transmission efficiency. The improving of transmission efficiency can not only contribute to reducing vehicle overall fuel consumption and increasing vehicle fuel economy, but also lowering the friction losses as well as the internal heat transmission, thus enhances the durability and reliability of gear and bearing [1]. Currently, researches in this area are sparse and most focused on the impact of the theory of traditional automobile transmission efficiency. Based on the electronically controlled powertrain test bench, a large number of experimental studies pertaining to the transmission efficiency of hybrid electric bus transmission with synthetic box assembly are firstly conducted in this paper, and transmission efficiencies of the transmission assembly at different speeds, torques and oil temperatures are obtained. Meantime, the derivative torques of the non-load transmission with a synthetic box are measured [2] - [5], which offers a reference for the hybrid electric bus transmission system design optimization.\nII. TEST BENCH AND INSTALLATION\nThis test employs electronically controlled powertrain test bench to conduct transmission efficiency test. The test bench is made up of an input electric power dynamometer and an output one, it shows in Fig.1and Fig.2. Under the control of the host computer, the dynamometers can work in different speeds, torques mode or\nvarious road conditions mode, the data of speed, torque, oil temperature and other parameters of the tested components are accurately recorded throughout the course of the experiment. Therefore, it can furnish the automotive powertrain design and development with the matched parameters and precise test data. In this\n978-1-4577-0547-2/12/$31.00 \u00a92012 IEEE", "experiment, the input electric power dynamometer is used for working conditions to realize the input, passing via the synthetic box assembly, and finally to the output electric power dynamometers to accomplish the transmission efficiency test on the hybrid electric bus transmission with synthetic box assembly\nIII. TEST CONTENT AND METHOD\nTransmission efficiency \u03b7 is the ratio of transmission output power to input power, as shown in Fig.3:\nThe transmission efficiencies are always less than 1, since the power loss occurs in the\ntransmission process of the power transmission assembly, including derivative power loss generated by the torque when in no-load and extra power loss when in load [6]. Therefore, this test consists of two aspects:\n(1) Test on the derivative torque of transmission with synthetic box assembly. Derivative torque refers to the resistance torque that the rotation input required while the output end is no-load. This experiment is mainly carried out to measure the transmission assembly derivative torques of hybrid electric bus transmission with a synthetic box assembly at different input speeds and oil temperature conditions.\n(2)The transmission efficiency tests with load Transmission efficiency tests of the hybrid electric bus transmission with synthetic box assembly are preceded respectively at different input speeds, torques and oil temperature conditions.\nIV. TEST CONDITIONS\nTo ensure the test being done smoothly, it is necessary to regulate the test conditions. Test conditions mainly based on this hybrid electric bus engine parameters, motor parameters, transmission parameters, test security, test bench capability and other factors.\n\u2460Determination of test speed: the peak motor speed of hybrid electric bus is 5500 r / min, so the input speed ranges from 800 r / min to 5500 r / min when tested;\n\u2461Determination of the test torque: motor peak torque of the hybrid electric bus is 240N.m, so the input torque ranges from 20N.m to 240N.m when tested;\n\u2462 Determination of experimental oil temperature: since the external oil temperature control device will cause the gear oil flow and fluctuation to the tested piece, thereby the efficiencies of the tested transmission components are affected [7]. Therefore, set this experiment not changing the oil circulation state within the transmission as the premise, through", "cooling fan or heater to keep the internal temperature constant, and select 80 \u2103and 120 \u2103 in this paper to conduct temperature test analysis.\n\u2463 As structural characteristic of the transmission assembly with synthetic box is the synthetic box and transmission cannot be separated, the test selects transmission in gear \u2162 with synthetic box to conduct representative test analysis for transmission efficiency.\nV. TEST RESULTS AND DATA PROCESSING\n\u2474Derivative torque of the transmission in gear \u2162 with synthetic box assembly\nThe input electric power dynamometer connects synthetic box's motor flange with transmission placed at gear \u2162 and the output device armed with no connection. The derivative torques of transmission with a synthetic box assembly is shown in Fig.4:\nAccording to the results, the following conclusions can be drawn: \u2460 the derivative torques of transmission assembly are approximately in proportional relationship with the input speeds. This is because the power loss of the oil mixing and the gear meshing increases as the speed rises. \u2461Under the same speed, derivative torque is slightly larger while the oil temperature is lower, and as the speed increases, the difference becomes larger. In general, the higher the oil temperature is, the lower the viscosity of the lubricating oil is, the power loss of mixing caused by the viscosity of the\nlubricating oil is smaller. However, when the viscosity of the lubricating oil is low, the oil film between gears is difficult to form, and the damage of gear meshing increases [8].According to the test data of derivative torques, when the transmission is in low torques rotation, the domination of power loss is mixing oil ,it poses a greater impact on the derivative torque.\n\u2475 Transmission efficiency of the transmission in gear \u2162 with synthetic box assembly.\nThe input electric power dynamometer connects synthetic box's motor flange with transmission placed at gear \u2162 . The output electric power dynamometer connects transmission\u2019s output flange, then the transmission efficiencies of transmission assembly with synthetic box are tested in a different input speeds, torques and oil temperatures, and the results are shown in Fig. 5 to Fig. 11." ] }, { "image_filename": "designv11_101_0001488_amc.2012.6197112-Figure6-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001488_amc.2012.6197112-Figure6-1.png", "caption": "Fig. 6. Schematic view of motions.", "texts": [ " Hand robot\u2019s motor is rotary geared motor and resolving power of the encoder is 1600 pulse/rev. Motor of master system is linear motors and position encoders whose resolutions are 100 nm. And C++ language is used for control and it is mounted by RTAI 3.7. A soft ball is used as a envrionment. Parameters of experimental system are shown in Table I In experiment, width of window is 10 s, and distance of next window is 1 s. In this paper, 4 motions are searched. Schematic view of motions that is used in experiment are shown in Fig. 6. The motions are shown in Fig. 6. Motion 1 and motion 2 are same trajectory, and the difference is only the force. Motion 3 and motion 4 are same relation, too. From motion 5 to motion 8 are same trajectory, and only the angle of force vector in last state is different. In this paper, each motion was searched 10 times. And, the motions are searched in series. The recognition results are shown in Table II. Table II shows that almost all motions were recognized with high rate. In spite of the fact that these motions are same trajectory, they were recognized with high rate" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001690_icra.2011.5979627-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001690_icra.2011.5979627-Figure2-1.png", "caption": "Fig. 2. Configuration at impact", "texts": [ " 2) Flight phase: During flight phases, the holonomic constraint force becomes zero and the dynamic equation becomes the same as Eq. (2) where \u03bb = 02\u00d71. 3) Collision phase: As previously mentioned, the telescopic-legs are assumed to be settled to the desired terminal lengths, L1 \u2192 Le and L2 \u2192 Ls, before the next collision. The inelastic collision with the ground is modeled as M (\u03b8\u2212)q\u0307+ = M(\u03b8\u2212)q\u0307\u2212 \u2212 JI(\u03b8\u2212)T\u03bbI , (6) JI(\u03b8\u2212)q\u0307+ = 02\u00d71. (7) In the following, the detail of JI(\u03b8\u2212) \u2208 R 4\u00d75 is described. As shown in Fig. 2, (x+, z+) is the tip position of the stance leg just after impact and its time derivative must be zero. The velocity conditions are then specified as d dt ( x\u2212 + Le sin \u03b8\u2212 + Ls sin(\u03b1 \u2212 \u03b8\u2212) ) = 0, (8) d dt ( z\u2212 + Le cos \u03b8\u2212 \u2212 Ls cos(\u03b1 \u2212 \u03b8\u2212) ) = 0. (9) These equations are arranged as x\u0307+ + Le\u03b8\u0307 + cos \u03b8\u2212 \u2212 Ls\u03b8\u0307 + cos(\u03b1 \u2212 \u03b8\u2212) = 0, (10) z\u0307+ \u2212 Le\u03b8\u0307 + sin \u03b8\u2212 \u2212 Ls\u03b8\u0307 + sin(\u03b1 \u2212 \u03b8\u2212) = 0. (11) Here, note that x\u0307+ = d dt (x +) = 0 and z\u0307+ = d dt (z +) = 0 because (x, z) is not updated. x\u0307+ in Eq. (10) and z\u0307+ in Eq" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001414_cobep.2013.6785219-Figure4-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001414_cobep.2013.6785219-Figure4-1.png", "caption": "Fig. 4. 3D prototype: (a) finite elements; (b) virtual model.", "texts": [ " Depending on the direction that the current is applied in the coils, the mover moves in one direction or another along of the x-axis. Numerical analyses of LSM are carried out by means of FEM using a commercial package. A three-dimensional model or virtual prototype was developed, since the distribution of magnetic flux density flux occurs in space, and the topology of the linear motor does not present symmetries to allow an analysis in two dimensions. A view of the virtual model of LSM is shown in Figure 4. The total number of finite elements in the virtual model is 125.365 and 145.546 differential equation for the solution. The magnetic flux density is analyzed in the air gap, in the region between the PM and the coil, when the PM is completely aligned in front of the coil that will receive excitation. These results are processed along the sampling line, as presented in the Figure 5. Moreover, the behavior of the magnetic field distribution in the area 0.5mm in front of the PM is monitored. Also, linear traction forces were computed using the Maxwell Stress tensor as function of excitation current applies in the coils, when the primary is completely aligned in front of the coil excited and keep static", " The magnetic flux density along the sampling line is simulated by means FEM and by experimental tests using a Hall effect sensor and the results are shown in the Figure 8, as function of the coil excitation. (a) (b) (c) with DC excitation of 2A. It is possible to conclude, from the figure 8, when the electrical current increases in the coil, it also increases the effect of armature reaction, but this effect does not cause a significant difference in the magnetic flux density average, making their average value in 0.332T. Also, by means FEM, it is possible to analyze the magnetic flux distribution in the area presented in the Figure 4b; it is important to notice that this analysis is really important to verify if the ferromagnetic material is magnetically saturated. Figure 9 and 10 present these results as function of the coil excitation. From the figures 9 and 10, it was realized that the magnetic flux density is not significantly altered, which had been verified in the analysis of figure 8. Figure 10 shows that the region where is concentered more magnetic flux density is in PM, because it is the magnetic flux source. Another important analyzis is about the behavior and distribution of the magnetic flux density in the area closer to the permanent magnetic, in the 3D view" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003197_s0368393100068802-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003197_s0368393100068802-Figure1-1.png", "caption": "FIGURE 1.", "texts": [ " What the missile had to do, therefore, was so to interpret the radar signal as to reckon its displacement from the beam axis and then to make the manoeuvres necessary to null this. This problem is very much like that of the autotracker on the ground, which has to reckon the (angular) displacement of the target from the scan axis and then move the latter to nullify it. How does the ground radar do this? In principle the mechanism is as follows. *A. C. Cossor Ltd. 348 By rotation of an off-set dipole, the polar diagram of the aerial is made to rotate (at \" scan\" frequency) about a skewed axis. In Fig. 1, OA is the axis of rotation and OM, the axis of the polar diagram which rotates about it, being offset by angle a- (the \"split\" angle). If the target is on sight line, this means that the target (7\\) is on the axis OA in which case no scanfrequency modulation results. However, for a target T2 off axis by angle s a finite scan frequency modulation appears which is a measure of the magnitude of s. The phase of the modulation relative to dipole rotation measures the direction of the displacement. When this problem is transferred to the missile it is seen that, when this is within the radar beam but off axis, it has quite similar intelligence", " This information must be supplied either over a separate radio link or by some form of \" coding \" applied to the ground signal. In \" Brakemine,\" use was made of the intrinsic coding provided by the fact that the plane of polarisation rotates with the ground dipole and so reveals its position to the missile. This condition inevitably contains a 180\u00b0 ambiguity, which can only be resolved on the ground, but will thereafter remain resolved so long as the signal is uninterrupted. The mathematics of this process is as follows. In Fig. 1 the right hand portion is a section normal to the beam axis OA through the target T2. AT2 is the dis placement vector of the target from the beam axis, and AZ is a fixed reference direction, e.g. AZ may lie in a vertical plane. AP denotes the polarisation plane which rotates at the scan-frequency w, rj denotes the direction of the displacement vector and \u00a3 the plane of the receiv ing dipole on the missile. We write &=\u2022\u00bb?-\u00a3. The received signal is then E{l+mcos(at-rj)} |sin(wf-\u00a3)| of which the fundamental component works out to 2 =- Em{(3 - cos 28) cos (wr - rj) + sin 28 sin (at -tj)} and the second harmonic to 4 - Ecos2(wf-\u00a3) The latter is independent of m and, subject to the 180\u00b0 ambiguity, is a perfect synchronising signal" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001012_2012-01-1936-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001012_2012-01-1936-Figure3-1.png", "caption": "Figure 3. Solid-axle movement", "texts": [ " This large lateral acceleration is simulated for actual driving, and thus enables the examination of a preferable steering control method technology for drivers. The driving simulator is based on CarSim (Version 6.06) (Mechanical Simulation Corp. (USA)), which is a full vehicle movement simulation software package. Figure 2 shows a schematic diagram of the vehicle model. Table 1 lists the main components in the vehicle together with the number of degrees of freedom associated with each component. For example, the rear axle is rigid and has two freedom degrees: vertical movement and rotation of the axle (see Figure 3). Details are provided in a previous report [6,7]. Table 2 shows the parameters for the vehicle model used in the experiment. The model of the vehicle considered herein has an FR layout (front engine mount and rear drive), which easily enters the drift state. The available tire cornering force follows the general tire characteristics shown in Figure 4. The tire characteristics during driving and braking were assumed to follow the data shown in Figure 5. By simultaneously considering the slip angle and the slip ratio, the friction circle concept can be applied based on the combined characteristics of CarSim" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000316_amm.351-352.250-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000316_amm.351-352.250-Figure1-1.png", "caption": "Fig. 1: The typical curves of metal-to-metal of test arrangement frictional factor under different pressure Fig.2: Test arrangement plan", "texts": [ " Because of the non-linear relationship, friction coefficient always changed with the load. 2. Bonding on the clean contacted metal pairs could lead to high friction factor and high wear rate. Bonding on the clean contacted metal pairs could lead to high friction factor and high wear rate. In addition, the presence of the metal surface oxidation film may change metal friction property. 3. Metal friction property not only relates to contact surface conditions, but also to the sliding speed, contact pressure, temperature, humidity, atmospheric environment and so on. Fig. 1 shows the universal rule in the variation of metal-to-metal frictional factor. The test schemes to analyze the friction coefficient There is no compression deformation on the surface of steel wire under load-free, and the line contact can be assumed. Coulomb friction law can be used in the test design. The following arrangement is used in this test, as shown in Fig. 2. All rights reserved. No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of Trans Tech Publications, www" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001255_icrms.2011.5979461-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001255_icrms.2011.5979461-Figure2-1.png", "caption": "Figure 2. wear between the thrust washer and the planet gear, bearing side.", "texts": [ " The lubrication oil flows from the center of the bearing to the two ends, most of which flows out through the gap between the planet gear and thrust washer. Only a little oil will flow out through the gap between the thrust washer and the planet carrier. Material composition of bearing is GCr15. The working environment of the supporting components for the planet gear is serious. The relative speed range of the planet gear is 0 ~ 5000r/min, and the radial load of the planet gear axle is in a range of 0~ 33000N [4]. After investigation, it can be found that there is a wear between the side of bearing outer ring and the thrust washer, as shown in Fig.2. There are two areas which have obvious wear, viz, one is the contact area between the planet gear and the thrust washer, the other is the contact area between the side of bearing outer ring and the thrust washer, which results into the second wear band. III. ANALYSIS OF TYPICAL WEAR FAILURE CASES According to a failure in the road test for power shift steering transmission, that is the damage to the left planet gear train. After the disassembly of the power shift steering transmission, it can be found that one planet gear bearing was burned seriously" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000265_epqu.2011.6128833-Figure5-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000265_epqu.2011.6128833-Figure5-1.png", "caption": "Figure 5. Simplified system with six-phase generator and a linear load.", "texts": [ " 4 illustrates the constituent parts of a full wind complex. The figure assumes the presence of a six-phase generator, in the terms argued in this work. The complex includes: the equivalent model of the wind incident in the wind turbine; a unit for rectification / inversion; a coupling transformer, and the point of coupling of the wind farm with the AC network connection. Although the recognition that the complete system has the indicated topology, for this paper analysis only the generator is taken into account. Due to this simplification, Fig. 5 illustrates the physical arrangement considered here, which shows the wind turbine, the special generator with permanent magnetic excitation, and a linear load being fed by the arrangement. The reasoning for the simplification is based on the fact that this paper is aimed at considering the generation behavior only. 2 p iF e IP i i n dk T F i d\u03b8 = iFdk d\u03b8 T e d T T J dt \u03c9\u2212 = [ ] [ ] [ ] e d V R I dt \u03bb = \u2212 \u22c5 \u2212 [ ]V eR The main objective of the computational studies is to show the following operational variables: voltages at the generator\u2019s terminals, currents supplying the loads, and magnetic flux in the machine" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000621_amm.284-287.854-Figure5-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000621_amm.284-287.854-Figure5-1.png", "caption": "Figure 5. Numerical model for journal bearing and measured points.", "texts": [ " Then, from the MFBD analysis, positions and velocities of journal and bearing are calculated. These data are transmitted to the EHD solver again. Fig. 4 shows the procedure of the fluid-structure interaction solving method between EHD and MFBD solvers. To implement the EHD module considering the bending stiffness effect with MFBD solver together, this study used the RecurDyn TM [15] MFBD environment. To validate the numerical results of this study, the experimental results of Okamoto et al. [16] are used. The detailed explanation about the numerical model is described well in [16]. Fig. 5 shows the numerical model and measured points of the oil film pressure. The rotational speed of shaft is 3250 rpm. Table 1 shows the simulation parameters used in the numerical model. Numerical results are compared with the experimental results of [16] at measured points to validate the model. Fig. 6 shows a pressure distribution which has a steep slope around the edge and a flat slope around the center. Generally, if the flexibility of the journal and bearing is not considered, the pressure distribution shows a parabolic shape along the depth direction" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000886_amr.317-319.764-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000886_amr.317-319.764-Figure2-1.png", "caption": "Fig. 2 Equivalent load model of the lifting systems during the forging process", "texts": [ " (2) Fy is proportional to \u2206y, so the work piece is equivalent to a spring whose stiffness can be expressed as 3 f 3 4l Ebh k = . The velocity of point P is half to that of the upper die as the deformation of the work piece is uniform [5,6], so 2 v v d p = . (3) dtv 2 v \u2206y t w d 0 \u222b \u2212= t . (4) Where vp, vd is the velocity of point P and the upper die respectively, and t0 is the time when the upper die touches the work piece. The load model of the lifting systems can be simplified, as shown in Fig. 2, based on the equivalences mentioned above. \u03b8r and \u03b8f, which are the angles between the lifting arms and the horizontal line, are treated as zero in this paper as they are always very small during the forging process. r rc ra m fc f fc fa m fcm w v l l l ll v l l l lll v \u22c5 + \u2212\u22c5 ++ = . (5) Where vw, vf, vr is the velocity of the work piece in the press point, the front lifting cylinder and the rear lifting cylinder respectively. Fy can be computed from equations (2), (4) and (5). From the theories of force balance and moment of force balance, the loads of the lifting systems can be derived by y fc fa m fcm f F l l l lll F \u22c5 ++ = " ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003440_icawst.2012.6469608-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003440_icawst.2012.6469608-Figure1-1.png", "caption": "Figure 1. Schematic drawing of fully-integrated electrochemical threeelectrode system composed of Ag/AgCl thin-film reference electrode, Pt thinfilm counter electrode and the pf-SWCNT-patterned working electrode.", "texts": [ " The results suggest that all the biomolecules listed above can be theoretically targeted to achieve some degree of immobilization efficiency. II. EXPERIMENTAL Plasma-patterned pf-SWCNT-based three-electrode microdevices were prepared on inexpensive glass substrates with low melting points. The SWCNT working electrode on the Pt support, Ag/AgCl reference electrode, and Pt thin-film counter electrode were integrated into the electrode system by conventional microfabrication and the plasma process presented in this work. Fig. 1 shows a schematic drawing of the miniaturized SWCNT-based electrode design. Pt conducting tracks are 0.25 mm wide and the working electrode is 1.1 mm in diameter with a CNT island patterned on a pre-patterned Pt thin film of 1 mm diameter. The dimensions of the reference electrode were 250 \u03bcm \u00d7 250 \u03bcm. The area of the counter electrode was approximately 3.358 mm2. The output pads were used to connect the electrodes on the substrate to a pin on a packaged chip or were interfaced to the external circuitry via a three-channel connector that was specifically designed to facilitate signal processing to the external circuitry", " The 80-nm thick SWCNT film was spraydeposited over the substrate using a previously prepared SWCNT suspension. The film thickness of the SWCNT was finely controlled by adjusting the flow rate of the SWCNT suspension while maintaining as constant all other spraying conditions, including the substrate temperature. Following SWCNT film deposition, the unwanted SWCNT layer deposited around the effective working electrode region was removed by 50-W O2 plasma etching at an oxygen flow rate of 10 sccm for 5 min. Figure 1a and 1b show SEM images of the fabricated SWCNT three-electrode device and magnified images of the edges of the patterned SWCNT working electrode and Ag/AgCl reference electrode, respectively. Figure 2. SEM image of (a) the fully-integrated electrochemical three-electrode system, and (b) the boundary between the pf-SWCNT working electrode and nitride passivation area. B. Immobilization of biomacromolecules Three different biorecognition molecules\u2014i.e., glucose oxidase (GOD), Legionella-specific antibody, and L" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000203_icmtma.2011.390-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000203_icmtma.2011.390-Figure1-1.png", "caption": "Figure 1. Rear multi-link independent suspension model", "texts": [ " In this paper, the research object is the rear multi-link suspension of one model, using multi-body dynamics and suspension kinematics theory, the influence of outer rubber bushing stiffness of the under control arm on the suspension performance were analyzed under the test of parallel wheel travel and loading brake force. II. REAR MULTI-LINK INDEPENDENT SUSPENSION MODEL The kinematic model of rear multi-link suspension is built according to the key hard point parameters of a car, as shown in the figure 1. Analyzing the variation characteristics of the suspension performance parameter under the simulation test of parallel wheel travel and loading brake force respectively while simulant bench exert a force on the wheel. In the analysis, axial, torsion, radical stiffness of outer rubber bushing of under control arm are increased to 5-fold and then compare the variation characteristics with the original. III. INFLUENCE OF OUTER RUBBER BUSHING STIFFNESS OF THE UNDER CONTROL ARM ON SUSPENSION Because of the rubber which is the composition of outer rubber bushing of under control arm, the static rigidity of rubber bushing is nonlinear" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001652_ecj.10392-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001652_ecj.10392-Figure1-1.png", "caption": "Fig. 1. Generation of a path.", "texts": [ " First, apparently L0, the locus of the fulcrum of the first link, can be constructed by connecting mutually 0-connective points from H0 to D0, because D0 is included in R(0). Now, let Ln\u22121 be a locus of the fulcrum of the n-th link consisting of a sequence of mutually (n \u2013 1)-connective points from Hn\u22121 to Dn\u22121, and let P, Q, and R be consecutive points on Ln\u22121. Then, because P and Q, and Q and R are (n \u2013 1)-connective, there exist pairs of feasible and adjacent attitudes [(P, S), (Q, T)] and [(Q, U), (R, V)] of the n-th link as shown in Fig. 1, and S, T and U, V are n-connective. Moreover, because A(Q, n) is a connected set in terms of n-connectivity, A(Q, n) includes a sequence of points that connects T to U, and consecutive points in the sequence are mutually n-connective. This means that {P, S}, {Q, T}, . . . , {Q, U}, {R, V} are a sequence of feasible and adjacent attitudes of the n-th link, and a sequence of n-connective points {S, T, . . . , U, V} constitute a locus of the movable end of the n-th link. By applying this process to every point in Ln\u22121, it is possible to constitute locus Ln of the movable end of the n-th link from Hn to Dn, which consists of n-connective points" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002063_ciima.2013.6682792-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002063_ciima.2013.6682792-Figure3-1.png", "caption": "Fig. 3. Ejemplo de los objetos a caracterizar. a)Llave b)Tuerca. En rojo los puntos de intere\u0301s", "texts": [ " Illave(x, y) = 255 para R > 160, G > 80 y 20 < B < 100 0 en cualquier otro caso (1) Ituerca(x, y) = 255 para 130 < R < 180 y 10 < G < 50 0 en cualquier otro caso (2) El resultado es una imagen binaria que contiene tanto los objetos segmentados como a\u0301reas de ruido producidas durante la segmentacio\u0301n (Fig. 2). De esta imagen se 4 obtienen los contornos, en los que pueden calcularse los momentos estad\u0131\u0301sticos [17]. Espec\u0131\u0301ficamente, el momento espacial m00 se refiere al a\u0301rea del contorno y lo utilizamos para realizar un filtrado del ruido dejando so\u0301lo el contorno de mayor a\u0301rea, es decir, el objeto que esta\u0301 siendo segmentado. Se definen los objetos a caracterizar como una llave y una tuerca (Fig. 3). Para cada punto en el contorno resultante, se calcula el a\u0301ngulo que genera con respecto a los puntos anterior y posterior a medida que se recorre el borde del objeto. Para esto, se calculan los vectores (~v1 y ~v2) que van desde el punto anterior (j \u2212 1) y posterior (j + 1) hasta el punto que se esta\u0301 procesando (j) (Fig. 4). En la Ecuacio\u0301n 3 se presenta la fo\u0301rmula para el ca\u0301lculo del a\u0301ngulo entre dos vectores, que se obtiene de la definicio\u0301n del producto escalar entre dos vectores. \u03b8 = arc cos ( ~v1 \u00b7 ~v2 \u2016~v1\u2016\u2016~v2\u2016 ) (3) En el caso de la llave, los a\u0301ngulos deben permitir hallar los puntos correspondientes a las salientes de la herramienta (Fig. 3a), mientras que en la tuerca se desea encontrar dos aristas de la parte inferior (Fig. 3b). La librer\u0131\u0301a utilizada para el procesamiento de ima\u0301genes se denomina OpenCV [18]. Entre las funciones de descripcio\u0301n de forma y ana\u0301lisis estructural se encuentra la funcio\u0301n minEnclosingCircle, la cual encuentra el c\u0131\u0301rculo de a\u0301rea m\u0131\u0301nima que encierra un contorno (Fig. 5). Con la informacio\u0301n del centro y el radio del c\u0131\u0301rculo, se limitan el nu\u0301mero de a\u0301ngulos encontrados en el punto anterior, restringiendo a\u0301reas que no son de intere\u0301s. Debido a que el robot NAO sostiene la llave con la mano derecha, se generan restricciones dadas por la movilidad del robot" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000937_1.4025818-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000937_1.4025818-Figure1-1.png", "caption": "Fig. 1 The kinematics chain of 6R-2P drill jumbo (home configuration)", "texts": [ "4025818] Keywords: mining drill jumbo, redundant manipulator, inverse kinematics, pose determination In mining industry, about 75% of the mineral materials are excavated underground using the drill and blast method [1]. Unfortunately, most drill jumbos in operation today are manually operated. In order to improve the operation efficiency, we need to develop an automation system for the drill jumbo. To this end, we have to solve the kinematics problem. This work focuses on one drill jumbo which is manufactured by MTI. Its kinematic chain is presented in Fig. 1. This drill jumbo has six revolute joints and two prismatic joints. Hence, it is actually an 8DOF redundant serial robotic manipulator, that has two more DOFs than required to achieve the desired position and orientation of the end-effector. Unlike most researchers who have addressed the inverse kinematics of the redundant manipulators at the velocity or the acceleration level [2\u20135], this work directly solves the inverse kinematics of the 6R-2P drill jumbo with two redundant DOFs at the displacement level. This work uses the method of successive screw displacements [6]. Hence, only two frames are needed as shown in Fig. 1. The reference frame Oxyz is attached to the drill jumbo at the first joint J1 with the Oxy plane lying in the horizontal plane and the z axis lying in the vertical direction. The mobile frame Quvw is attached to the end-effector. At the home configuration, the mobile frame is parallel to the reference frame. For every joint Ji, we define a screw $1 whose position soi and direction si in the home configuration of the drill jumbo are listed in Table 1. 2.1 Direct Kinematics. If all joint variables \u00f0h1; h2; d3; h4; h5; h6; h7; d8\u00de are given, the resultant transformation matrix Q can be obtained as Q \u00bc Q01Q12Q23Q34Q45Q56Q67Q78 (1) where Qk 1,k is the 4 4 homogeneous transformation matrix from screw $k to screw $k 1\u00f0k \u00bc 1; 2; :::; 8\u00de" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001002_amm.86.850-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001002_amm.86.850-Figure3-1.png", "caption": "Fig. 3. Movement and forces between the roller and the inner race", "texts": [ " The geometrical parameters of the slipping clutch include: (1) the gorge circle radius, R; (2) the angle between roller and race axes, \u03a3; (3) the axial displacement of rollers, ZC, and (4) the radius, length and number of rollers, r, b and Nr. All rights reserved. No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of Trans Tech Publications, www.ttp.net. (ID: 128.210.126.199, Purdue University Libraries, West Lafayette, USA-31/05/15,22:55:34) The resisting torque of SRSC is controlled by the applied axial force Fa (see Fig. 3). Suppose that the inner race, cage and rollers rotate relatively around their own axes with angular velocities of \u03c9i, \u03c9c, and \u03c9r, respectively, while the outer race is stationary. At the moment t, a point A on the inner race and a point B on a roller coincide with one another. After a time interval of \u2206t, point A moves to position A\u2019 and the roller to position B\u2019, therefore, the movement of the roller relative to the race is from A\u2019 to B\u2019. This movement can be further decomposed into two processes: (1) the rolling movement from A\u2019 to C\u2019; (2) the slipping movement from C\u2019 to B\u2019", " In coordinate system o-xyz, the differential force and moment in matrix form can be expressed as [ ] [ ] [ ] ( ) ( ) dtdz ztp ztp CdMdMdMdFdFdFdF a o iT zyxzyx \u222b == , , ,,,,, (5) where [C] is a coefficient matrix associated with positions of the contact points on roller and velocity parameters of roller and cage. The resultant forces and moments acting on the entire roller are calculated by integrating the Eq. 5 over the roller length b. If the mass of roller is neglected, the dynamic equilibrium equation can be expressed as [ ] [ ] [ ] [ ]0 ),( ),( 0 = == \u222b \u222b\u222b b a i b dtdz ztp ztp CdFF (6) Dynamic Equilibrium Equation of Races. As shown in Fig. 3, the applied axial force, Fa, is balanced by Fprj and Fr along the Z-axis. Therefore, the axial equilibrium equation of either inner or outer races can be given as ( ) 0),(sincoscoscossin )0()()()( =\u00b1\u22c5\u22c5\u03a3\u2212\u22c5\u03a3\u00b1=\u00b1 \u222b \u222b\u222b ai b a oziprjotiprjra b oZir FdtdzztpNFdFN \u03b3\u00b5\u03b3\u00b5\u03b3 \u2213 (7) Based on the FEA solver developed by the first author [3-4] , the normal contact pressures pi(o)(t, z) are calculated by simultaneously solving the deformation-pressure Eq. 1 and Eq. 4 and the dynamic equilibrium Eq. 6 and Eq. 7. Furthermore, the von Mises stress can be calculated" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003245_b978-0-08-096979-4.00004-9-Figure4.20-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003245_b978-0-08-096979-4.00004-9-Figure4.20-1.png", "caption": "FIGURE 4.20", "texts": [], "surrounding_texts": [ "The descriptions of the components and processes presented serve to illustrate that the functions and construction of the types of structure used in F1 motor racing do not lend themselves easily to simple analytical cases. While finite element analysis may be employed in panel design and \u2018failure indices\u2019 based on a chosen criterion obtained, the current state of practice is to regard these results, although valuable, as being only part of the process. The demonstration of the suitability of a design is still determined by the mechanical testing of key structural elements and, wherever possible, a complete structure as proof of its integrity. Real data measured while a car is The role of demonstration, concept and competition cars running at a circuit Impact proof loading may be captured and processed to emulate a realistic loading system for application to a captive test structure on an appropriate rig. This is always the best way of developing confidence in a design\u2019s capability. Advanced as FEA techniques have become, it is still not possible or practical to use them to predict failure quickly in some detailed areas of real structures, and particularly those of a sandwich construction. A rudimentary understanding of composite analysis is all that is required to appreciate the fact that the most likely site of an unexpected failure will be at a detail where there is a significant amount of loading applied across the low strength resin matrix or joint adhesive in a tensile sense. Despite this knowledge it is inevitable that, with the types of structures used in F1, it is very often the case that their shape will be dictated by aerodynamic considerations and by those of manufacturing practicality or expediency rather than best structural suitability. It is also inevitable, therefore, that areas of uncertainty will be built in to any given design. A range of different quality assurance techniques are employed but these will only highlight some of the problems." ] }, { "image_filename": "designv11_101_0000126_j.1747-1567.2011.00748.x-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000126_j.1747-1567.2011.00748.x-Figure1-1.png", "caption": "Figure 1 Force system in the involute spur gearing.", "texts": [ " The kinetostatic analysis of spur gears indicates that the normal force changes its value in the pitch point due to the action of friction forces, when the torsional moment on the gear is constant, so, the principal aim of this work is to demonstrate this change. Let us assume that the friction moments in bearings are much smaller than the moments caused by friction forces between teeth. Therefore, if we neglect the friction moments in the bearings of a gear pair, the force system in the gearing can be represented in the following form (Fig. 1). Here, the total force between teeth F has the normal component Fn and the tangential component, the friction force, Ff. In involute gearing the friction force changes its direction when the contact point K between teeth passes the pitch point P, causing excitation of noise and vibration in the gearing. Experimental Techniques (2011) \u00a9 2011, Society for Experimental Mechanics 1 Let us restrict this analysis by the one pair teeth zone where the major change of the forces takes place. When the load is being transmitted by only one pair of teeth and the point of contact K is in the approach action zone N1P, the torque on the pinion is determined by the formula T1 = F\u2032 nRb1 \u2212 (a \u2212 x)Ff, (1) and when the point K is in the recess action zone PN2, T1 = F\u2032\u2032 n Rb1 + (a + x)Ff, (2) where a = Rb1 tan \u03c6, F\u2032 n, F\u2032\u2032 n are the normal forces in the approach and recess action zones, respectively, Rb1 is the base radius of the pinion, and x is the distance between the point of contact K and the pitch point P" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000694_s1063776113110034-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000694_s1063776113110034-Figure1-1.png", "caption": "Fig. 1.", "texts": [ " It is equal to (9) The friction coefficient can be found using Eqs. (3) and (5). For simplicity, we assume that C1, 2(x) = C1, 2\u03b4(x). As a result, obtain the following expression for the friction coefficient: (10) (11) It is important in deriving Eqs. (10) and (11) that some positions of the fixed end of the rod correspond to the position of a free rod. One of this positions is realized in motion from left to right, and the other, in motion from right to left. Quite a different picture is realized at low pressures (see Fig. 1b), when the normal force does not exceed the instability threshold FE. In this case, the elastic forces acting in the contact at sym metric points of the roughness have equal and oppo sitely directed tangential components. In calculating the total force, we should average over possible posi tions of the fixed end of the rod. Such an averaging describes both the sum of the forces arising at different moments of the rod motion and the sum of the forces acting at the system of randomly arranged rods at rest. If such an averaging is performed in the situation shown in Fig. 1b, the resulting force is zero. Therefore, the friction force does not arise in the \u201cbrush\u201d model at low pressure. \u02dc \u02dc 4SEh 4Slh \u03c02I 1\u2013 .= kfr A p E \u239d \u23a0 \u239b \u239e 1/4p pc\u2013 p ,= pc \u03c02 12 E IRC1C2 Sl2 ,= A 2 35 \u03c010 \u239d \u23a0 \u239b \u239e 1/4 S E l9 C1C2RS \u239d \u23a0 \u239b \u239e 1/4 .= In this section, we consider contact of two rough surfaces similar to those considered in Section 2. The difference is that we here neglect the adhesion. We assume that there is no metastable states in each individual contact. The tangential forces fi, acting at the points of contact can be assumed to be random and to depend on the position of the contact: If the mutual influence of contacts is not taken into consideration, the mean tangential force ffr is where N the number of contacts" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002278_978-1-4471-4628-5_3-Figure3.113-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002278_978-1-4471-4628-5_3-Figure3.113-1.png", "caption": "Fig. 3.113 Bond graph model of inverting amplifier", "texts": [ " With the first assumption (rin \u2192 \u221e), current through the resistor will be zero, while with second assumption (rout \u2192 0), the voltage drop across the resistor will be zero. Thus power loss in both these resistors (effort times flow) is zero and both the resistors rin and rout can be dropped from the bond graph model. The reduced bond graph model is then shown in Fig. 3.111. fR:R As the name implies an inverting amplifier inverts and amplifies the input voltage. To achieve this, two external resistors are connected to op-amp as shown in Fig. 3.112. Figure 3.113 shows the bond graph model of the inverting amplifier. From the bond graph model, we can analyze for the gain of amplifier as follows: e6 = \u03bce5 = \u2212\u03bce4 = \u2212\u03bce9 = \u2212\u03bc (e7 \u2212 e8) = \u2212\u03bc (e6 \u2212 RF f8) (3.106) or e6 = \u03bcRF f8 1 + \u03bc (3.107) Again assuming f4 = f5 = 0, f8 = f9 = \u2212 f3 = \u2212 f2 or f8 = \u2212 (V1 \u2212 e3) R1 = \u2212 (V1 \u2212 e9) R1 (3.108) Thus e6 = \u2212 \u03bcRF R1 (1 + \u03bc) (V1 \u2212 e9) (3.109) or \u2212\u03bce9 = \u2212\u03bcRF (V1 \u2212 e9) R1 (1 + \u03bc) or e9 = RF V1 R1 ( 1 + \u03bc + RF R1 ) (3.110) From the above equations, e6 = \u2212\u03bce9 = \u2212 \u03bcRF V1 R1 ( 1 + \u03bc + RF R1 ) Thus gain of the inverting amplifier can be defined as k = e6 V1 = \u2212 \u03bcRF R1 ( 1 + \u03bc + RF R1 ) = \u2212 RF R1 ( 1 \u03bc + 1 + RF \u03bcR1 ) (3" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002066_isas.2011.5960957-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002066_isas.2011.5960957-Figure1-1.png", "caption": "Fig. 1. Mobile-hapto", "texts": [ " \ud835\udc47 + \u2202\u03a6 \u2202\ud835\udc99 \ud835\udc96 (13) = ?\u0304? \u2202\ud835\udc3b(?\u0304?) \u2202?\u0304? \ud835\udc47 + ?\u0304?. (14) Here, ?\u0304? and ?\u0304? are defined as follows: ?\u0304? = \u2202\u03a6 \u2202\ud835\udc99 \ud835\udc71 \u2202\u03a6 \u2202\ud835\udc99 \ud835\udc47 (15) ?\u0304? = \u2202\u03a6 \u2202\ud835\udc99 \ud835\udc96. (16) III. DYNAMICS In this section, controllers for constrained mobile-hapto are proposed. Hamiltonian of robot systems is given as follows: \ud835\udc3b = 1 2 \ud835\udc91\ud835\udc47\ud835\udc8e\u22121\ud835\udc91 (17) ?\u0302? = \ud835\udc3b + \ud835\udc5d0 (18) Here, \ud835\udc8e is a generalized mass matrix and momentum \ud835\udc91 are given as follows: \ud835\udc91 = \ud835\udc8e?\u0307? (19) Mobile-hapto realizes bilateral control between handle (local) robots and mobile (remote) robots [17]. Fig. 1 is an overview of mobile-hapto. \ud835\udc5e, \ud835\udc53,\ud835\udc5a are position, force and mass. Variables with subscripts 1 and 2 mean that with the master robot and the mobile robot, respectively. The master robot has 1 DOF and the slave has 2 DOF. In this research, we treated the slave as 1 DOF system modeled in Fig. 2. \ud835\udf03\ud835\udc5f, \ud835\udf03\ud835\udc59, \ud835\udf0f\ud835\udc5f, \ud835\udf0f\ud835\udc59 are the angler positions and torques of the slave actuators. Therefore, there are relations that \ud835\udc5e2 = \ud835\udc45 \ud835\udf03\ud835\udc5f+\ud835\udf03\ud835\udc59 2 , \ud835\udc532 = \ud835\udf0f\ud835\udc5f+\ud835\udf0f\ud835\udc59 \ud835\udc45 with the radius of wheels \ud835\udc45. The dynamics of the system is parameterized by our motion equation as follows: \ud835\udc99 = \u23a1 \u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a3 \ud835\udc61 \ud835\udc5e1 \ud835\udc5e2 \u2212\ud835\udc3b \ud835\udc5d1 \ud835\udc5d2 \u23a4 \u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a6 ,\ud835\udc96 = \u23a1 \u23a2\u23a2\u23a2\u23a2\u23a2\u23a2\u23a3 0 0 0 \u2212\ud835\udc531\ud835\udc5e1 \u2212 \ud835\udc532\ud835\udc5e2 \ud835\udc531 \ud835\udc532 \u23a4 \u23a5\u23a5\u23a5\u23a5\u23a5\u23a5\u23a6 (20) \ud835\udc8e = [ \ud835\udc5a1 0 0 \ud835\udc5a2 ] ,\ud835\udc71 = [ 0 \ud835\udc703 \u2212\ud835\udc703 0 ] (21) C" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003921_iros.2011.6048332-Figure9-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003921_iros.2011.6048332-Figure9-1.png", "caption": "Fig. 9 Phases of stairs-descending assist.", "texts": [ " The robot might be able to assist the user by moving automatically according to the prepared trajectory in order to realize natural human motion when the user descends stairs. However, it is difficult to recover the user\u2019s balance when the user loses own balance for some reason because the COG of the user might move outside the supporting leg. In the stairs-descending assist which is proposed in this paper, the COG of the user is moved to the next step after the user puts the swing leg on the next step to avoid these situations. The phases of the stairs-descending assist are shown in Fig. 9. In the first phase as shown in Fig. 9(a), the user moves own swing leg to the next step. Then the robot generates the additional torque for the perception-assist as necessary to prevent the COG and ZMP of the user from moving outside the supporting leg. When the robot wants to modify ZMP, the robot generates the additional torque at the ankle joint motor of the supporting leg. If ZMP moves near the boundary of the support polygon, the additional torque of the ankle joint motor is generated to move ZMP to the center of the support polygon. On the other hand, when the robot wants to modify COG, the robot generates the additional torques at the hip, knee and ankle joint motors of the supporting leg and tries to change the posture of the upper body region because the movement of the mass of the upper body is the most effective way to change COG. In the second phase as shown in Fig. 9(b), the robot tries to put the user\u2019s swing leg on the next step while the COG and ZMP of the user remain in the supporting leg. To realize this motion, the robot rotates the user\u2019s hip, knee and ankle joint of the supporting leg automatically until the swing leg is put on the next step. The robot tries to move the COG of the user down to the z-axis direction with having kept the COG of the x-axis direction in the support leg. In this proposed method, since the COG and ZMP of the user remain in the supporting leg during the motion, it is comparatively easy to keep the user\u2019s balance, and the robot can try to modify the user\u2019s balance by adding the additional torque to the supporting leg. The robot judges the end of first phase by the position of the swing leg based on the encoders. After that, the robot starts the perception-assist of the second phase when the user moves the swing leg down to the next step. The robot shifts to the final phase after the user\u2019s swing leg is put on the next step. In the final phase as shown in Fig. 9(c), the user tries to move the COG and ZMP to the next step, and move another leg to the next step. The supporting leg changes during this motion. The robot helps the user\u2019s motion based on ZMP in order to prevent the user from losing the balance. In the proposed stairs-descending assist, although the user\u2019s motion is not always a natural motion, the robot can assist the user\u2019s motion always because the COG and ZMP of the user remain in the support polygon, and the user can descend stairs safety while being supported by the robot", " 10 and 11, the subject\u2019s right leg entered the virtual wall at about 3 seconds. Then, the robot judged that the subject might stumble on the stairs, and generated the additional motion modification force along the virtual wall. The subject could ascend the stairs without stumbling on the stairs by the perception-assist of the robot. The experimental result when the subject descended the stairs is shown in Figs. 12 and 13. Each line expresses the same as Figs. 10 and 11. In Figs. 12 and 13, (a), (b), and (c) show each phase as shown in Fig. 9. In the first phase as shown in range (a), the subject moved own right leg to the next step in the direction of x axis. In the second phase as shown in the range (b), the robot rotated the hip, knee, and ankle joint of the subject\u2019s left leg (supporting leg) while keeping the posture of right leg (swing leg) in order to put the right leg on the next step. From Fig. 12, ZMP is the almost constant value and remains in the supporting leg in the range (b). After the second phase, the robot encouraged the subject to shift ZMP and the weight of the subject from the left leg to the right leg in final phase as shown in the range (c)" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001430_s12206-012-1266-x-Figure5-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001430_s12206-012-1266-x-Figure5-1.png", "caption": "Fig. 5. Stresses and displacements in case that the spacer without a gap was used.", "texts": [ " It is discovered that even when the wheel shaft was supported by the well preloaded bearings, it rotated slightly in plane, not out of plane, with respect to the carrier housing. The lock nut slid by 0.18 mm on the planet carrier face, as shown in Fig. 4(c). Second, it was assumed that there was no initial gap between the spacer and the inner race. The structural analysis result shows that both bearings were not preloaded and only the inner races were loaded as the lock nut tightening force was applied (Fig. 5(a)). Next, as the dual tire load was applied, the upper part of the left bearing was the most severely compressed and the lower part of the right bearing was compressed, as shown in Fig. 5(b). Compared to the first case, the compressed portion of the roller was smaller in the second case. The contact stresses on the lock nut were also roughly less than 1 MPa. The stresses were also too small to initiate yielding. Since the wheel shaft was supported by the not-preloaded bearings, it rotated severely in plane against the carrier housing. The lock nut slid by 0.44 mm on the carrier face, as shown in Fig. 5(c). Compared to the first case with an initial gap of 1.0 mm, the slippage increased to be 2.4 times larger. The slippage is seen to be related to the loosening. The loosening process in Fig. 6 is very similar to the present loosening failure. The lowered annular surfaces were observed with a lowresolution USB digital microscope. The height, i.e. the radial thickness, is 3.5 mm. Many circular scratches were easily observed. It seems as if they were stacked one after the other along the circumferential direction" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000026_amr.430-432.1524-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000026_amr.430-432.1524-Figure1-1.png", "caption": "Fig. 1 Gear pair dynamic model", "texts": [ " In early design stage, dynamics analysis parameters often can be got by using theoretical calculation, and the finite element is one of the major methods [2]. Through the modal analysis we can get the structure parts of frequency and vibration mode. For the structure design in dynamic loading conditions, frequency and vibration model is very important parameters, in the progress of spectrum analysis, harmonic response and transient analysis is needed [3]. A pair of gear pair dynamic model is as shown in figure 1[4]. Where the spiral angle of active wheel is \u03b2 , then the relationship of axial vibration and the transverse vibration is z=tan \u03b2 . Without considering the friction of the systems, six degrees of freedom py , pz , p\u03b8 , gy , gz , g\u03b8 . py , gy of the system are the transverse vibration displacement. pz , gz are the axial vibration displacement. p\u03b8 , g\u03b8 are angular displacement around the z axis rotation. All rights reserved. No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of Trans Tech Publications, www.ttp.net. (ID: 128.210.126.199, Purdue University Libraries, West Lafayette, USA-24/05/15,03:33:36) Without considering the load, system motion differential equations can be got by Figure 1. yygg zzgzggzgg yggyggygg yypp zppzppzpp yppyppypp RFI Fzkzcym Fykycym RFI Fykzczm Fykycym \u2212= \u2212=++ =++ \u2212= =++ \u2212=++ \u03b8 \u03b8 (1) where pyk , gyk are tangential support stiffness, pzk , gzk are axial support stiffness, pI , gI are inertia moment, pyc , gyc are normal damping, pzc , gzc are axial damping. )(cos)(cos gggpppmggppppmy RyRycRyRykF \u03b8\u03b8\u03b2\u03b8\u03b8\u03b2 +\u2212+++\u2212+= (2) )](tan)(tan[(sin )]tan()(tan[(sin ggpgppppmggggppppmz RyzRyzcRyzRyzkF \u03b8\u03b2\u03b8\u03b2\u03b2\u03b8\u03b8\u03b2\u03b2 \u2212+\u2212+\u2212+\u2212+\u2212+\u2212= . (3) where mk is normal mesh stiffness, mc is normal mesh damping" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000186_amm.312.307-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000186_amm.312.307-Figure1-1.png", "caption": "Fig. 1 Three-dimensional solid meshed model of gearbox casing", "texts": [ " In the meanwhile, we extract the main vibration modal to make the vibration coupling analysis, which provides the basis for reasonable structural design of gearbox casing and lays the foundation for the further study on the analysis of fatigue and reliability. This gearbox is a five gears MT (manual transmission) which has three axles. The casing of gearbox can be divided into top casing and bottom casing, which is made of casting Aluminum (ADC12). When three-dimensional solid model is built, with respect to calculation accuracy and calculation economy, original structure is simplified and modified reasonably. After mesh optimization, the established model has 79624 tetrahedral units and 150486 nodes, as shown in Fig.1. In the paper, we focus on the impact of internal motivation on casing, ignoring the effect of road surface irregularity on casing. Therefore, the bottom surface of gearbox casing is restrained. All rights reserved. No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of Trans Tech Publications, www.ttp.net. (ID: 130.88.90.140, The University of Manchester, Manchester, United Kingdom-08/04/15,14:39:59) The finite element static analysis The force on the casing comes mainly from load of each bearing hole, and analysis is performed at the first gear and reverse gear of the gearbox under the maximal static working condition" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000662_gt2011-45999-Figure4-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000662_gt2011-45999-Figure4-1.png", "caption": "Figure 4: Image of Modified Bearing Cage with Cutaway Diagram. The Quarter is shown for Scale.", "texts": [ "org/about-asme/terms-of-use The passive sensor was fabricated on a polyetheretherketone (PEEK) ring that was pressed into a modified bearing cage. A cavity for the sensing capacitor was milled into the material extending into the cage and a shallow hole drilled through and filled with conductive epoxy. The sensing capacitor was then bonded into the cavity. A channel was cut onto the top surface of the PEEK ring and a wire was pressed into it. Both ends of the wire were soldered directly to the sensing capacitor, creating the resonator. A photograph of the completed sensor as well as a cross- sectional diagram is provided in Figure 4. The transceiver was also fabricated out of a PEEK ring with a channel milled out for the inductor wire, much like the passive sensor. A 1 meter coaxial cable was soldered to the ends of the inductor so as to allow an interface to the telemeter from outside of the bearing housing. This ring was then mounted on the inside of the bearing with the inductor placed 3mm away from the inductor on the cage sensor. Figure 5 shows the finished components of the telemeter and how they fit inside of the housing" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000982_12.904997-Figure4-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000982_12.904997-Figure4-1.png", "caption": "Figure 4. Makeup of the double flank micro gear rolling tester", "texts": [ " The design concept of double flank rolling tester for micro gears is based on the criterion that the test micro gear is rotated by the master gear and the center distance is variable to ensure no radial backlash during rotation5. The power source of this tester is a servo motor that drives the master gear. To achieve desired requirements for engagement, the test micro gear is mounted on a translational stage such that one another degree of translation is allowed. A spring is used to pull the test micro gear against the master gear and its elongation could be adjusted by a screw. The variation of the center distance is recorded by the high resolution encoder on the translational stage for accuracy analysis. Fig. 4 illustrates the system makeup of the double flank micro gear rolling tester. The master gear employed for double flank gear rolling test is shown in Fig. 5. Specifications are as follows: 380 teeth, a module of 0.12mm and the consequent pitch diameter of 45.6mm. iF \u2032\u2032 and if \u2032\u2032 are 11\u00b5m and 3.7\u00b5m respectively, which corresponds to grade 5 and grade 7 if ISO 1328-2 is specified. Fig. 6 is the close view of the test micro gear in mesh with the master gear. Fig. 7 shows the experimental data of four continuous clockwise revolutions of the micro gear, where the vertical axis is the encoder reading of the translational stage while the horizontal axis is the index of sampling data" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003069_978-1-4419-9305-2_30-Figure5-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003069_978-1-4419-9305-2_30-Figure5-1.png", "caption": "Figure 5: Five degree of freedom bearing-pedestal model [1]", "texts": [ " Different mathematical models [7-9] have been developed to study the dynamic effects on the transmission error (TE) of the UNSW gearbox. These were lumped parameter models (LPM), which assume that each shaft mass and inertia is lumped at the bearings or at the gears. In all these models, rolling element bearings (REBs) were modelled as a single degree of freedom (mass-spring) system with constant stiffness. In [1, 2] Sawalhi and Randall combined the gear model with a bearing model, which has the capacity to model faults. This resulted in a 34 DOF model (Fig.4) . The 34 DOF in the LPM included a 5 DOF bearing model (Fig. 5). The translational degrees of freedom were considered both along the Line of Action (LoA) and perpendicular to it. The casing model considered was a simple one and contained only two modal frequencies from hundreds available. This still gave a valid simulation of the gearbox for the purpose of studying its behavior for a spalled bearing (envelope signal for demodulation of a high frequency resonance) and also in studying the different interactions that exist in the system by comparing simulations with real measurements, for a variety of localized and extended faults in both gears and bearings [1, 2]" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001235_amr.424-425.838-Figure9-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001235_amr.424-425.838-Figure9-1.png", "caption": "Fig. 9. The initial position of mould", "texts": [ " 0)2( 2 2 =++ p dy d G \u03bd \u03bb baffles are not effected by roller\u2019s gravity, so value of p is 0, so solution is BAy +=\u03bd RBA \u2200\u2282, (9) According to the elastic equation of lame formula y Gy \u2202 \u2202 += \u03bd \u03bb\u03b8\u03c3 2 Because MPayy 800][ 0 =\u2212 =\u03c3 , so put (9)into the elastic equation,the solution is G A y 2+ \u2212= \u03bb \u03c3 therefore By G y + + \u2212= 2\u03bb \u03c3 \u03bd (10) place 10 mm to boundary of mould does not have displacement, so 0)( 01.0 ==y\u03bd Put (10) into it 01.0)2( \u00d7+= GB y \u03bb \u03c3 Put the equation above into (10) 01.0)2( 2 \u00d7++ + \u2212= Gy G yy \u03bb \u03c3 \u03bb \u03c3 \u03bd solution is 51004.3 \u2212\u00d7=\u03bd m According to calculation results, corresponding correction in mold can be done. Mould initial state is shown in Fig. 9. Comparing cold rolling technology with traditional technology, the production efficiency is improved at least 10 times. if mass production, you can greatly reduce machine tool equipment,venue and labour. The advantages of mould in rolling process is high reliability, smooth movement, and good rolling effect, which makes it has good economic efficiency. thus cold rolling is of a future of vast technology, should be widely popularized and applied. Control [1]. Liping Jiang, Shougao Tang and Junmin Wang" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001322_amr.791-793.718-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001322_amr.791-793.718-Figure3-1.png", "caption": "Fig 3 Stress contours of rack and pinion Fig 4 Strain distribution of rack and pinion", "texts": [ " COSMOS/WORKS can be divided into a number of processors by functional role: includes a preprocessor, a solver, two post-processor and several auxiliary processor, such as design optimization; COSMOS/WORKS preprocessor is used COSMOS/WORKS solver is used to apply loads and boundary conditions, and then finish solving arithmetic; COSMOS/WORKS post-processor is used to obtain and examine the results to the model evaluation, and then make other interesting calculations. This paper can get under load model stress distribution and deformation in post-processor, and according to the actual needs of the model cut, observation model internal stress distribution. Using time - history postprocessor can watch model in different time periods or sub-step course on the results, commonly used in processing transient or dynamic analysis results. Rack and pinion stress contours are shown as follows, Fig. 3. Strain distribution of rack and pinion as follows, Fig. 4: Roadheader is likely to encounter some unforeseen circumstances in working, such as the roof off will be on turret enormous impact, generally may reach 65000N .When considering the limit state, using the finite element analysis results are as follows, Fig. 5: Strength analysis refers to parts the load under the action of damage which does not occur, to ensure safety work in the design of period of internal energy. As the actual work, the gear is the operation, while it can only consider under a certain moment's static stress, the pair of gears' freedom degree" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000759_000370260774614689-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000759_000370260774614689-Figure1-1.png", "caption": "FIG. 1. CALCULATING BOARD", "texts": [ " Processes Laboratory, Large Steam Turbine-Generator Department, General Electric Company, Schenectady, New York A spectrochemical calculating board'vesigned for two-step data evaluation has been modified to reduce data evaluation to a single operation without the use of special concentration scales for each analytical line pair investigated. The board is similiar in application to the calculator described by Dewey (1 ) , but is simpler to manipulate and utilizes the Seidel density function. A Seidel density scale has been inscribed on the left edge of the vertical cursor (B) in addition to the two cycle log scale provided on the right edge (Figure 1) ) . Using this scale, Seidel calibration curves can be drawn on the plotting surface directly from transmittance data. A family of calibration curves was constructed and projected onto narrow plastic strips. These cover a wide range of emulsion contrast variations in small increments. Two sliding metal blanks (1', T') at the bottom of the plotting surface carry, respectively, a two cycle log scale and one of the calibration scales. Windows were cut in these metal blanks so that an index line (F) drawn on the plotting surface can be aligned with values on either the log scale or the calibration scale", " Thus i t is often desirable to compensate for the fixed energy losses in the sample beam (or in effect, to expand the scale by attenuation of the reference beam). This has been accomplished through the use of wire mesh screens or slotted shutters. These methods are inconvenient since they either produce a \"step-by-step\" attenuation, have only a limited attenuation range, or the degree of attenuation can not be easily adjusted. A reference beam attenuator that will compensate the energy losses in the sample beam from 0 to 100% transmission in a continuous manner and can be adjusted to energy differences as small as 1 % is shown in Figure 1. This attenuator was constructed for use with the Perkin-Elmer Model 21 Spectrophotometer and its mounting is shown in Figure 2. The brass vane structure was constructed of a rectangular frame 1% in. x 2 in. into which the 4 mil thick brass vanes were positioned by slotting the upper and lower sections of the frame. The vanes are 5/16 in. wide and are spaced 5/16 in. apart so that when the frame makes an angle of more than 45' with the reference beam there is no transmission. At an angle of O 0 with the beam there is about 96% transmission" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001298_s0016-0032(59)90424-7-FigureI-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001298_s0016-0032(59)90424-7-FigureI-1.png", "caption": "FIG. I. The three-gear drive.", "texts": [], "surrounding_texts": [ "In the field of automatic machinery, it is frequently necessary for the output shaft to have an intermit tent motion, while the input shaft is turning uniformly. It is usually desirable that the curves representing the velocity and acceleration of the output shaft be smooth and free from discontinuities. The masses are then accelerated and decelerated gradually, and the mechanism is free from shock and impact.\nThe three-gear drive illustrated in Fig. 1 can be arranged to serve such a purpose. By suitable choice of dimensions, the output gear 3 can be given a short dwell period or even a reversal of motion during part of the cycle. Ordinary spur gears can be used if desired. Gear 1 is mounted off center at O,a. The gears are held in mesh by links A B and BOo~. The arrangement of parts thus gives a four-bar linkage with lengths of links a, b, c, and d. Let it first be assumed that these lengths are known. Link a or gear 1, is the driver which is turning at a constant angular velocity of col rad/sec. Equations will now be derived which will give the rotation e of driven gear 3 for an angular rotation ~'a of the driving link.\nWhen links a and b are in a straight line as shown in Fig. la, the angles of the resulting triangle are/3, v, and 6.\ns = ~ ( a + b + c + d ) .\nBy trigonometry 2\nCO8 = s(s - d)\n2 (a + b)c\ncos8 = 2cos ' ~ - 1\nLet\n(1)\n2 s ( s - d)\n(a + b)c\n(2)\n1. (3)\n1 Professor of Mechauical Engineering, Northwestern University, Evanston, I11. 2 See O. W. ESHBACH, \"Handbook of Engineering Fundamentals,\" New York, John\nWiley & Sons, Inc., second edition, 1952, pp. 2-74.\n464", "Dec., x959.] THE \"I~ttREE-GEAR DRIVE 465\nI", "By sine theorem\nc sin 7 = ~ sin/~ (4)\nf3 = 1 8 o \u00b0 - (-y + a) . (s)\nWith the links in the posit ions of Fig. l a , let the drive shaft at O., be removed, and let gear 1 be ro ta ted through angle ~. This produces rota t ions in gears 2 and 3 indicated by the heavy arcs and the designated angles.\nLet gears 1 and 2 now be a t t ached to member A B , and let the mechanism be so moved tha t the drive shaft can be reassembled at 0,, , . The parts are now in the locations shown in Fig. lb with the drive shaft: ro ta ted through angle yi. Angle ~ is increased to 61, and the inclination of link BO,a is changed by the amoun t /3 - (~1 + 9e). Gears 2 and 3 underwent some addit ional rotat ion.\nThe equat ion for rota t ion ~ of gear 3 follows directly.\n0 = u + 71 (6 )\ne 2 = a 2 + d 2 - 2 a d c o s 0 = b\"- + c ~ - 2 b c c o s 6 ~ <7) ()F\nb 2 + c \" - - a 2 - d e ad cos 51 = 2bc + bc cos 0 = K1 @ K,2 cos 0 (,R)\nwhere b 2 + c ' 2 - a'-' - d e\nK 1 = 2bc (9)\nad K 2 = -b~ (1{))\n6 s i n r = - s i n a i (11)\ne\n~2 = 180 -- ( ~ 1 + r) (12)\na sin ~ 1 ~--- - - sin 0 (13)\ne\n3& = 0 -4- So! -- r, (14)" ] }, { "image_filename": "designv11_101_0001806_peds.2013.6526987-Figure5-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001806_peds.2013.6526987-Figure5-1.png", "caption": "Fig. 5. Experiment of Load test on the equipment apparatus", "texts": [ " The magnetically levitated part must be cordless and the power source must involve eighteen Ni-MH AA size rechargeable batteries connected in series, equating to a normal voltage of DC 23.4V. The battery capacity is 1900mAh. The target air gap length at steady state levitation was 10.5mm. At steady state levitation, each hybrid magnet had a controlled current at 0.1A due to the balanced mass between the levitation part and the force of attraction. The total current of coils was 0.4A, but this did not include the power of the control circuit. III. EXPERIMENTAL RESULTS Figure 5 shows Experiment of Load test on the equipment apparatus. A Weight was water bottles. We added a weight to each upper surface of a levitated part as load Experiment result ware as follows. Figure 6 shows the digital oscilloscope wave at four point levitation. When the power was switched turn on, the coil A current was excited to the maximum value of 5.44A; the coil B current was then excited to the maximum value of 1.63A. After excitation of the maximum current, the levitated part was able to attain steady state levitation" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001047_pamm.201110013-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001047_pamm.201110013-Figure1-1.png", "caption": "Fig. 1 Time-optimal motion planning with dry friction and power constraints", "texts": [ " Similarly, any state-dependent dynamic constraint j at the joint velocities q\u0307, joint accelerations q\u0308 and actuator forces Q, as well as at the target velocities tE and target accelerations t\u0307E, can be written compactly as b\u0302 1j(s, s\u0307) \u2264 m\u0302j(s) s\u0308 \u2264 b\u0302 2j(s, s\u0307), which represent limits on the accelerations s\u0308 and velocities s\u0307 of the form b\u03021j(s, s\u0307) |m\u0302j(s)| \u2264 sgn [m\u0302j(s)] s\u0308 \u2264 b\u03022j(s, s\u0307) |m\u0302j(s)| . . . m\u0302j(s) 6= 0 (3) b\u03021j(s, s\u0307) \u2264 b\u03022j(s, s\u0307) . . . m\u0302j(s) = 0 . (4) The time-optimal function s(t) subject to equations of the form (3), (4) has been shown to be bang-bang [1], thus the main task of the solution algorithm is to search for the switching points Sk where the acceleration s\u0308 switches from maximal to minimal value in the s-s\u0307 plane (see examples fig. 1a and fig. 1b). The time-optimal motion algorithms have been applied up to now to problems with constraints that are at the most linear in the velocities. In this contribution, the method is extended to handle two new problems, namely (a) dry friction constraints of \u2217 Corresponding author: email francisco.geu@uni-due.de, phone +00 49 203 3792715 \u2217\u2217 email andres.kecskemethy@uni-due.de, phone +00 49 203 3793344, fax +00 49 203 3792494 \u2217\u2217\u2217 email alois.poettker@de.bucyrus.com, phone +00 49 231 9224350 c\u00a9 2011 Wiley-VCH Verlag GmbH & Co. KGaA, Weinheim the form |A t\u0307E(KE, tE)| \u2264 t\u0307 max E , where t\u0307E = [\u03c9E , vE ] T is the velocity twist of the end-effector containing its angular velocity \u03c9E and its linear velocity vE , as well as (b) power consumption constraints of the form |q\u0307T Q| \u2264 Pmax. The case of dry friction constraints can be illustrated by the so called waiter motion problem (fig. 1a), where a 6-degrees of freedom manipulator has to move a tablet carrying a number of glasses from an initial pose to a final pose as fast as possible such that objects placed on it do not slide, and the joint velocities and accelerations remain inside given limits. The no-slipping condition corresponds to the constraints \u221a [a\u0302kx] 2 + [a\u0302ky ] 2 \u2264 \u00b50a\u0302 k z , where a\u0302kz is the acceleration of each object k perpendicular to the tablet. These constraints are non linear in s\u0308 and thus incompatible with the known algorithms", " The case of overall power consumption constraints is shown by means of an example of the mining industry, where the shovel of an excavator with closed loop kinematics has to follow a given trajectory in minimal time without violating the actuator force maxima as well as the maximally allowed overall hydraulic power. These power constraints can be written as a function of s as \u2212Pmax m\u0302(s) s\u0307 \u2212 b\u0302(s, s\u0307) m\u0302 (s) \u2264 s\u0308 \u2264 Pmax m\u0302 (s)s\u0307 \u2212 b\u0302(s, s\u0307) m\u0302 (s) , (6) where m\u0302 = qJ T sM(s)qJs > 0. Both cases can be treated by the existing algorithms successfully. Figures fig. 1a and fig. 1b show the time-optimal solutions in the s\u2212 s\u0307 plane. This paper extends classical fixed-geometry time-optimal path planning methods to consider on the one hand sticking conditions and overall power consumption constraints, and on the other, smooth interpolation of angular acceleration as well as arbitrary multibody systems comprising multiple loops. The constraints can be straightforwardly constructed in terms of velocities, accelerations and forces at any place along the kinematical skeleton. The practical applicability of the introduced approach is illustrated by two applications from robotics and mining industry, [1] J" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003026_6.2012-855-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003026_6.2012-855-Figure1-1.png", "caption": "Figure 1. Views of the UAV with aerodynamic vectoring feature.", "texts": [ " Aerodynamic vectoring refers to the fact that the angle of incidence of the wing with respect of the fuselage is decoupled and dealt as an independent control variable. The aerodynamics is modeled over the whole maneuver range based on the wind-tunnel experimentation. The aircraft motion is studied in longitudinal mode. First, open-loop stability of the transition dynamics is studied using, a recently developed tool in nonlinear control, contraction theory. Then the nonlinear control synthesis is carried out using feedback linearization II. Aircraft Description and Optimal Trajectory Generation Figure 1 shows the UAV platform considered, which has a conventional wing-tail configuration with a tractortype propeller. The fuselage length as well as the wing span is 1 m. The aspect ratio of the wing is 4.31. Typical dimensional attributes of the model include a rectangular wing with the mean aerodynamic chord of 0.24 m and a T D ow nl oa de d by C O L U M B IA U N IV E R SI T Y o n O ct ob er 5 , 2 01 6 | h ttp :// ar c. ai aa .o rg | D O I: 1 0. 25 14 /6 .2 01 2- 85 5 American Institute of Aeronautics and Astronautics 3 propeller with a diameter of 0" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000105_978-3-642-20222-3_3-Figure3.38-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000105_978-3-642-20222-3_3-Figure3.38-1.png", "caption": "Fig. 3.38 Induction motor starting characteristics (relative values) for the medium power motor during the soft-start free acceleration in relation to delay angle \u03b1: a) starting current b) break torque c) starting torque d) idle run free-acceleration time", "texts": [ "36) since the value of the constant component of the motor torque increases during the initial stage of the start-up. The current waveform in the phase winding of the motor for such a supply is presented in Fig. 3.37. The delay angle in the range of around 40\u00b0 is virtually the sharpest one for which it is possible to conduct start-up of the motor during idle run within a sensible time, due to the considerable reduction of the value of electromagnetic torque of the motor. The approximate illustration of the effect of delay angle \u03b1 on characteristics of the motor is presented in Fig. 3.38. Soft-starters find application in drives with an easy start-up due to the considerable reduction of the torque following the fall of the value of the supply voltage. It is possible to control slip in an induction motor when electric power is delivered through the rotor windings to the external devices. This comes as a consequence of the fact that for a constant electromagnetic torque Te and constant supply frequency fs the power P\u03c8 delivered by the rotating field from the stator to the rotor has to remain constant" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000265_epqu.2011.6128833-Figure9-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000265_epqu.2011.6128833-Figure9-1.png", "caption": "Figure 9. Generator output current profile.", "texts": [ " The frequencies will naturally be changed with the imposed speed. If a real complex were to be simulated, these voltages would be applied to the double three-phase rectifier units; however, for the aims of this paper, a linear load has been fed. Fig. 8 is associated with the output voltage zoom at the region of 10 m/s wind speed. The first figure shows the six output voltage, and the second a comparison between phases a and x. The displacement angle of 30o can be seen. Thus, by supplying the linear load instead of the uncontrolled three phase rectifiers, Fig. 9 illustrates the line currents provided by the special generator. Once again, the profile behavior is similar to the voltage. This has occurred due to the fact the load has been modeled by constant parameters representation. Details associated to the six line currents around the 10 m/s wind speed region are given in Fig. 10. The magnetic fluxes associated with the individual six phases are shown in Fig. 8. The waveforms are clear enough to show that the displacement angle have been accomplished. V" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001172_ijmee.39.4.7-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001172_ijmee.39.4.7-Figure1-1.png", "caption": "Fig. 1 Force applied to a cylinder.", "texts": [ "nternational Journal of Mechanical Engineering Education, Volume 39, Number 4 (October 2011), \u00a9 Manchester University Press http://dx.doi.org/10.7227/IJMEE.39.4.7 Keywords rolling cylinder; frictional force; roll expectation It is no surprise that the force applied to the cylinder (e.g. a wheel) in Fig. 1 will cause it to move to the right and roll clockwise on the fl at surface [1]. Yet a seemingly similar situation is surprising. The following situation was described by an engineering instructor who interrupted his lecture to ask a question [2]. Imagine a rod rigidly attached to a cylinder as in Fig. 2. The rod extends beyond the surface, and a force is applied to the right at the end of the rod. Which way will the cylinder roll [3]? The instructor claimed that the cylinder could move only to the right" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003942_iros.2011.6048398-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003942_iros.2011.6048398-Figure2-1.png", "caption": "Fig. 2. Configuration q = (\u03b81, l1, \u03b82, l2, \u03b83, l3) for a concentric tube robot with N = 3 nested tubes.", "texts": [ " Problem Formulation We consider a concentric tube robot with N nested tubes numbered in order of decreasing cross-sectional radius. Each tube i consists of a straight transmission segment of length Li followed by a pre-curved portion of length Ci, i = 1, . . . ,N. The pre-curved portions of the tubes are curved with constant radii of curvature of ri, i = 1, . . . ,N. We assume that the device is inserted at a point xstart in 3D space and oriented along the vector vstart. Each tube contributes two degrees of freedom to the entire concentric tube robot, as shown in Fig. 2. Each tube may be (1) inserted or retracted from the previous tube, and (2) axially rotated. We therefore define a configuration of a concentric tube robot as a 2N dimensional vector q = (\u03b8i, li : i = 1, . . . ,N) where \u03b8i is the axial angle at the base of the i\u2019th tube and li > 0 is the length of insertion of tube i past the tube just before it (and l1 is the length of the first tube past the insertion point xstart). For a given configuration qi, we define the device\u2019s shape using x(q, s) : R2N \u00d7 R 7\u2192 R3" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000794_j.jcp.2012.04.020-Figure7-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000794_j.jcp.2012.04.020-Figure7-1.png", "caption": "Fig. 7. Plots of velocity vectors (a), pressure (b), x- and y-components of velocity (c) and (d) at t = 0.7T.", "texts": [ " During the recovery stroke, the tangential component of the boundary force is sufficiently large that the jump discontinuity and unboundedness of the normal derivatives of the velocity can be seen in Fig. 5(b) and (c). Note also that u and v are approximately zero near the floor (y = 0 along the solid lines), as expected. To illustrate the hydrodynamics effects generated by the cilium at different phases of its beat cycle, we show p and u everywhere in the 2D fluid computational domain [ 2,2] [ 2,2] during the effective stroke (t = 0.3T, Fig. 6) and during the recovery stroke (t = 0.7T, Fig. 7). The computational domain is discretized using 320 subintervals along each spatial domain. The configuration of the cilium is overlaid on the velocity plot (see Figs. 6 and 7, panels (a)). During the effective stroke, the cilium is approximately straight; thus, it exerts a relatively large effect on the fluid (compare the maximum pressure values in Fig. 6(b) and Fig. 7, and compare the size of the regions in Fig. 6(c) and Fig. 7 with near maximum velocity magnitude). During the recovery stroke, its effect on the surrounding fluid is reduced because the cilium has a more curved shape. Note also that, in both cases, fluid motion is larger near the tip of the cilium. Owing to the asymmetry in the effective and recovery strokes, a net propulsion effect is generated in the direction of the effective stroke. During one period of simulation, the wall remains approximately stationary. The maximum speed and displacement of the wall were computed to be 0" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000105_978-3-642-20222-3_3-Figure3.4-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000105_978-3-642-20222-3_3-Figure3.4-1.png", "caption": "Fig. 3.4 Diagram with cross-section of induction motor: a) 3-phase stator and rotor windings b) rotor squirrel-cage windings", "texts": [ " The currently solved tasks in drive control apply the following procedure: simple and functional control models in terms of calculations are accompanied with the correction of discrepancies resulting from parameter estimation using signals that are easily accessible by way of measurement. From the point of view of the current book the principal interest focuses on the mathematical models designed for determination of the characteristics of the drives and the ones applied for the purposes of control. The question of the number of the degrees of freedom (2.33) is encountered in systems with lumped parameters whose motion (dynamics) is described by a system of ordinary differential equations. For the case of an induction machine (Fig.3.4) this means one degree of freedom of the mechanical motion sm = 1, for variable \u03b8r denoting the angle of rotor position and the adequate number of the degrees of freedom se for electric circuits formed by the phase windings. For the case when both the stator and rotor have three phases and the windings are independent, in accordance with the illustration in Fig. 3.4a, the number of electric degrees of freedom is sel = 6. The assumption that electric circuits take the form of phase windings with electric charges Qi as state variables does not exclude the applicability of a field model for the calculation of magnetic fluxes \u03c8i linked with the particular windings of the motor. This possibility results from the decoupling of the magnetic and electric fields in the machine and the consideration of electric currents ii Qi = in the machine as sources of magnetic vector potential (2.180), (functions that are responsible for field generation). In this case we have to do with field-circuit models [48], in which the model with lumped parameters describing the dynamics of an electromechanical system (in this study the induction motor) is accompanied by an interactively produced model of the electromagnetic field in which the present flux linkages \u03c8i are defined. Hence, the model of an induction motor whose diagram is presented in Fig 3.4a has 7=+= em sss (3.1) degrees of freedom. In this place one can start to think about the state encountered in the windings of a squirrel cage motor (Fig. 3.4b), which does not contain a standard three-phase winding, but has a cage with m = Nr number of bars. The squirrel cage winding responds to the MMFs produced by stator winding current. The induced EMFs in squirrel cage windings display the same symmetry properties on condition that the squirrel cage of the rotor is symmetric in the range of angular span corresponding to a single pole of the stator winding or its total multiple. Hence, the resulting number of degrees of freedom ssq for a symmetrical squirrel cage winding [101] is expressed by the quotient us v u p m sq == 2 (3", "2) where: m - the number of bars in the symmetrical cage of a rotor u,v - relative prime integers The number of the degrees of freedom of the electric circuits of a rotor\u2019s squirrelcage winding ssq = u corresponds to the smallest natural number of the bars in a cage contained in a span of a single pole of the stator\u2019s winding or its multiple. This is done under the silent assumption that the stator\u2019s windings are symmetrical. If the symmetry is not actually the case, the maximum number of the degrees of freedom of a cage is equal to 1+= mssq (3.3) which corresponds to the number of independent electric circuits (meshes), in accordance with (2.195), in the cage of a rotor (Fig 3.4b). For the motor in Fig. 3.4b, we have p = 2, Nr = m = 22, hence the quotient: v u p m === 2 11 4 22 2 and, as a result, the number of electric degrees of freedom for a squirrel cage winding amounts to ssq = u = 11. This means that in this case the two pole pitches of the stator contain 11 complete slot scales or slot pitches of the rotor, after which the situation recurs. The large number of the degrees of freedom of the cage makes it possible to account in the mathematical model for the parasitic phenomena [80], for example parasitic synchronic torques", " For such an assumption of monoharmonic field the number of the degrees of freedom decreases to ssq = 2 regardless of the number of bars in the rotor\u2019s cage. In the studies of induction motor drives and its control the principle is to assume the planar and monoharmonic field in the air gap. Nevertheless, at the stage when we are starting to develop the mathematical models of induction machines, it is assumed for the slip ring and squirrel cage machines that the rotor\u2019s winding is three-phased (as in Fig. 3.4a) for the purposes of preserving a uniform course of reasoning. Hence, as indicated earlier, under the assumptions of a planar and monoharmonic field in the air gap, slip-ring and squirrel-cage motors are equivalent and can be described with a single mathematical model with the only difference that the winding of a squirrel-cage motor is not accessible from outside, in other words, the voltages supplying the phases of the rotor are always equal to zero. According to (2.189) and (2.210), Lagrange\u2019s function for a motor with three phase windings in the stator and rotor can take the form: \u2211\u222b = += 6 1 0 1 2 2 1 ~ ),0,0, ~ ,( k Q krkkr k QdQQJL \u2026\u2026 \u03b8\u03c8\u03b8 (3", " However, these tend to be less precise than the ones that result from field calculations since they account only for the major term of the energy of the magnetic field, i.e. the energy of the field in the machine\u2019s air gap. In the fundamental notion (Fig. 3.2), under the assumption of monoharmonic distribution of the field in the gap, the coefficients of mutual inductance take the form: )cos( lkkl MM \u03b1\u03b1 \u2212= (3.17) where: M - value of inductance coefficient for phase coincidence \u03b1k,\u03b1l - angles which determine the positions of the axes of windings k,l In accordance with Fig. 3.4 these angles are: pp pp rrr 3/2,3/2,,, 3/2,3/2,0,, 654 321 \u03c0\u03b8\u03c0\u03b8\u03b8\u03b1\u03b1\u03b1 \u03c0\u03c0\u03b1\u03b1\u03b1 \u2212+= \u2212= (3.18) - for the stator\u2019s windings and rotor\u2019s windings, respectively. The number of the pole pairs p reflects p-time recurrence of the system of windings and spatial image of the field along the circumference of the air gap. On the basis of relations in (3.15 \u2013 3.18), the matrix of the inductance coefficients of stator\u2019s windings takes the form \u23a5 \u23a5 \u23a5 \u23a6 \u23a4 \u23a2 \u23a2 \u23a2 \u23a3 \u23a1 \u2212\u2212 \u2212\u2212 \u2212\u2212 += 1 1 1 2 1 2 1 2 1 2 1 2 1 2 1 3 sssph ML 1M \u03c3 (3" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001430_s12206-012-1266-x-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001430_s12206-012-1266-x-Figure2-1.png", "caption": "Fig. 2. Wheel shaft assembly.", "texts": [ " The failed lock nut was as shown in Fig. 1(b). The annular region was lowered. The lowered region\u2019s radial thickness is 3.5 mm. This is just the radial thickness at the region with overlapping the planet carrier face. The lock nut appeared as if it were being collapsed by the excessive compressive contact stress. The contact stress was to be checked. CATIA structural analysis module was used as a finite element analysis tool. Some parts irrelevant to the problem were deleted from the assembly. Only six parts in Fig. 2 were modeled in the structural analysis. (1: wheel shaft, 2: bearing 32012, 3: bearing 33013, 4: housing carrier, 5: locking nut, 6: planet carrier) The wheel load from the tire is transmitted to the wheel shaft *Corresponding author. Tel.: +82 52 259 2141, Fax.: +82 52 259 1680 E-mail address: sjchu@ulsan.ac.kr \u2020 Recommended by Editor Jai Hak Park \u00a9 KSME & Springer 2013 through the hub housing. The wheel shaft is supported by the taper roller bearing set. The spacer ring, not shown in the Fig. 2, limits the minimum distance between the bearings. Thus the preloads on the bearings can be bounded to a preset level. Since it was anticipated that the preloads play important roles in the failure, the bearings were treated in more detail as composed of three separate parts: i.e. the inner race, the rollers and the outer race. Thus, eight contact bodies are connected, as in Table 1. Several types of connections are available in CATIA structural analysis such as the contact connection, fastened connection, bolt tightening connection and so on" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000430_omems.2012.6318778-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000430_omems.2012.6318778-Figure1-1.png", "caption": "Figure 1: Process flow : 1. Deposit ITO (yellow part) on a glass substrate, 2. Patterning the ITO and make electrode and addressing lines, 3-4. Spin coat SU-8 on the substrate at the thickness of 25\u03bcm and pattern SU-8.", "texts": [ " Here we suggest a new EWOD structure and a simple fabrication process that best facilitate the idea of self dosing and enhanced the VAR. Our structure has grid walls that have openings through which every cell is connected. These openings are also used as a good storage of the oil which is pushed by the EWOD, leading to a high VAR (>90%). Also the side walls of the grids are not hydrophilic but hydrophobic, resulting in a different mode of EWOD operation. The whole fabrication procedure is described in Figure 1. The cell size was 200 \u00b5m\u00d7200 \u00b5m with the height 65(\u00b12) \u00b5m and the channel gap was 10 \u00b5m. Driven by the oleophilic capillary forces through the channel openings, the colored oil was uniformly filled up by spreading into every cell once a certain amount was placed at the edge or the center of the structure. We used 9:1 oil mixture of chloronaphthalene and dodecane which was dyed with oil blue N pigment. Electrolyte was first filled up over the whole area, and 1\u00b5l oil droplet was deposited at the edge of the structure that consists of 20\u00d720 pixels of 200 \u00b5m\u00d7200 \u00b5m" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000105_978-3-642-20222-3_3-Figure3.3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000105_978-3-642-20222-3_3-Figure3.3-1.png", "caption": "Fig. 3.3 A frequent shape of a squirrel-cage winding of a rotor of induction motor", "texts": [ " With these rings and by adequate butting contact using brushes slipping over the rings it is possible to connect an external element to the windings in a rotor. This possibility is used in order to facilitate the start-up of a motor and in many cases also to control its rotational speed. The squirrel cage forms the other variety of an induction motor rotor\u2019s winding that is more common. It is most often made of cast bars made of aluminum or, more rarely of bars made from welded copper alloys placed in the slots. Such bars are clamped using rings on both sides of the rotor. In this way a cage is formed (Fig 3.3); hence, the name squirrel cage was coned. The cage formed in this way does not enable any external elements or supply sources to be connected. It does not have any definite number of phases, or more strictly speaking: each mesh in the network formed by two adjacent bars and connecting ring segments form a separate phase of the winding. Hence, a squirrel cage winding with m bars in a detailed analysis could be considered as a winding with m phases. Moreover, a squirrel cage winding does not have a defined number of pole pairs" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003263_b978-0-08-097016-5.00007-3-Figure7.15-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003263_b978-0-08-097016-5.00007-3-Figure7.15-1.png", "caption": "FIGURE 7.15 Enhanced transient tire model in top and side view showing carcass compliances and contact patch mass.", "texts": [ " The model to be discussed automatically accounts for the property that the lag in the response to wheel slip and load changes diminishes at higher levels of slip that in the previous section was realized by decreasing the relaxation length. This latter approach, however, appeared to possibly suffer from computational difficulties (at load variations). Also, combined slip was less easy to model. In developing the enhanced model, we should, however, try to maintain the nice feature of the relaxation length model to adequately handle the simulation at speeds near or equal to zero. Figure 7.15 depicts the structure of the enhanced transient model. The contact patch can deflect in circumferential and lateral direction with respect to the lower part of the wheel rim. Only translations are allowed to ensure that the slip angle seen by the contact patch at steady state is equal to that of the wheel plane. To enable straightforward computations, a mass point is thought to be attached to the contact patch. That mass point coincides with point S* the velocity of which constitutes the slip speed of the contact point" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000653_s10846-013-9993-5-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000653_s10846-013-9993-5-Figure3-1.png", "caption": "Fig. 3 Six degree of freedom of helicopter in body frame", "texts": [ " The Trex 600 ESP remote control (RC) helicopter is a model helicopter chosen as the platform for carrying out the data measurement on real flight due to its sufficient payload capacity, great manoeuvrability and low cost replacement parts. It is equipped with Bell-Hiller stabilizer bar and has a two-bladed rotor of 0.6 m radius. The dry weight is 3.3 kg and allows payload of 2 kg with operation time of about 15 min. Trex 600 with the necessary instrumentation equipment installed is shown in Fig. 2. 2.2 Modelling Structure For modelling a model-scale helicopter, a standard six degree of freedom (DOF) model is used. The axes of rotation of the helicopter are shown in Fig. 3. The standard equation describing angular velocity in the body frame is defined: \u03c9\u0307 B = I\u22121 (I\u03c9B \u00d7 \u03c9B)+ I\u22121\u03c4 B (1) \u03c9B = [ p q r ]T (2) \u03c4 B = [ M\u03c6 M\u03b8 M\u03c8 ] (3) I = \u239b \u239c \u239d Ixx 0 \u2212Ixz 0 Iyy 0 \u2212Ixz 0 Izz \u239e \u239f \u23a0 (4) where \u03c9B is the angular velocity in body fixed reference frame, \u03c4 B is the moment components along body axes, and I is the fuselage inertial matrix in body coordinates. Due to the symmetry of the helicopter with respect to the xB \u2212 zB plane, the terms Ixz and Iyz are zero. Although Ixz is non-zero, but the value is typically much smaller than the other terms, thus it will be ignored in the model", " The Flybar flapping angle derivation starts from defining dynamics of flybar in the Euler equation [18]: \u03c4 F = \u03c9F \u00d7 IF\u03c9F + IF \u03c9\u0307F (8) \u23a1 \u23a2 \u23a2\u23a2 \u23a3 \u03c4F1 n \u222b BR2 R1 rdL \u03c4F3 \u23a4 \u23a5 \u23a5\u23a5 \u23a6 = \u23a1 \u23a2 \u23a3 0 \u2212I f \u03c9F1\u03c9F3 I f \u03c9F1\u03c9F2 \u23a4 \u23a5 \u23a6+ \u23a1 \u23a2 \u23a3 0 I f \u03c9\u0307F2 I f \u03c9\u0307F3 \u23a4 \u23a5 \u23a6 (9) where I f is the unified rotational inertial of the flybar. The parameters \u03c4F1, \u03c4F2, \u03c4F3 are the external moment applied to flybar, which can be obtained by integrated lift elements dL along the length of flybar and \u03c9F1, \u03c9F2, \u03c9F3 are the angular velocities of flybar around iB, jB and kB axes in Fig. 3. After defining the flybar angular velocity, inertial and external forces, the flybar flapping angle is defined [18]: \u03b2 = \u03b2o ( \u03b11L7L8 L5L6L9 (\u2212\u03b4\u03b8 cos \u03be \u2212 \u03b4\u03c6 sin \u03be ) +\u03b12 ( \u03c6\u0307 cos \u03be + \u03b8\u0307 sin \u03be ) ) + 2 ( \u03c6\u0307 sin \u03be \u2212 \u03b8\u0307 cos \u03be ) , (10) \u03b2o = 1 2 n 2\u03c1ac2 2 ( B4 R4 2 \u2212 R4 1 ) 8I f (11) The constants a1, a2 are a correction factor to compensate for simplified flybar aerodynamics, \u03b4\u03c6, \u03b4\u03b8 are the roll and pitch input command respectively, \u03c6\u0307, \u03b8\u0307 are the roll rate and pitch rate of helicopter body frame respectively, \u03be is the orientation angle of flybar and L5,6,7,8,9 are the linkages in the rotor hub assembly" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000868_12.918709-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000868_12.918709-Figure1-1.png", "caption": "Figure 1. Hole patching in the WSAN.", "texts": [ " When nodes detect that multiple UGVs have selected the same destination, the nodes send a rejection message to some of the UGVs. Upon receiving the message, the UGVs will make adjustments in their path planning. Due to its distributed manner, there is a problem in distinguishing each destination group. Since we are considering problem domains such as in our coverage hole patching work3 where the destination is a hole in a sensor network composed of multiple hole-boundary nodes, nodes in the same set form a single destination; see Fig. 1. However, since we are working in a distributed environment, we cannot easily assign the set a unique identifier. Therefore, prior to the task allocation step, we propose using a timer-based leader election algorithm both for assigning a unique identifier to each set and for having each node identify their respective set. After the leader election, each destination set will have a single leader whose ID becomes the ID for the entire set. Deciding whether two nodes are in the same or different destination sets is, in general, application specific" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001740_9781782421702.12.749-Figure5-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001740_9781782421702.12.749-Figure5-1.png", "caption": "Figure 5 \u2013 The forces and torques at the cylinder (vane is in contact at V1)", "texts": [], "surrounding_texts": [ "To determine the angle of the cylinder, the triangles CS0S1, CV1R and RV0V1 shown in Figure 3 are used. By observing the triangles, the relationship between the various geometrical parameters according to Equation (1) can be derived. If the contact point is at the left hand side of the vane, the equation should be adjusted accordingly. Equation (1) can then be used to find the angle of the cylinder, \u03d5c, when the vane is in contact with the vane slot.\n( ) ( ) r v vCSScccrCSScc w\nwerrer \u03d5\u03b8\u03d5\u03d5\u03b8\u03d5 cos 2\n25.0cos2sinsin 222 1010 +\u2212+\u2212+=+ (1)\nDuring the operation, the vane extends and retracts in and out of the vane slot. To find the lengths of the vane at either sides of the vane that are exposed to the fluids inside the suction and discharge chambers, Equation (2) is needed.\nr\nv RVVRVV r w bb 2 sinsin 2010 == \u03b8\u03b8 (2)\nFrom the triangle RV1Vb1, the expression for the length of the exposed vane, lve, at its right hand side is obtained as expressed in Equation (3). If the contact point is at the left hand side of the vane, the equation should be adjusted accordingly.\n5.0 22\n2 2\n222 1 444 2 11\n\u23aa\u23ad\n\u23aa \u23ac\n\u23ab\n\u23aa\u23a9\n\u23aa \u23a8\n\u23a7\n\u239f \u239f \u239f \u23a0\n\u239e\n\u239c \u239c \u239c \u239d\n\u239b +\u239f\n\u239f \u23a0 \u239e \u239c \u239c \u239d \u239b \u2212\u239f \u239f \u23a0 \u239e \u239c \u239c \u239d \u239b \u2212\u2212+= vv RV v rRVrve ww l w rlrl (3)\nThe gap between the sides of the vane to the tips of the vane slot opening is necessary to check if the vane is in contact with the vane slot opening. These gaps may also cause the working fluid to leak from the high pressure chamber to the low pressure chamber. To help the understanding of the procedure, a geometrical model when the vane is not in contact with the vane slot is shown in Figure 4.\nThe vane is not in contact with the vane slot when neither the distance between R1 and S1 nor between R2 and S2 is equal to half of the vane width. The triangles CRS1 and R1RS1 are used and give the gap between the right hand side of the vane and the right hand tip of the vane slot opening as expressed in Equation (4). A similar approach can be performed to find the gap between the left hand side of the vane and the left hand tip of the vane slot opening.\n2111 v\nSRS w ld \u2212= (4)", "The torque balance equation of the cylinder is expressed in Equation (5).\n( ) ( )1 0 1 1 0 1 1 , ,\n1\ncos sinve c c v n c CV R V RV v f c CV R V RV c\nve\ndl dtI F r F r T dl dt \u03b1 \u03b8 \u03b8 \u03b8 \u03b8= \u2212 \u2212 \u2212 + \u2211 (5)\nWhere: vnvfv FF \u03b7,, = (6)\nThe \u03a3Tc term consists of all the torques at the cylinder which are not caused by the vane contact force nor the vane side friction. Rearranging and generalizing Equation (5) to be also applicable when the vane is in contact with the vane slot at V2 and using Equation (6) result in Equation (7). In this equation, the contact points V1 and V2 have been replaced by a general Vcont. If the vane contact point is at V1, then\n, , 1v n v nF F = . If the vane contact point is at V2, then , , 1v n v nF F = \u2212 , and if the vane\nis not in contact with the slot, there is no contact force and therefore , , 0v n v nF F = .\n0 0\n,\n, , ,\n, , ,\ncos sin cont cont cont cont\nc c c v n\nv n v n v nve c CV R V RV v c CV R V RV\nvev n v n v n\nI TF F F Fdl dtr r\ndl dtF F F\n\u03b1\n\u03b8 \u03b8 \u03b7 \u03b8 \u03b8\n\u2212 \u2211= \u239b \u239e \u239b \u239e\n\u2212 \u2212 \u2212\u239c \u239f \u239c \u239f\u239c \u239f \u239c \u239f \u239d \u23a0 \u239d \u23a0\n(7)\nThe forces and torques acting on the rotor are shown in Figure 6.", "Based on Figure 6, and assuming that the areas of the vane exposed to P1 and P2 are the same, the torque balance equation of the rotor when the vane is in contact with the vane slot at V1 can be expressed as shown in Equation (8).\n( ) 1 1 1 2 1 , , 12 2 ve ve v r r P P r v n RR v f r ve l dl dt wI F F r F l F T dl dt \u03b1 \u239b \u239e= \u2212 + \u2212 \u2212 + \u2211\u239c \u239f \u239d \u23a0\n(8)\nwhere: P ve chamF Pl l= (9) The \u03a3Tr term consists of all the torques acting at the rotor which are not caused by the vane contact force, the vane side friction nor the fluid pressures acting on the vane. Rearranging and generalizing Equation (8) to be also applicable when the vane contact is at V2 and using Equation (6) give Equation (10). Points R1 and R2 have been generalized as Rcont in this equation.\n( )2 1\n, 2\n2cont\nve r r P P r r\nv n ve v\nRR v ve\nlI F F r T F dl dt wl\ndl dt\n\u03b1\n\u03b7\n\u239b \u239e\u2212 + \u2212 + + \u2211\u239c \u239f \u239d \u23a0=\n+ (10)\nTo simplify the subsequent discussions, let us introduce the new parameters shown in Equations (11)-(14).\n0 0\n, , , , ,\n, , ,\ncos sin cont cont cont cont v n v n v nve v n c c CV R V RV v c CV R V RV\nvev n v n v n\nF F Fdl dtR r r dl dtF F F \u03b8 \u03b8 \u03b7 \u03b8 \u03b8 \u239b \u239e \u239b \u239e \u239c \u239f \u239c \u239f= \u2212 \u2212 \u2212 \u239c \u239f \u239c \u239f \u239d \u23a0 \u239d \u23a0\n(11)\nc cT T= \u2211 (12)\n, , 2cont\nve v v n r RR v\nve\ndl dt wR l dl dt \u03b7= + (13)\n1 2 1 22 2 ve ve r P r P r r l lT F r F r T\u239b \u239e \u239b \u239e= \u2212 + + + + \u2211\u239c \u239f \u239c \u239f \u239d \u23a0 \u239d \u23a0\n(14)\nEquations (7) and (10) calculate the contact force at the cylinder and the rotor, respectively. They are action and reaction forces. These forces are zero when the vane is not in contact with the vane slot. Substituting Equations (11)-(14) into Equations (7) and (10) and equating them give Equation (15).\n, , , ,\nc c c r r r\nv n c v n r\nI T T I R R \u03b1 \u03b1\u2212 \u2212 = (15)\nThe conservation of energy equation before and after the vane hits the vane slot is expressed in Equation (16) with timp is the impact time between the vane and the vane slot. This parameter should be obtained experimentally. If necessary, coefficient of restitution can be introduced in Equation (16) by adding a term into Tr and Tc in Equations (12) and (14).\n( ) ( )2 2 2 2 1 1 1 2 1 2 2 2 1 1 1 1 1 1 2 2 2 2 2 2r r c c r r imp r c c imp c r r c cI I t T t T I I\u03c9 \u03c9 \u03c9 \u03c9 \u03c9 \u03c9 \u03c9 \u03c9+ + + + + = + (16)\nwhere: 2 1r r r impt\u03c9 \u03c9 \u03b1= +" ] }, { "image_filename": "designv11_101_0002428_s10483-012-1604-x-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002428_s10483-012-1604-x-Figure1-1.png", "caption": "Fig. 1 Principal axis coordinate system of cross section and Frenet coordinate system", "texts": [ " (iv) Its cross section is perpendicular to the center line, and the shear deformation caused by the bending is ignored. (v) The tensile strain of the center line is ignored. (vi) The cross section can twist around the center line, and the angle is the continuous function of the arc length. To describe the deflection of the curved bar under the action of the external load, two coordinate systems are introduced: the principal axis coordinate system P -123 and the Frenet coordinate system P -NBT [5\u20136], as shown in Fig. 1. The axis T directs along the tangent direction of the center curve C. The coordinate system P -123 can turn around the center line C with the deformation of the curved bar. The coordinate system P -NBT describes the space shape of the center line. The inertial frame is named as O-xyz. The variable representing the arc length of the curve C is s, and the twist angle of the cross section is represented as the variable \u03d1(s). Denote the three components of the twisting vector with respect to the principal axes 1, 2, 3 by \u03c91, \u03c92, and \u03c93, respectively" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002830_6.2012-1918-Figure11-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002830_6.2012-1918-Figure11-1.png", "caption": "Figure 11. Cross-sectional model", "texts": [ " Configuration of cross-section A-A after buckling In this section, a simple analytical model is formulated to determine the dominant parameters and to identify the conditions for the local buckling based on the mechanical properties obtained by the finite element simulations. The results of the finite element simulations indicate that the local buckling is induced around the crease. Also, the layer thickness around the crease is sufficiently larger than that of the other area as shown in Fig.9 and 10. Thus, the crease region is modeled as the essential part of the wrapping fold membrane. Figure11a indicates the cross-section around the crease, where a and t represent the layer pitch of creased membrane and the membrane thickness, respectively. To simplify the analysis, the cross-section is modeled as a cylindrical configuration as shown in Fig.11b. The results of the finite element simulations also indicate that the local buckling is not induced when the membrane contacts the center hub, but it is induced when the membrane is wrapped around the center hub to some extent. This process translates to that the membrane is creased and the layer pitch becomes a by the normal force from the center hub after the membrane contacts the center hub, and then, the local buckling is induced. Thus, the analysis is performed based on the following two processes: 6 of 15 American Institute of Aeronautics and Astronautics D ow nl oa de d by C A R L E T O N U N IV E R SI T Y L IB R A R Y o n N ov em be r 28 , 2 01 4 | h ttp :// ar c" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002278_978-1-4471-4628-5_3-Figure3.4-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002278_978-1-4471-4628-5_3-Figure3.4-1.png", "caption": "Fig. 3.4 Schematic diagram of four-bar mechanism", "texts": [ " The resulting bond graph model is shown in Fig. 3.3. Note that a detailed model of the system can be developed from multibody approach with the model of the revolute joints as given in [2] (see the Rapson slide and Andrew\u2019s or seven-body mechanism models given therein) and then the differential causality problem can be avoided. However, such a model turns out to be very complex and the number of states in the model increase. On the other hand, one need not perform complex kinematic analysis to create those models. Four-Bar Mechanism Figure 3.4 represents a typical four-bar mechanism driven by a flexible shaft at constant speed. It consists of four links, the first link (ground link) being fixed. An elastic load is attached to output link 4. Figure 3.5 represents the bond graph model of four-bar mechanism. This can be drawn similar to slider crank mechanism. In the bond graph input crank rotation is represented by 1\u03b8\u03072 junction and output link rotation is represented by 1\u03b8\u03074 junction. The velocity source SF drives the elastic crank or link 2" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002126_icma.2011.5985978-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002126_icma.2011.5985978-Figure2-1.png", "caption": "Fig. 2. The origin and the coordinate axis of a robot", "texts": [ " Therefore, the laser scanner, Hokuyo UBG-04LX-F01 (UBG) is selected in this paper. The main specifications of UBG are listed in Table I. UBG has a small size and light weight of 260g. UBG can measure the distances up to 5600mm with 1mm resolution and 10mm tolerance. The scan frequency and the angular resolution of UBG is 36Hz and 0.352\u25e6, respectively. UBG can measure the intensity of the returned laser beam, which is used to determine the traffic lane on the road. Note that the detailed characteristics of UBG is provided in [6]. The coordinate axis of a robot is shown in Fig. 2. The origin O is located at the center of both feet with the xaxis going forward, the y-axis going left, and the z-axis going up. Pi j=(xi j,yi j,zi j) is the j-th measured point by ith scan, the point S=(0,0,h+l) is the place where a laser scanner is installed, di j is the distance between S and Pi j. When the mounting position of the laser scanner is high, the incidence angle to a target is increased and the accuracy of the measured distance is increased, which is the known characteristic of the laser scanner" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001141_amr.383-390.4096-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001141_amr.383-390.4096-Figure2-1.png", "caption": "Fig 2. 24/16 pole DSEG", "texts": [ " Differing from large-scale grid-connected wind power system, working state of a stand alone wind power system is mainly determined by wind speed, load and battery status. Generally speaking, there are three different operating modes .Generator provides power to the load and the battery, the battery is in the charging mode. Generator and battery both provide power to the load. the battery is in the discharge mode. Generator provides power to the load alone\uff0cthe battery doesn\u2019t work. Doubly Salient Electro-Magnetic Generator (DSEG) is a new type of brushless DC generator. Fig. 2 illustrates a 24/16 pole DSEG. Its stator and rotor both adopts salient structure with no windings mounted on the rotor. All the excitation windings and armature windings are mounted on the stator. Armature windings that are spatially opposite are connected in series to form a phase [4-5]. According to literature [4], the phase voltage of DSEG could be described as following equation: ( ) pf p f p p f p pf p p p dL dL di di u i i L L i r d d dt dt \u03c9 \u03c9 \u03b8 \u03b8 = \u2212 + \u2212 \u2212 \u2212 . (1) where, p denotes phase A, B or C, ip is the phase current, rp is the equivalent resistor of armature winding, if is the excitation current, \u03c9 is the rotor speed, Lp is the phase self-inductance, and Lpf is the mutual-inductance between armature winding and the excitation winding" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001539_educon.2011.5773230-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001539_educon.2011.5773230-Figure2-1.png", "caption": "Figure 2. Parallel-plate capacitor with suspension.", "texts": [ " Accelerometers measure acceleration, which is the rate of change of velocity. In case of crash, sudden change in velocity takes place, which results in an increased acceleration that will exceed a certain threshold and then the crash is detected by the system. For the crash detection systems, the precision of the sensor is crucial in saving lives. The \u2018Circuit Model\u2019 is explained next. The capacitor between conductors is a function of conductor size A, spacing z, and the permeability of the material between the conductors. Fig. 2 shows a simple parallel-plate capacitor suspended with flexures permitting the structure to move in the z direction. A more depth explanation with images and equations to better understand the concept is provided in the real example. Fig. 3 shows a simple electrical model of the interface circuit for the crash detection sensor. The sense capacitor is placed in series with a reference capacitor Cref. Several parasitic capacitors CPi are also used. In the real example, the output voltage of the circuit Vo is calculated, which represents the sensed acceleration" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001526_powereng.2013.6635832-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001526_powereng.2013.6635832-Figure1-1.png", "caption": "Figure 1. Flux deviation", "texts": [ " DTC PRINCIPLES For sensorless application, the \"voltage model\" is usually employed for the stator flux estimation: ( )dtiRvt ssss \u2212=)(\u03c6 (1) where vs is the stator applied voltage, is and Rs are the stator current and winding resistance respectively. Neglecting the resistive drop and considering time intervals Te sufficiently short, the vector that represents the stator flux variation has the same direction of the applied stator voltage space vector. By supposing that the vector \u03c6s is in the position shown in Fig. 1, it can be seen that to increase the module of \u03c6s can be applied the small vector v3 for example, in the case of two-level inverter. 978-1-4673-6392-1/13/$31.00 \u00a92013 IEEE POWERENG 2013 Further to the calculation of the components of the flux, the estimated torque is determined from the following equation. ( )qsdsdsqse iip \u03c6\u03c6 \u2212=\u0393 (2) As shown on Fig. 2, DTC strategy only needs the output voltages and currents of the inverter which feeds the induction machine, the instantaneous values of flux and torque in the machine are then calculated and the error can be gotten after compared with the referring values ( eref and \u03c6sref)" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000562_amm.201-202.517-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000562_amm.201-202.517-Figure3-1.png", "caption": "Fig. 3 Coordinates applied for tooth contact analysis", "texts": [ " The envelope to the family of grinding disk and the meshing equations between grinding disk and the machined face gear determine the surface of face gear, the family represented in S2 by transforming grinding disk coordinates in Sts to S2 Fig.2(b), and the meshing equations are the dot products of grinding disk normal vector and relative velocity between the points of grinding disk surface tangency to the face gear surface. TCA that provides under-load contact information must be simulated prior to LTCA. Fig.3 (a) and (b) respectively represent installation process of the pinion and face gear in coordinate system Sf for TCA simulation. Where \u03b3f=\u03c0-\u03b3m+\u2206\u03b3 and parameters \u2206\u03b3, \u2206q, and \u2206E are errors of alignment. The surfaces of the pinion and face gear will be in tangency contact in Sf if the following vectors equations are observed: 2f cs gs g g 2 1f cp gp 1( , , , , ) ( , , )s su L u L\u03c8 \u03d5 \u03a6 \u03a6=R R (3) 2f cs gs gs gs 2 1f cp gp 1( , , , , ) ( , , )u L u L\u03c8 \u03d5 \u03a6 \u03a6=n n (4) g2(1) 1 cs gs gs gs gs( , , ) 0f u L v\u03c8 = \u22c5 =n (5) g2(2) 2 cs gs gs gs gs gs( , , , ) 0f u L v\u03c8 \u03d5 = \u22c5 =n (6) Where R2f, n2f, R1f, and n1f are respective position and normal vectors of the pinion and face gear in Sf, f1 and f2 are respective meshing equations in the grinding disk feeding and face gear generating motion direction, \u03c6gs=\u03c6s+\u2206\u03c6s, \u03c62+\u2206\u03c62=m2s\u22c5\u03c6gs, m2s is the contact ratio" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003406_tmee.2011.6199359-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003406_tmee.2011.6199359-Figure1-1.png", "caption": "Figure 1. Vehicle dynamic model", "texts": [ " Therefore, a SMO for vehicle state estimation based on 3-DOF nonlinear vehicle model, rather than the commonly-used 2-DOF bicycle model, is developed here. This paper is organized as follows. Section II describes the nonlinear vehicle dynamic model. Section III presents the synthesis of SMO for estimating the land vehicle dynamics in detail. Simulation results are provided in Section IV. Section V is devoted to the conclusion. II. VEHICLE MODEL To the land vehicles, the dynamic model used in this paper is shown in Fig.1, which has 3-DOF for longitudinal motion, lateral motion and yaw motion of the vehicle, respectively. Longitudinal velocity, lateral velocity and yaw rate are defined in the body frame (i.e., b-frame) with the origin at the vehicle center of gravity (COG). According to Newton-Euler equations, the vehicle dynamic model can be described in the b-frame as [9] 2( ) cos sin sin cos ( ) sin cos x z y xr xf yf D x z z xf yf yr y z x yr xf yf m v v F F F k v I aF aF bF m v v F F F \u03c9 \u03b4 \u03b4 \u03c9 \u03b4 \u03b4 \u03c9 \u03b4 \u03b4 \u2212 = + \u2212 \u2212 = + \u2212 + = + + (1) where xv , yv , and z\u03c9 are the longitudinal velocity, the lateral velocity, and the yaw rate, respectively", " Substituting (2)-(3) into (1), and considering that is commonly a small angle, we can obtain 2 2 2 ( ) ( ) ( ) x L x y z y z f L x D x x b mv F f mg k v mv a b v a a C f mg k v k v v a b \u03c9 \u03c9 \u03b4 = \u2212 \u2212 + + + + \u2212 \u2212 \u2212 + (4) 2( ( )) ( ) ( ) z z L x y z y z f r x x b I a F f mg k v a b v a v b aC bC v v \u03c9 \u03b4 \u03c9 \u03c9 \u03b4 = \u2212 \u2212 + + \u2212 \u2212 \u2212 + (5) 2( ( )) ( ) ( ) y z y L x r x y z f x z x v bb mv F f mg k v C a b v v a C mv v \u03c9 \u03b4 \u03c9 \u03b4 \u03c9 \u2212 = \u2212 \u2212 \u2212 + + \u2212 \u2212 \u2212 (6) Note that the vehicle dynamics described by (4)-(6) is a 3-DOF nonlinear model. It is different from the commonlyused bicycle model. In the bicycle model, the longitudinal velocity is assumed to be constant, and thus the vehicle dynamics is further simplified as a 2-DOF linear model, which only includes the lateral velocity and the yaw rate. Moreover, in terms of kinematic relationships as shown in Fig. 1, we can get 2 2 W RL x z W RR x z T V v T V v \u03c9 \u03c9 = \u2212 = + (7) where VRL and VRR represent the left and right wheel velocities at the non-steering rear axle, respectively. TW is the rear width of the vehicle. In practice, VRL and VRR can be obtained using two wheel-speed sensors that are normally mounted at the vehicle rear axle, as applied in this paper. Therefore, for the SMO discussed later, xv and z\u03c9 are regarded as measurable vehicle states. III. SLIDING MODE OBSERVER DESIGN Compared to the conventional observers, sliding mode observer has been proven to be an effective method for handling uncertain systems with disturbances and modeling uncertainties [7-8]" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000300_kem.572.269-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000300_kem.572.269-Figure1-1.png", "caption": "Figure 1. (a): CAD model of an exhaust manifold. (b): Illustration of tool half, bronze & taps within in the crucible (c): Manufactured tool insert assembly.", "texts": [ " However, the implications of this for the part are significant: this split would necessitate in a join at a critical point in the part geometry and consequently hamper assembly of the finished part. Instead, for this application the tool was split into three sections, allowing the 300mm middle section to remain intact. For the example described in this paper the two outer cylindrical sections (enjoying a comparatively simple geometry) were instead produced using a traditional CNC process. Manufacture of Green Part. Production of the Green Part requires the fabrication of the two insert halves as shown in Figure 1 (a), either in separate builds or within the same build chamber. Whilst there are some financial advantages in terms of machine utilization and reduction in set-up operations where a single build is conducted, the use of mutli-builds can be considered as a means of mitigating risk for these parts which are in excess of some of the recommended parameters in terms of physical size and weight. For this example, the total sintering time for a tool half was noted as 24 hours. After the part was allowed to cool to the room temperature, the part was carefully cleaned by raising the part bed piston in a regular interval of 20 mm and removing the surrounding powder", " The inability to prepare the bottom of the part resulted in subsequent difficulties in the normal process of lapping, necessitating several attempts to achieved the desired result. Infiltration of the tool halves. After the part was cleaned and lapped, it was placed in a crucible together with surrounding taps. In this case it is very important to make sure that the taps are in firm contact with the side wall of the part. It was also necessary to fill any potential gaps with the ST- bronze cubes was calculated and evenly placed on the top of the taps as shown in Figure 1 (b). Alumina powder was used to cover on the top to avoid the direct radiation from the oven to the tool surface. After the crucible was placed inside the oven, it underwent the infiltration cycle according to the time and temperature settings given by 3D Systems [6]. After a 24 hour oven cycle, the part was taken out, and the same procedure was followed for the manufacture of the second tool half. After the oven cycles, the taps (which were found to remain attached to the side wall) were sawn off and the sides of the both halves machined to size", " The weight of each tool half was measured and the infiltration efficiency was calculated by assessing the weight of the part. The achievement of 98% efficiency together with the visual inspection of not being able to trace any leftover of bronze infiltrate around the tool confirmed that both halves were fully infiltrated. Table 1 shows the building time and the material usage for the complete tool. After that both halves were assembled with pre-machined tool steel blocks to achieve the specified size as shown in Figure 1 (c). Finally, the tool finishing operation and dimensional inspection based on 3D laser scanning and Coordinate Measurement Machine (CMM) were implemented. Since this process involved the prolonged thermal exposure in the fabrication of a relatively large tool, visual and physical inspection checks were performed, including: (i) Assembly conformance tests to check the mismatch between contact surfaces, and (ii) Dimensional inspection based on 3D laser scanning and CMM to generate error maps, comparing the actual surface versus original CAD model, for the tool evaluation and analysis" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002761_s12204-013-1363-8-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002761_s12204-013-1363-8-Figure1-1.png", "caption": "Fig. 1 Schematic of an active heave compensation system", "texts": [ " [10] and a discontinuous projection-type adaptation law, where the latter one has been widely implemented and tested in the motion control of hydraulic systems[11-15]. By implementing the observation results and using the back-stepping technique, a controller is then designed to be robust with respect to the unknown force acting on the rod of the hydraulic system. The effectiveness of the proposed control strategy is proved theoretically and validated by means of simulations. Consider an active heave compensation system illustrated in Fig. 1. This system is composed of a doublerod hydraulic cylinder and a four-way servo valve. The cylinder is fixed to the ship. A rigid cable connects the payload and lower rod tip via a ball joint. Hence, we can assume that only the forces on the vertical direction need to be considered. The displacement of the piston with respect to the ship is denoted by xh; L is the cable length and is assumed to be constant; u is the control current input of the servo valve. The reference water level is a predefined horizontal level fixed to the earth", " Then, to simplify the control design procedure, the following practical assumption for the uncertain parameters is made. Assumption 1 The unknown parameters \u03b8i and the disturbance f\u0304(t) satisfy \u03b8i \u2208 \u03a9\u03b8i \u0394= [\u03b8i min, \u03b8i max], i \u2208 {1, 2, \u00b7 \u00b7 \u00b7 , 5}, and |f\u0304(t)| \u03c3f , where \u03b8i min, \u03b8i max and \u03c3f are some known constants. In this paper, our objective is to design a control law for the input u such that the distance from the reference water level to the payload, d, tracks a bounded reference trajectory dr(t) as closely as possible in spite of model parametric uncertainties and external disturbances. From Fig. 1, one can see that the distance d can be expressed as d(t) = z(t) \u2212 xh(t) \u2212 L. Hence, the control objective is equivalent to the problem of regulating the following tracking error: e(t) = d(t) \u2212 dr(t) = z(t) \u2212 x1(t) \u2212 L\u2212 dr(t). (6) For simplicity, we make the following assumption. Assumption 2 The tracking error e(t) is measurable and the time derivatives of z(t) are all bounded. In this section, an adaptation law will be developed for the system Eq. (4) to estimate the uncertain parameters. Since the uncertain parameter set \u03b8 only occurs in the second equation of Eq" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001713_transducers.2011.5969296-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001713_transducers.2011.5969296-Figure3-1.png", "caption": "Figure 3. Schematic diagram of wiring elect up test devices. Each electrode is coupled potentiostat terminal. The data obtained transfer to a computer through USB cable.", "texts": [ " Next, the waf at 80\u00b0C for 10 min to remove absorbed that could prevent uniform coating o e selectivity and over different catecholamine scan rates (1- PRINCIPLE for detecting on the oxide, a thin electrode ited by physical covery 18) and andard micro (Suss MicroTec er was prebaked water molecules f the CD-based \u2018catcher\u2019 molecular layers. T was dipped into a solution amino-groups (2nd step in F step was followed by ad modification steps: coating acid (CA) with NHS/EDC coating of each CDs wit immobilization, as described On top of the fab electrochemical cell was polycarbonate tube (OD = (Loctite, Quick SetTM Epox confines the testing object solutions. To avoid therma we utilized room temperatu adhesive paste (Alfa Aesa fabrication processes were published at microTAS 2010 Figure 3 shows the te neurotransmitter sensor. Eac sensor was connected t instruments, Reference 60 voltage variations under monitoring was recorded ut recording system (VFP 600 additional Ag/AgCl electro electrode). Target catechola contained in polycarbonate testing. The fabricated se mixtures of catecholamine L-Dopa, and L-tyrosine), c and scan rates (1-100mV/s) electrode were measured in solution to measure the indiv coated electrode. Next, I while increasing the conc sensitivity. Finally, the optimize the detection condi describing the --cyclodextrins at on describing the ers' for physical rodes and setting to a designated by potentiostat hen, the moisture-free wafer of 1mM cysteamine to form ig" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000940_icelmach.2012.6349890-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000940_icelmach.2012.6349890-Figure3-1.png", "caption": "Fig. 3 shows the applied line voltage waveforms. Table I shows the analysis conditions for the magnetic field analysis. This analysis is performed using eight nodes (64 CPUs) of the ES2 [5].", "texts": [ "5\u00b0 is approximately 10% of the average torque. In addition, the torque around x-axis is almost zero. Fig. 7 shows the off-center angle and the torque around yaxis. The average torque around y-axis increases linearly by the influence of the off-center of the rotor. Fig. 8 shows distributions of nodal force vectors of stator core. The nodal force in the circle as shown in Fig. 8 is large, because the gap of the motor with off-centered rotor is small. Table II shows the discretization data for the magnetic field analysis. Fig. 3. Applied line voltage waveforms. TABLE I ANALYSIS CONDITIONS (MAGNETIC FIELD ANALYSIS) Fig. 4. Current waveforms (phase U). Fig. 6. Torque around y axis of rotor core. TABLE II DISCRETIZATION DATA (MAGNETIC FIELD ANALYSIS) Computer used: ES2 cu rr en t (p .u .) electrical angle (\u00b0) off-center angle\u03b8(\u00b0) -1.5 -1.0 -0.5 0.0 0.5 1.0 1.5 0 120 240 360 0.0 0.5 Rotation speed (min-1) 935 Frequency of power supply (Hz) 46.75 Number of coil turns 7 Coil resistance (p.u.) 0.038 Magnetization of magnet (T) 1" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000690_icems.2011.6073983-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000690_icems.2011.6073983-Figure2-1.png", "caption": "Fig. 2. Structure of three phase 6k/4k DSEG", "texts": [ " 6/4-pole DSEG is the basic unit of three-phase DSEG, and we define it as DSEG element. Multi-poles DSEG is combined by DSEG elements, and can be expressed by 6k/4k-pole DSEG, where k is positive integers. Taking 12/8-pole DSEG as an example, k is equal to 2 and it is repeated twice by 6/4- pole DSEG elements, as shown in Fig. 1(b). III. THE CONFIGURATIONS OF MULTIPLE ROTOR POLES DSEGS If we remove the odd stator pole or the even stator pole of 12/8-pole DSEG and double the winding turns of the remain stator pole, the 6/8-pole DSEG element is obtained, as shown in fig. 2(a). Compared with 12/8-pole DSEG, their EMFs are equal, while the field winding element number is reduced by 50% as the stator pole number is halved. So the loss of field windings and stator core can be reduced. Moreover, the stator pole width is a quarter of the stator pole distance and the stator slot area is increased by more than two times. The 12/16-pole DSEG is combined by two 6/8-pole DSEG elements, as shown in Fig. 2(b). The stator pole width of 12/16-pole DSEG is half the pole width of 6/8-pole DSEG. If the airgap flux densities of the two generators are the same, the flux per pole of the 12/16-pole DSEG is decreased by 50%. However, the rotor pole number of 12/16-pole DSEG is two times than that of 6/8-pole DSEG, and so as the flux frequency of the stator pole. If the phase winding turns of the two DSEG are the same, the EMFs induced in phase windings are equal. In the same way, if we remove the odd stator pole or the even stator pole of 12/16-pole DSEG and double the winding turns of the remain stator pole, the 6/16-pole DSEG element is obtained, as shown in Fig. 2(c). The stator pole width is one eighth of the stator pole pitch and then the stator slot area is further increased. Fig. 2(d) shows the configuration of 24/16-pole DSEG, which is repeated four times by 6/4-pole DSEG elements. Compared with 6/16-pole DSEG, we can get the following results: \u2666 The stator pole and field winding element of 6/16-pole DSEG are a quarter of those of 24/16-pole DSEG. And the total stator slot area is larger than that of 24/16-pole DSEG if the inner stator diameter and the height of stator pole are the same. \u2666 The phase EMFs of them are also equal if the inner stator diameter, the air gap flux density, the rotation speed, and the phase winding turns are the same" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001965_ijvnv.2012.046176-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001965_ijvnv.2012.046176-Figure1-1.png", "caption": "Figure 1 Schematic of biped robot", "texts": [ " A popular theory regarding legged locomotion is that humans (or animals) move in a way that minimises energy in one form or the other (Minetti and Alexander, 1997; Srinivasan and Ruina, 2007; Srinivasan, 2009; Kuo et al., 2005). Inverted pendulum walking is one of the idealised gaits that can be used to describe walking. In an inverted pendulum walking the hip moves in series of circular arcs, vaulting over a straight leg. The transition from one circular arc to the next is realised via impulsive push-off (Kaneko et al., 2004) that happens just before heel strike. For the first part of the paper, a simple biped robot with torso (Grizzle et al., 2001) is considered, which is shown in Figure 1. The biped robot, confined to the sagittal plane, is composed of two identical rigid legs and a torso, which are connected at the hip. The biped robot has four point masses. The values of the robot parameters are shown in Table 1. One typical step is composed of a swing phase and a collision phase. In the swing phase, only the stance foot is in contact with the ground without slipping. In the collision phase, which is thought to be instantaneous, the swing foot touches the ground and the swing leg becomes the new stance leg and vice versa" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003185_s0001925900001827-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003185_s0001925900001827-Figure3-1.png", "caption": "FIGURE 3. Integral curves in the region of the singularities, B < 0", "texts": [ " Tufts Univ, on 24 Mar 2018 at 17:33:39, subject to the Cambridge Core terms of use, available at S H O R T - P E R I O D M O T I O N It is now appropriate to plot the integral curves in the phase plane. In the cases where Ax and A3 are of the same sign, then the singularity occurs only at the origin and the non-linear terms do not change its character. When Ax and A3 are of different sign, then two additional symmetrically disposed singularities arise whose natures depend on the signs ofH,Ai,A3 and the discriminant H2 + AG. The three interesting cases are those where the singularities away from the origin differ from those at the origin and are sketched in Fig. 3. The curves of Fig. 3(a) correspond to an M (w) curve such as c in Fig. 2, and imply that for initial values of w between '\u00b1ws the airframe will perform a damped oscillation. When zw, and hence B19 is large, then BX 2 + 4AX > 0 and the motion near the origin degenerates into a subsidence. In the quasi-linear approximation mw will change sign at the minimum in the Miyv) curve, and (17) will become unstable when df2 (w)/dw changes sign. Differentiation gives dw which changes sign when ={U0 + zQ)mw-mtzw + 3 {([/0 + zg)m3-m,,z3} w2, w -+\\~^; + Zimw + maZwYl2=wm" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001817_isie.2011.5984290-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001817_isie.2011.5984290-Figure3-1.png", "caption": "Fig. 3. Contact points Fig. 4. Definition of ffi", "texts": [ " Each joint has an actuator and an encoder. The 3D model of the robot is shown in Fig. 2. x-axis is a forward direction, y-axis is a horizontal direction, and z-axis is a vertical direction. TABLE I shows parameters of links. In this research, the robot is constrained in the sagittal plane and considered as a 2-dimensional biped robot. Besides, in order to measure ground reaction forces during walking, force sensors consisted by strain gauges are attached. Each sole of the robot is a rectangle, as shown in Fig. 3. Each corner of the robot has a uniaxial force sensor that detects a vertical force. The reaction force at each corner of the sole is represented as e1, e2, e3, and e4. The robot has the gyroscope on the body. The gyroscope measures an angular velocity of the body, and the angle of the body \u03c6body is calculated by the numerical integration. As shown in Fig. 4, \u03c6 is defined as a pitch angle of the sole plane. In case that the sole plane is parallel to the horizontal, \u03c6 = 0 is achieved. In this section, environmental mode compliance control [7] is explained. If the sole is a rigid rectangular plane and the number of contact points is four (Fig. 3), environmental modes are classified into four kinds, heaving, rolling, pitching, and twisting. These modes are expressed as shown in Fig. 5. Environmental modes are calculated from sensor information as follows:\u23a1 \u23a2\u23a2\u23a3 eh er ep et \u23a4 \u23a5\u23a5\u23a6 = \u23a1 \u23a2\u23a2\u23a3 1 1 1 1 1 \u22121 1 \u22121 1 1 \u22121 \u22121 1 \u22121 \u22121 1 \u23a4 \u23a5\u23a5\u23a6 \u23a1 \u23a2\u23a2\u23a3 e1 e2 e3 e4 \u23a4 \u23a5\u23a5\u23a6 = H4 \u23a1 \u23a2\u23a2\u23a3 e1 e2 e3 e4 \u23a4 \u23a5\u23a5\u23a6 (1) where eh, er, ep, and et are defined as the environmental modes of heaving, rolling, pitching, and twisting, respectively. H4 is the fourth-order Hadamard matrix" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000282_icems.2011.6073966-Figure6-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000282_icems.2011.6073966-Figure6-1.png", "caption": "Fig. 6 The axial edgy of Motor-A", "texts": [ " In this case, as the field generated by the armature winding currents in the axial edge is normally much weaker then the field generated by the magnet, the magnetic pull produced by the permanent magnet is constant. It also means, for the slotless PMAC motor, if the UMP in one rotor position can be known, the UMP on the other positions can also be known. However, the existence of the stator slots makes the geometrical relationship between the rotor and stator be complicated. The influence of the edgy magnetic field varies with the rotor rotation. This makes the UMP varies with the rotor rotation. One example is the motor shown in Fig. 5 and Fig. 6. It is a spindle motor with outer rotor. From the viewpoint of the magnetic circuit, the effective airgap of the motor axial edgy can be expressed with the simplified model shown in Fig. 7. From he viewpoint of the rotor, the length of the local airgap changes when the rotor is at different position. IV. UMP CHARACTERISTIC IN THE SPACE DOMAIN From Fig. 7, the effective permeance of the axial edgy airgap can be expressed as 0 0 ( ) [1 ( )]\u03b8 \u03bb \u03b8 = \u039b = \u039b + \u22c5\u2211A A An n Cos nZ , (1) where, \u039bA0 is the effective average permeance of the airgap, which is related with \u03b5, the difference between the lengths of rotor and stator; see Fig", " In this case, as the denominators of (11) and (12) are formed by the even harmonics whose order is also the multiple of the magnetic poles, it can be concluded that the variation of the UMP center is formed by the even harmonics in both X and Y directions, and the order of the harmonic is the multiple of the motor magnetic poles. VI. NUMERICAL ANALYSIS ON THE UMP For verifying the analytical results obtained in the Section IV and V, several PMAC motors are calculated with 3D finite element method (FEM), and the FEM results are compared with the analytical ones. Here, the calculation results for two spindle motors are introduced, and they are named as Motor-A and Motor-B, separately. Motor-A has 12 stator slots and 4 magnetic pole-pairs, whose structure has been shown in Fig. 5 and Fig. 6. Motor-B is shown in Fig. 9. It has 3 stator slots and 2 magnetic pole-pairs. Both these two types of motor EM structure can be found in many applications. Fig. 10 shows the flux lines of Motor-A obtained with 3D FEM. From the analytical analysis in the Section-IV, as Motor-A has 12 stator slots and 4 magnetic pole-pairs, its minimum common multiple of Z and 2p is 24. Therefore, the cycle width of fundamental harmonic of the UMP is 15\u00b0. This analytical deduction is confirmed by the UMP curve shown in Fig" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001045_icma.2013.6618057-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001045_icma.2013.6618057-Figure3-1.png", "caption": "Fig. 3. Nut attachment task.", "texts": [ " (n - Sni) T (25) Using this, we define the objective function gi as eni gi (ai, j3i) = L 1'; (nT; ai, j3i) (26) Thus, the optimization problem (23) is equivalent to the optimization problem with respect to g;: mmUlllze gi (ai, j3i) subject to ai > j3i j3i > 0 Dmax a\u00b7 < -t - 2M 2 _ j3 2 < Kmax at t - M (27) This optImIzation problem includes the nonlinear objective function and nonlinear inequality constraints. Hence, a nu merical scheme is required to solve the problem. IV. ApP LICATION To verify the effectiveness of the method described in the previous sections, we did skill acquisition experiments. In order to compare the distinctions between our previous method and the proposed one, we chose the same task treated in our previous work [lO][ll]. The target task is the nut attachment task. A screw nut held by the slave arm is fastened to the bolt fixed to the ground, as shown in Fig. 3. The base coordinate system \ufffdb and nut coordinate system \ufffdn are placed as shown in the figure. The initial position of the origin of \ufffdn was set to [0 , 0 , 0.02f (m) with reference to \ufffdb. The initial orientation of \ufffdn relative to \ufffdb was set to [0 , 0.15 , of (rad) , where the orientation is represented by roll-pitch-yaw angles. We developed a bilateral teleoperation system illustrated in Fig. 1. A compact 6-DOF haptic interface, developed at Tohoku University [18][19], was utilized as the master arm, and a 7-DOF robot manipulator PA-1O was used as the slave arm, as shown in Fig" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000653_s10846-013-9993-5-Figure4-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000653_s10846-013-9993-5-Figure4-1.png", "caption": "Fig. 4 Angle of helicopter relative to earth reference frame", "texts": [ " The quaternions vector [ q0 q1 q2 q3 ] is solved by the standard set of differential equations: \u23a1 \u23a2 \u23a2 \u23a3 q\u03070 q\u03071 q\u03072 q\u03073 \u23a4 \u23a5 \u23a5 \u23a6 = \u23a1 \u23a2 \u23a2 \u23a2\u23a2 \u23a2 \u23a2\u23a2 \u23a2 \u23a2\u23a2 \u23a3 \u22121 2 (q3r + q2q + q1 p) 1 2 (q2r \u2212 q3q + q0 p) 1 2 (\u2212q1r + q0q + q3 p) 1 2 (q0r + q1q \u2212 q3 p) \u23a4 \u23a5 \u23a5 \u23a5\u23a5 \u23a5 \u23a5\u23a5 \u23a5 \u23a5\u23a5 \u23a6 (25) where: \u221a q2 0 + q2 1 + q2 2 + q2 3 = 1, q0 (0) = 1, q1 (0) = 0, q2 (0) = 0, q3 (0) = 0 (26) The time-varying rotational matrix corresponding to the unit quaternion [ q0 q1 q2 q3 ] is defined as: Rt = \u23a1 \u23a2\u23a2 \u23a3 2 ( q2 0 + q2 1 )\u2212 1 2 (q1q2 + q0q3) 2 (q1q3 \u2212 q0q2) 2 (q1q2 \u2212 q0q3) 2 ( q2 0 + q2 2 )\u2212 1 2 (q0q1 + q2q3) 2 (q0q2 + q1q3) 2 (q2q3 \u2212 q0q1) 2 ( q2 0 + q2 3 )\u2212 1 \u23a4 \u23a5\u23a5 \u23a6 (27) The orientation of the helicopter in terms of the body fixed frame is obtained using the equations: RA = \u23a1 \u23a2 \u23a3 RAx RAy RAz \u23a4 \u23a5 \u23a6 = Rt \u2217 X A0 PA = \u23a1 \u23a2 \u23a3 PAx PAy PAz \u23a4 \u23a5 \u23a6 = Rt \u2217 Y A0 Y A = \u23a1 \u23a2 \u23a3 Y Ax Y Ay Y Az \u23a4 \u23a5 \u23a6 = Rt \u2217 Z A0 (28) where XA0 = [ 1 0 0 ] , YA0 = [ 0 1 0 ] , ZA0 =[ 0 0 1 ] are the unit vectors in earth reference frame, RA, PA, YA are the resulting rigid body axis vectors. The angle of body frame axis with respect to earth reference axis is then calculated using trigonometric functions: \u03c6 = sin\u22121 (r) , r = PAz/cos (\u03b8) \u03b8 = sin\u22121 (RAz) \u03c8 = cos\u22121 (RAx/ \u221a RAx2 + RAy2) (29) \u03c6 \u2261 roll angle about RA axis \u03b8 \u2261 pitch angle about PA axis \u03c8 \u2261 directional angle about YA axis (30) Figure 4 shows the axes and angles relative to the helicopter body axis, where: RA \u2261 Axis parallel to the tail of helicopter from front nose PA \u2261 Axis pointing directly to the right of helicopter (directly south if the nose is pointing East) YA \u2261 Vertical axis perpendicular to main helicopter blades point down in the formation The parameters \u03c6, \u03b8 and \u03c8 in Eq. 30 by right hand rule are used to give an intuitive and easily visualized idea of the range of attitude dynamics of the helicopter during flight. 2" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000097_ijtc2011-61123-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000097_ijtc2011-61123-Figure2-1.png", "caption": "Fig 2 Four tilting pad cage design", "texts": [ " The analysis to determine the range of design preload as a function of oil and housing temperature has been automated in an Excel file. This allows a quick determination of the min and max expected clearance and preload for the cold build and also both no-load and loaded operation. The design is predicted to have a nominal 3 mil on diameter clearance with a preload range of 0.35 to 0.57. An offset of 60% was selected to increase the load capacity and reduce the pad operating film temperature. The bearing cage and pads are shown in Fig 2-a. The bearing assembly in a check fixture is shown in Fig 2-b. The predicted response to in-phase imbalance of 0.001 ozinch at compressor, station 8, and turbine, station 9 are shown in Fig 3 for the 3 mil on diameter clearance conditions. The second mode is predicted to be higher than with the floating ring bearing experimental data. The typical spectrum content on the stock bearings is shown in Fig 4 where the turbocharger is shown to be unstable in the first mode with a frequency in the 12-20 kcpm frequency range, even at the engine idle speed. The second mode instability comes in at a shaft speed near 80 krpm in the 35-38 kcpm frequency range" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003419_icase.2013.6785546-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003419_icase.2013.6785546-Figure1-1.png", "caption": "Fig. 1. A-3 Airliner Fig. 1 is a wide body airliner; it is the airliner in which 2 crew\u2019s cockpit is built. It has twin jet engines, supercritical wing design and T-tail. Detailed characteristics of A-3 are given in Table I.", "texts": [], "surrounding_texts": [ "Stability derivatives are required to calculate the transfer functions, which are determined by using DATCOM software. Longitudinal stability derivatives are in Table II. Theta/del e (change in pitch angle with the change in elevator deflection) transfer function is calculated by using longitudinal stability derivatives that is given below. Theta/del e: -1.6 s^2 - 0.96 s - 0.01 -------------------------------------------------- s^4 + 0.9 s^3 + 3.3 s^2 + 0.02 s - 0.003 Now comparison and root locus study is done on different displacement autopilots." ] }, { "image_filename": "designv11_101_0000011_amr.631-632.817-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000011_amr.631-632.817-Figure1-1.png", "caption": "Fig. 1 Applied coordinate systems of meshing spur gears", "texts": [ " Fong [6] proposed a mathematical model of parametric tooth profile of spur gear where the line of action is given. The paper illustrates the mathematical model with an example of a combinative curve that comprises a straight line (involute) and a circular arc (extended cycloid). Our paper aims to establish the mathematical model of tooth profiles of spur gears using a parabola as line of action. Geometrical modeling of tooth profiles based on line of action In the following discussions, coordinate systems \u03a3 1 (O1, x1, y1), \u03a3 2 (O2, x2, y2) and \u03a3 (P, x, y) are designated as illustrated in Fig. 1. Coordinate system \u03a3 is a fixed coordinate system whose origin O coincides with the pitch point P, while coordinate systems \u03a3 1 and \u03a3 2 are moving coordinate systems rigidly connected to the center of the driving gear and the driven gear, respectively. The line of action passes through the centrode point O. Assume that the equation of the line of action can be described in coordinate system \u03a3 as ( ) ( ) 0 0 0 0 x x y y \u03b8 \u03b8 = = (1) where \u03b8 is the parameter of the line of action. According to the meshing theory, transforming the equation of the line of action from coordinate system \u03a3 (X, O, Y) to \u03a31 (X1, O1, Y1), the equation of tooth profile of the driving gear can be expressed as follows" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001593_nems.2011.6017546-Figure4-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001593_nems.2011.6017546-Figure4-1.png", "caption": "Fig. 4 Driving principle of the CNC-based nano-robot by a magnetic field; the yellow arrows are the direction of the magnetic field and the blue arrows are the orientation direction of the nano-robot.", "texts": [ " The CNCs coated with MPC polymer have biocompatibility and are generally not harmful to living cells [6]. This work was supported in part by the Industrial Research Program of NEDO, Magnetic Health Science Foundation, Osaka Science & Technology Center, Sasagawa Science Research Grant, and Grants-in-Aid for Science Research from the Ministry of Education, Culture, Sports, Science and Technology in Japan (Nos. 20860031, 21676002, 20034017, and 21111503). 1089978-1-61284-777-1/11/$26.00 \u00a92011 IEEE Fig. 4 shows the driving principle of the nano-robot. The constant magnetic field aligns the magnetized CNC parallel to the direction of the applied magnetic field. This makes it possible to rotate the CNC by turning the applied constant magnetic field. A pair of coils facing each other, e.g. Helmholtz coils, can generate a uniform magnetic field between them. Results of a simulation for an applied magnetic field of facing coils are shown in Fig. 5. A rotating magnetic field generated by two pairs of coils can be obtained with the current through each coil as tII tII cos sin 2 1 (1) where I is the amplitude of the current and is the circular rotation frequency" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001425_ijptech.2011.038108-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001425_ijptech.2011.038108-Figure3-1.png", "caption": "Figure 3 Conversion of AutoCAD model into UNIGRAPHICS (see online version for colours)", "texts": [ " Following objectives have been set for present experimental study: 1 to study feasibility of decreasing the shell thickness from recommended one (12 mm) for statistically controlled RC solution of aluminium alloy in order to reduce the production cost and time 2 to evaluate the dimensional accuracy of the castings obtained and to check the consistency of the tolerance grades of the castings (IT grades) as per allowed IS standards for casting process 3 proof of concept, to present the concept in physical form with minimum cost by avoiding the cost of making dies and other fixtures for a new concept. In order to accomplish the above objectives, \u2018aluminium casting\u2019 has been chosen as a benchmark (Figure 1), representative of manufacturing field, where the application of RT and RC technologies is particularly relevant. The experimental procedure started with drafting/ model creation using AutoCAD software (Figure 2). For the process of RC process based on 3DP, following phases have been planned: 1 After the selection of the benchmark, the component to be built was modelled using a CAD (Figure 3). The CAD software used for the modelling was UNIGRAPHICS Ver. NX 5. 2 The upper and lower shells of the split pattern were made for different values of the thickness. The thickness values for shells thickness were 12, 9, 7, 6, 5, 4, 3, 2 and 1 mm. 3 The CAD models of upper and lower shells were converted in to STL (standard triangulation language) format also known as stereo lithography format (Figure 4). 4 Moulds were manufactured in 3DP (Z Print machine, Model Z 510) with Z Cast 501 powder, and parts were heat treated at temperature of 1100 C for one hour, The upper and lower shells were placed in such a way that the central axis of both the shells was collinear" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001277_amm.391.72-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001277_amm.391.72-Figure1-1.png", "caption": "Fig. 1 Spherical engine Fig. 2 Spherical cam", "texts": [ " The conventional piston engine was successful over time reducing consumption and meeting lower exhaust limits and increasing power at the same time any new concept face very high requirements to push through. In these days there is a talk about a novel spherical engine which is compact, very light and has high specific power and consumes less fuel. It is possible to mandate curvy pistons through a curvy way inside the sphere as opposed to conventional crank shaft and piston rod. The drive device of the spherical engine can be simplified as a spherical cam [1]. Similar to rotary piston engines, the pistons of the engine rotaries in a dimensional sphere as show in Fig. 1. Within the curvy channel two groups of pistons move. Each is mandated through the curvy way through a conic follower. Apart from that the two swinging pistons are connected through a central x-axis moving with them and thus form two groups of connected pistons. Due to the curvy channel, the rotary axis of the pistons rotates about the z-axis, thus the output speed of the engine can be obtained. All rights reserved. No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of Trans Tech Publications Ltd, www" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000836_ccdc.2012.6244435-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000836_ccdc.2012.6244435-Figure1-1.png", "caption": "Figure 1: The \"S\" shape characteristic of the pumped storage unit", "texts": [ " There is a lot of technical problems of the large pumped storage unit should be solved during its localization progress. One of which is how to lower the impact of the \"S\" shape characteristic of the pumped storage unit . The \"S\" shape characteristic has a unique impact on the stability of the unit because of its intricate condition transition and frequent start and stop. In the oval area of the full characteristic figure of the pump turbine which called the \"S\" shape characteristic of the pumped storage unit there are three different fluxes in one equal-opening line (Figure 1), which indicate the turbine mode, turbine brake mode and anti-pump mode, in this area the pump turbine has a bad running situation. The unit rev of the pump turbine is 1 / 2/Nn n D H= which n indicate the rev, D indicate the unit diameter of the pump turbine and H indicate the water head. When the pump turbine is starting under a lower water head, the no-load unit rev of the turbine is relatively higher, which make the turbine easily impacted by the inhibited \"S\" shape characteristic and then switched into the anti-pump mode, this will cause the no-load running of the turbine unstable and the rev fluctuating around the grid frequency; under this situation the generator has a big problem to connect to the grid" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002278_978-1-4471-4628-5_3-Figure3.81-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002278_978-1-4471-4628-5_3-Figure3.81-1.png", "caption": "Fig. 3.81 Schematic diagram of vane rotary actuator", "texts": [ " The spool valve position xv is decided by the controller and it decides the port opening areas (discharge coefficients), and the chemical potentials and temperatures at the flow ports. The four MSe elements in the model impose external effort variables on the system. For some range of the spool valve position, \u03bca = \u03bcs , Ta = Ts , \u03bcb = \u03bcr , and Tb = Tr where subscripts s and r refer to supply and return sides. For other values of spool valve position, \u03bca = \u03bcr , Ta = Tr , \u03bcb = \u03bcs , and Tb = Ts . The cross-port leakage can be modeled between the two MR elements. Similarly, we can model the vane rotary actuator. Figure 3.81 shows the schematic diagram of vane rotary actuator. The equation of motion (mechanical part) for vane actuator can be written as V\u03b8 2\u03b80 (P1 \u2212 P2) \u2212 \u03c4e \u2212 \u03c4r \u2212 I \u03b8\u0308 = 0 (3.84) In Eq. 3.84, V\u03b8 represents vane motor displacement, \u03b80 represents half stroke, P1 and P2 are the absolute pressures at two sides, \u03c4e and \u03c4r are external torque and friction torque, respectively, I represents the rotary inertia of vane actuator and \u03b8 is the vane actuator displacement. The bond graph model of the vane actuator maintains the same structure as that for the linear actuator with some minor modifications: the rotary inertia must be substituted in place of the mass as parameter of the I element and both the transformer moduli A1 and A2 should be replaced by V\u03b8 2\u03b80 " ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001227_j.engstruct.2013.08.033-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001227_j.engstruct.2013.08.033-Figure2-1.png", "caption": "Fig. 2. (a) Configuration: Applied loads on a given geometry (b) Kinematics: Configuration with imposed virtual velocities.", "texts": [ " Therefore, as long as the subject is deemed important for engineering students to learn, it may be helpful to relate it to concepts of Statics that are already known to them. In reference to Fig. 1a and b, the virtual work rate done by the applied forces is given by: P t2y \u00fe Q 1 \u00bc 0 \u00f018\u00de Substituting Eq. (17) produces: Q \u00bc P 1 tan h1 \u00fe tan h2 \u00f019\u00de This example shows that, after establishing relations of Kinematics for a problem, static equilibrium becomes trivial when using virtual velocities. Assuming that all dimensions and angles shown in Fig. 2a are known, it is desired to find the relation between forces P and Q to satisfy the conditions of static equilibrium of the system. The work done by the applied forces can be expressed as: P dY1 Q d\u00f0Y1 \u00fe Y2\u00de Q d\u00f0X1 \u00fe X2\u00de \u00bc 0 \u00f020\u00de Referring to Fig. 2a and letting L12 and L23 be the lengths of element 1\u20132 and element 2\u20133, the vertical and horizontal distances from joint 1 to joint 2 and from joint 1 to joint 3 are given by: Y1 \u00bc L12 sin h1 Y1 \u00fe Y2 \u00bc L12 sin h1 \u00fe L23 sin h2 X1 \u00fe X2 \u00bc L12 cos h1 \u00fe L23 cos h2 \u00f021\u00de Differentiating: dY1 \u00bc L12 cos h1 dh1 d\u00f0Y1 \u00fe Y2\u00de \u00bc L12 cos h1 dh1 \u00fe L23 cos h2 dh2 d\u00f0X1 \u00fe X2\u00de \u00bc L12 sin h1 dh1 L23 sin h2 dh2 \u00f022\u00de Substituting the differentials above into Eq. (20) gives: P L12 cos h1 dh1 Q \u00bdL12 \u00f0cos h1 sin h1\u00de dh1 \u00fe L23 \u00f0cos h2 sin h2\u00de dh2 \u00bc 0 \u00f023\u00de To find dh2 in terms of dh1 the following geometric compatibility equations can be used: X1 \u00fe X2 \u00fe X3 \u00bc L12 cos h1 \u00fe L23 cos h2 \u00fe L34 cos h3 \u00bc L Y1 \u00fe Y2 \u00bc L12 sin h1 \u00fe L23 sin h2 \u00bc H \u00fe L34 sin h3 \u00f024\u00de Differentiating these two expressions (Eq", " (24)) gives: L12 sin h1 dh1 L23 sin h2 dh2 \u00bc L34 sin h3 dh3 L12 cos h1 dh1 \u00fe L23 cos h2 dh2 \u00bc L34 cos h3 dh3 \u00f025\u00de Dividing the last two expressions allows eliminating dh3: L12 sin h1 dh1 L23 sin h2 dh2 L12 cos h1 dh1 \u00fe L23 cos h2 dh2 \u00bc tan h3 \u00f026\u00de Rearranging: dh2 dh1 \u00bc L12 \u00f0sin h1 \u00fe cos h1 tan h3\u00de L23 \u00f0sin h2 \u00fe cos h2 tan h3\u00de \u00f027\u00de Substuting this in Eq. (23) produces: Q \u00f0sin h1 cos h1\u00de \u00fe \u00f0cos h2 sin h2\u00de \u00f0sin h1 \u00fe cos h1 tan h3\u00de \u00f0sin h2 \u00fe cos h2 tan h3\u00de \u00bc P cos h1 \u00f028\u00de Solving for Q: Q P \u00bc 1 \u00f0tan h1 1\u00de \u00fe \u00f01 tan h2\u00de \u00f0tan h1 \u00fe tan h3\u00de \u00f0tan h2 \u00fe tan h3\u00de \u00f029\u00de Referring to Fig. 2b, as a unit (virtual) angular velocity is applied to member 1\u20132 (x12 = 1), the velocities of joints 2 and 3 can be found using Eqs. (3) and (4) as follows: ~V2 \u00bc k\u0302 \u00f0X1 i\u0302\u00fe Y1 j\u0302\u00de \u00bc Y1 i\u0302\u00fe X1 j\u0302 \u00f030\u00de ~V3 \u00bc ~V2 \u00fex23k\u0302 \u00bdX2 i\u0302\u00fe Y2 j\u0302 \u00bc \u00f0 Y1 x23Y2\u00de\u0302i\u00fe \u00f0X1 \u00fex23X2\u00de\u0302j \u00f031\u00de The angular velocity of member 2\u20133, x23, can be computed from the condition that joint 4 cannot move: ~V4 \u00bc ~V3 \u00fex34k\u0302 \u00bdX3 i\u0302 Y3 j\u0302 \u00bc~0 \u00f032\u00de Solving for ~V3; equating to Eq. (31), and separating into orthogonal i- and j-components, x23 can be found from the resulting system of two equations: x23 \u00bc \u00f0X1Y3 \u00fe X3Y1\u00de \u00f0X2Y3 \u00fe X3Y2\u00de \u00f033\u00de It is interesting to note, by comparison with Eq" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001814_icmech.2011.5971188-FigureI-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001814_icmech.2011.5971188-FigureI-1.png", "caption": "Figure I The principle sketch of a safety clutch", "texts": [], "surrounding_texts": [ "using an original experimental setup dedicated for studying the influence of surface material and topography on tribological behavior. The experimental setup is provided with a safety clutch with two steel disks. On the frictional disk, permanently maintained, a ferodo ring is bonded. The other disk, with the material to be tested, was changed each time. Silicon microtextured wafers and chromium (non)textured thin layers deposited on silicon wafers were tested. The size and shape of the depressions were varied and their influence on the friction coefficient was studied. The angular speed of driver and driven shaft was measured with two non contact tachometers.\nKeywords-- safety clutch, surface material, tribological behavior, silicon microtextured wafer\nI. INTRODUCTION\nIt is well known that the tribological perfonnance of materials is highly dependent on the surface topography [1,2]. The introduction of specific microstructures on a sliding surface, involving flat and smooth zones interrupted by local depressions arrays of quadratic openings, microgrooves or micropores, can improve the tribological properties. These textures act as lubricant pockets retaining the oil in the contact even under high pressures; they also trap wear particles generated during movement, reducing the ploughing component of friction. Thus the friction is reduced and the lifetime of components increased. The high accuracy, high resolution and freedom in choice of shapes, make the etched silicon wafers an interesting alternative to study the effect of surface material and texture [3,4].\nThe paper deals with a study concerning influence of surface material and topography on tribological behaviour. The determinations were developed using an experimental setup, which mainly consists in a safety clutch. There were considered different materials and technologies for the mentioned clutch.\n978-1-61284-985-0/11/$26.00 \u00a92011 IEEE\nThe safety clutches, which can be considered a part of the group of intermittent clutches, permit the automatic disengage of the link between two shafts, when the load or the speed of these ones exceeds the admitted limit. They are, therefore, automatic clutches, used especially in automatic systems, which are not overseen during the operating time. The most practical ones are the slipping clutches that, at overcharge, when the transmitted torque is greater than the friction torque, permit the relative slipping of the surfaces making the link [5, 6].\nA usual safety clutch, which can be a slipping clutch, is represented in Fig. 1. If the transmitted torque is greater than the friction torque obtained by the elastic force of the spring, the shafts connection stops. The magnitude of the friction torque and of the transmitted torque too, is adjusted through corresponding tensioning of the spring, by means of a screw - nut device. The dependence between force and spring deflection was determined experimentally using a Hans Schmidt force setup. The frictional disk made from ferodo is ring-shaped with inside radius rm, respectively outside radius rM. The friction surface lies between rm and rM.\nFor the tested materials, well-defmed surface textures were produced by photolithography and anisotropic etching of silicon wafers (only covered with a very thin layer of native oxide 50 A) and, also, by photo etching a thin film of chromium wear resistant, which was deposited using a standard PVD process on a smooth (non-textured) silicon wafer [7,8]. Friction coefficient determining tests at the textured silicon / chromium - ferodo disk interface were perfonned.", "II. EXPERIMENTAL SETUP AND PROCEDURE\nA. Experimental Setup Presentation\nThe sketch of the experimental setup is represented in Fig. 2. The following notations were made: 1 - ALU 427 - power supply for stepping motor: input voltage 220V, output voltage 42 V, and output current 7 A; 2 - MS - power supply for Inter CNC control board, 24 V c.c., lSOW; 3. - SD S042 - stepping motor driver (servo-amplifier) SOY, 4.2 A; 4 - MPP - stepping motor (1.8 Nm, 1,8\u00b0); S - Inter CNC - control board for maximum 4 axes control (here is used only one axis); 6 - Tl, T2 - non contact tachometers (Tl for motor shaft and T2 for driven shaft) RM-ISOO; 7 - PC - personal computer.\nThe actuating system for stepping motor consists of: AL 427 power supply, SD S042 stepping motor driver (servo amplifier), MS - power supply for Inter CNC control board, Inter CNC control board. The system is developed by SOPROLEC.\nThe driver shaft was marked with I and the driven shaft was marked with II. The Tachometer Tl shows the rotational speed of shaft I and the tachometer T2 shows the rotational speed of shaft II. The tachometers are connected to PC via two RS232 interfaces. The tachometers are developed by TECPEL.\nThere are two screw - nut devices, which control the spring forces, one for the brake spring and the other one for the clutch spring. The deflection of each spring can be read on graduated rulers, not figured in the picture.\nThe characteristics for brake spring and clutch spring were determined using the apparatus shown in Fig. 3. Here, one can determine the force and the corresponding deflection of the spring at the same time.\nFor both springs, the characteristics were determined. It is important to control the spring forces in order to determine the friction torques.\nThe operating procedure for angular speed detection is presented in Fig. 4.\nThe tachometers are non contact detection means. Each tachometer detects the angular speed of the steel disk (half clutch), respectively for driver and driven shaft.\nOn each disk, a reflective marker for incident beam reflection has been stuck, as it is shown in figure 4 sketch. When the experimental setup is actuated in the first step, the brake does not act on the driven shaft, so that the angular speeds for driver and driven shafts are equal. This is shown on the tachometer displays.\nIn the second step, the brake acts gradually, the angular speed of driven shaft slowly goes down, meanwhile the driver shaft has the same angular speed from the beginning of setup operation.\nAt the early beginning of the second step, when the brake starts to operate, the angular speed has fluctuations similar to stick-slip phenomenon.\nThe angular speed has small variations, increases and decreases, for a short time. After the brake force increasing, the angular speed of the driven shaft goes down and then, in the third step, the driven shaft stops. At this moment, the friction torque equals the clutch torque." ] }, { "image_filename": "designv11_101_0000552_amm.86.688-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000552_amm.86.688-Figure3-1.png", "caption": "Fig. 3 Assembled model of the gear of a spiral bevel gear for cutting", "texts": [ " Through a NC manufacturing dynamic transformation, the relative motion of a cutting tool and a workpiece can be determined by the machine tool adjusting parameters. Fig. 2 shows a simplified machine tool with a cutting tool tilting crandle mechanism. The cutter tool path relative to the workpiece is: mwmmmw RMMR \u2032= . (4) mwmmmw cMMc \u2032= . (5) The above equations give a way to decode every moving coordinate axis and machine tool adjusting parameters. Gears share the same calculating process as the above. A model of the tool and the gear was constructed and assembled in CATIA. Fig.3 shows the mesh model and boundary conditions definitions in Abaqus. Cutter teeth rotated around the axis of the cutting tool, cutting the roughcast from inner side to the outer. Meanwhile, the origin of the cutter rotated around the origin of the whole machine tool. The roughcast is also defined to rotate around axis of the gear. Their rotating velocities honor the roll ratio. In this paper, analysis was performed with cutting speed and cutting depth as variables. Velocities of 35m/min, 30m/min and 45m/min were considered, with a total cutting depth of 1mm" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000733_amr.479-481.1409-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000733_amr.479-481.1409-Figure1-1.png", "caption": "Fig. 1 Hob parameters diagram Fig. 2 Hobbing coordinate system diagram", "texts": [ " Based on ANSYS /LS-DYNA software, this paper establishes a complete model of full radius hob processing gear according to full radius hob processing theory and completes dynamic contact simulation of gears. The simulation gets accurate data and has done some analysis which demonstrates feasibility of full radius hob processing gear\u2019s application. Relative to the other method of gear cutting, hobbing has the advantage of high efficiency. As a result of it, it is widely used in gear production. Parameters of hob are shown in Fig. 1. Full radius hob only has one fillet on top of cutter profile. It is say that tooth tip of the hob is one continuous full arc. Fillet cure of gear tooth and root circle arcs are envelope formed by the hobbing of arcs on hob teeth\u2019s top. In accordance with the trajectory of hob tooth fillet, the envelope curve is a continuous normal offset curve of a prolate involute. All rights reserved. No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of Trans Tech Publications, www.ttp.net. (ID: 130.216.129.208, University of Auckland, Auckland, New Zealand-18/06/15,04:50:21) In Fig.1, \u03b1\u2014\u2014active cutting edge profile angle; ra\u2014\u2014 root fillet radius; hc\u2014\u2014protuberance height \uff1b Full radius hob are widely used in heavy industry to produce gear before grinding because of its products with high bending fatigue strength compared to products of hob with two fillet. Model building. In order to exclude minor factors to facilitate data analysis, this paper set up two standard spur gears with identical parameters for dynamic contact analysis. Gear\u2019s modulus is 5 and teeth numbers are 20.Gear\u2019s face width is 20" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002117_icef.2012.6310331-Figure8-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002117_icef.2012.6310331-Figure8-1.png", "caption": "Figure 8. the distribution of von Mises stress", "texts": [ " 6, it can be seen that \u02c6 /, ,F Vi z i\u2212 and \u02c6 /, ,~F Vi z i in segments 14-16 (the nose part) are much larger than in other parts. \u02c6 /, ,F Vi z i\u2212 mostly lies in positive z-direction expect segments 11-13 and segments 18-20. The distribution of the deformation is shown in Fig. 7.The maximum displacement is 0.851\u00d710-4 m appears in the nose part. It is clear that the value gradually is decreasing from the nose part to the straight part. The von Mises criterion is one of the yield criteria for ductile materials. The distribution of von Mises stresses is shown in Fig. 8, the maximum value is 5.29Mpa which appears in the straight part. Compared with the yield strength of copper, 69MPa, it is small. V. CONCLUSION This paper studies the steady-state magnetic forces on the end-winding of Multi-phase induction machine. The magnetostructural coupling field is used to analyze the deformation and the stress caused by the forces. The force is divided into the constant component and sinusoidal component. The result shows that the force density owns the maximum value in the involute part" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003004_ut.31.067-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003004_ut.31.067-Figure2-1.png", "caption": "Fig 2: (a) Inverted pendulum on a cart, 2 DOF {q, x}, usually only actuated in x; and (b) Pendubot, 2 DOF {q1, q2}, usually only actuated on one of the two joints", "texts": [ " In section 4, two energy shaping controllers for the robotic fish are outlined. Section 5 briefly discusses the addition of directional control to the system. In section 6 the method of plant simulation used are presented, along with simulation results for the two controllers. Section 7 contains a discussion of the results presented in the previous section; and finally, section 8 presents some conclusions from this study. The majority of researchers in the field of underactuated systems have focused on two main problems: the cart pendulum as shown by Fig 2(a); and the Pendubot (Spong, 1995) as shown by Fig 2(b), each being a simple 2 DOF under-actuated device, with non-linear dynamics. Each of these devices can be described by a system of non-linear ordinary differential equations of the form of equation 1. Joint variables q = [q, x] for the cart pendulum and q = [q1, q2]\u2032 for the Pendubot both are describable through state spaces given by x = [q\u2032 q \u2032]\u2032. Researchers have focused on two main control tasks for these devices, swing up control and orbital stabilisation. Swing up control of the cart pendulum requires the state space transition [p,0,0,0]\u2032 \u2192 [0,0,0,0]\u2032" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001236_elektro.2012.6225652-Figure4-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001236_elektro.2012.6225652-Figure4-1.png", "caption": "Fig. 4. Phasor diagram of synchronous motor.", "texts": [ " It is obvious that a big emphasis must be made to look for the possibilities to prevent the decrease of the conventional synchronous generators installed capacity or on the other hand to outfit the distributed energy resources with the means of their own inertia contribution. From this it is apparent that the need of the load angle changes investigation is important in the process of preservation of secure and reliable power system operation. III. LOAD ANGLE ESTIMATION METHOD The proposed method of load angle estimation is based on the phasor diagram construction. It is known that phasor diagram represents the voltage and current conditions of synchronous machine (generator) at different voltage or load levels. The principle of this idea is shown in Fig. 4 where the phasor diagram of synchronous motor is shown. Voltages in synchronous motor must be kept in balance which means that phasor sum of voltages Ef, Xq.Iq and Xd.Id have to conclude at a point identical with apex of stator phase voltage UsphN as it is shown in Fig. 4. Given that all parameters of synchronous machine as Rs, Ld, Lq and Ef (stator resistance can be neglected) are known it is possible to quantify load angle without the position sensor. The input values of proposed estimation model are stator phase voltage, stator current and power factor. The load angle estimation model works with iteration calculations. In every iteration the load angle is increased and subsequently synchronous machine currents Id and Iq are calculated. With known values of Ld and Lq inductances the voltage drops can be easily calculated and connected to phasor of induced voltage Ef" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002278_978-1-4471-4628-5_3-Figure3.60-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002278_978-1-4471-4628-5_3-Figure3.60-1.png", "caption": "Fig. 3.60 End view of single phase squirrel-cage induction motor", "texts": [ " A squirrel cage, which is an exercise tool for pet or captive hamsters and squirrels allowing them to run fast in a limited space, is made of free to rotate drum fixed on an axle with two end plates and short bars connecting the two drums. An electric motor with a rotor resembling a squirrel cage (squirrel-cage rotor) is termed a squirrelcage motor. The actual rotor is a cylinder mounted on a shaft where longitudinal conductive bars (usually, copper or aluminium bars) are set into grooves inside the rotor drum and these bars are connected at both ends by shorting rings. Figure 3.60 shows the end view of the single phase squirrel-cage induction motor. These conductor bars fit into slots in end rings and form the circuit in which current is induced. There is no external connection to the rotor. The stator consists of set of windings. When an alternating current passes through the stator windings an alternating magnetic field is produced. Because of the alternating magnetic field, electromagnetic induction takes place in rotor and emf are induced. Thus, current flows through the rotor conductor bars" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001653_imece2013-63679-Figure13-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001653_imece2013-63679-Figure13-1.png", "caption": "Figure 13: Receiver connection design, Team B.", "texts": [ " Comparing these final designs to the original selected designs in the options reviews, Figures 5 through 8, it is clear that each team improved the details of their design while maintaining the same concept. Team A continued with the \u201cStanding Rack\u201d design with minimal changes in the folding arms. Team B improved the \u201cBi-Fold Style\u201d significantly by having a double swing arms instead of one to decrease the needed clearance area behind the car. One further development by team B was the unique design of the hitch connection. They developed a full detailed design of a connection that can be used for 2 and 1.25 inch receivers. Figure 13 shows the connection design for team B. Figures 14 through 17 are the summaries that the teams presented with each design to show how they met the design weight and price target. Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 01/30/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use 7 Copyright \u00a9 2013 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 01/30/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use 8 Copyright \u00a9 2013 by ASME Figure 16: Cost details, Target Budget is $391" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001526_powereng.2013.6635832-Figure6-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001526_powereng.2013.6635832-Figure6-1.png", "caption": "Figure 6. Two different switching combinations of vector v1 b) (onn)", "texts": [ " As the voltage evolution for a given capacitors will be different for each state as shown in this section, this redundancy permits to control the capacitors voltages while the requested vector voltage is supplied [7]. The large vectors and the zero vectors do not change the voltage of neutral point. For the medium vector, there is only one vector for a specific direction. The line current flows through the neutral point for a given vector and the NPP is then affected. The compensation of voltage capacitor balance has to be given to the next medium vector because this vector could flow opposite current from the capacitor bank. Fig. 6 shows an example of one small vector given by two different switching combinations. Both combinations produce POWERENG 2013 the same output voltage v1, but when the first combination is applied, the current flows into the neutral point and produces discharge in the capacitor Cp and with the second it follows out and produces a charge in the same capacitor. This property provides the freedom to control the voltage of the neutral point. Based on this property, a control strategy will be developed and applied to a three-level NPC VSI" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000391_isam.2013.6643498-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000391_isam.2013.6643498-Figure3-1.png", "caption": "Fig. 3. Suppose the stress at angular posit force balance equations for angular contact ba", "texts": [], "surrounding_texts": [ "Keywords\u2014 Ball bearing; Non-uniform prel\nI. INTRODUCTION It is well known that the preload control\nof the most effective ways to adjust performance. Reasonable preload can stiffness, restrain bearing skidding, and the life, reduce the noise and vibration of spindl the accuracy of rotation. Related studies show in high-speed spindle can improve by 6 temperature also drops by 41.6% under reas 2].\nThe constant pressure preload is a popul method in the high speed spindle. Commonly apply uniform force on the bearing oute considering the shaft error resulting from manufacturing, assembling process and the s the spring, bearing preload actually is uniformly, as shown in Fig. 1. Under the non which means a combination of a force alo moment acting on the bearing, the bearing ou at a small angle with respect to inner ring, a In this occasion, the stresses, deformations o at different azimuth angles are also diffe stiffness character is also different from th preload.\n978-1-4799-1657-3/13/$31.00 \u00a92013 IEEE\nsufficiently ensure onally, rigid and rm forces on the errors, assembly preload actually ball bearing. The g characters under niform preload the le with respect to d deformations of diverse from each different from that aring deformation -uniform preloads ethod. The results tiffness and axial stiffness increases nder non-uniform\noad; Stiffness\ntechnology is one bearing serving increase bearing n extend bearing e system, improve that the stiffness 5% and bearing onable preload [1-\nar preload-control spring is used to r ring. However, the designing, tiffness change of distributed non-uniform preload, ng the axis and a ter ring will rotate s shown in Fig. 2. f rolling elements rent. So bearing at under uniform\nMany researches have been\nstiffness under uniform preload stiffness characteristics under presents an elastic deformatio under axial preload [4]. Based statics model, Yang Zuowei in the speed, preload, thermal def and invented a method to calcu bearing [5]. Mohammed studi machine spindle under axial pre a numerical iterative method ba analyse the bearing stiffness u those researches do not cons condition, and further research stiffness with non-uniform prelo\nFig.1. uniform and\nFig.2. the rotate unde\nIn this paper, the mathema under preload is introduced element method, in which the u an axial force and the non-uni combination of an axial force bearing outer ring, the bearin under the uniform and no Comparing the data, the axial s characteristics can be achiev preload conditions.\ncarried out on ball bearing\n. Aramaki verified the bearing the preload [3]. Igor Zverv n model of high speed spindle on the rolling bearing Quasitegrated various factors such as ormation and oil film thickness late the dynamic stiffness of the ed the vibration of a grinding load [6]. Liu Xianjun presented sed on finite element analysis to nder preload [7]. However, all ider the non-uniform preload is needed to study the bearing ad.\nnon-uniform preload\nr non-uniform preload\ntical model of bearing stiffness [8]. By using nonlinear finite niform preload is simplified to form preload is simplified to a and a moment acting on the g deformation data is obtained n-uniform preload condition. tiffness and the angular stiffness ed separately under different", "II. MATHEMATIC MODE\nThe bearing load can be expressed by [ xF the basis of 5DOF quasi-static mathematic m\nAs shown in Fig. 4, the preload is induce points of action are uniformly distributed o Different stiffness ki of spring will lead distributed preload. Based on the equivalent\ncan be simplified to an axial force ixF\u2211 iyM\u2211 izM\u2211 respectively acting on XZ-pla\nFig.4. the constant preload induced by n\n1\nn\nix i i F F = =\u2211 \u2211\n1 =1 2= cos( ( 1) ) 2iy\nn\ni i M F L i n \u03c0 \u03c0\u03b5\u23a1 \u2212 + +\u23a2\u23a3 \u2211 \u2211\nL\n; ; ; ; ]y z y zF F M M on odel, as shown in ion \u03c8 is Q\u03c8 , the ll bearing are:\nd by n springs and n the outer ring. to non-uniformly theory of statics it\nas well as torque\nne and XY-plane.\nsprings\n(6)\n\u23a4 \u23a5\u23a6\n(7)\n=1\n2= sin( ( n\niz i i M F L n \u03c0\u23a1 \u23a2\u23a3 \u2211 \u2211\nIf there is only non-uniform outer ring in the static analys come as follows x ixF F=\u2211 0yF = 0zF =\n1\n= iy\nn\ny i M M = \u2211\n1\n= iz\nn\nz i M M = \u2211\nOn the basis of Pythagore Fig.5 that\n( ) ( )2 1 1 2 2j j j jA X A X\u2212 + \u2212\n)(2 2 1 2 0.5j j oX X f\u23a1+ \u2212 \u2212\u23a3\nWhere 1 BDsinj iaA \u03b1 \u03b4 \u03b8= + + \u211c 2 BD cos cosj rA \u03b1 \u03b4 \u03c8= +\n)( 2cos 0.5 j oj o X f D \u03b1 \u03b4 = \u2212 +\n)( 1sin 0.5 j oj o X f D \u03b1 \u03b4 = \u2212 +\n)( 2 2cos 0.5 j j ij\ni\nA X f D \u03b1 \u03b4 \u2212 = \u2212 +\n)( 1 1sin 0.5 j j ij\ni ij\nA X f D \u03b1 \u03b4 \u2212 = \u2212 +\nFig.5. Positions of ball center and race\npositi\nThe relationship between n deformations in bearing is as fo\n11) 2 i \u03c0\u03b5 \u23a4\u2212 + + \u23a5\u23a6 (8)\npreload acting on the bearing is, the force balance equations\n(9)\n(10)\n(11)\n(12)\n(13)\nan Theorem, it can be seen in\n)( 22 0.5 0i ijf D \u03b4\u23a1 \u23a4\u2212 \u2212 + =\u23a3 \u23a6\n(14) 2\n0ojD \u03b4 \u23a4+ =\u23a6 (15)\ncos j\u03c8 (16)\nj (17)\noj\n(18)\noj\n(19)\nij\n(20)\n(21)\nway groove curvature centersat angular on j\u03c8\normal loads and normal contact llows", "1.5 oj oj ojQ K \u03b4= (22)\n1.5 ij ij ijQ K \u03b4= (23)\nConsidering the horizontal and vertical direction force, equilibrium equations of the j-th ball are\n( )sin sin cos cos 0gj ij ij oj oj ij ij oj oj\nM Q Q\nD \u03b1 \u03b1 \u03bb \u03b1 \u03bb \u03b1\u2212 \u2212 \u2212 =\n(24)\n( )cos cos sin sin 0gj ij ij oj oj ij ij oj oj cj\nM Q Q F\nD \u03b1 \u03b1 \u03bb \u03b1 \u03bb \u03b1\u2212 + \u2212 + =\n(25)\nWhere cjF is centrifugal force.\n2\n21 2 m cj m\nj\nF md \u03c9\u03c9 \u03c9 \u239b \u239e= \u239c \u239f \u239d \u23a0\n(26)\nConsidering the equilibrium condition of the whole bearing, equilibrium equations of the forces are\ni 1 b\nsincos cos 0 j Z\nij gj ij ij ij\nj\nM Q D \u03bb \u03b1 \u03b1 \u03c8 = = \u239b \u239e \u2212 =\u239c \u239f \u239d \u23a0 \u2211 (27)\ni 1 b\nsincos sin 0+ j Z\nij gj ij ij ij\nj\nM Q D \u03bb \u03b1 \u03b1 \u03c8 = = \u239b \u239e =\u239c \u239f \u239d \u23a0 \u2211 (28)\n1 1 b\ncos 0in s j Zi n\nij gj ix ij ij ij\ni j\nM F Q D \u03bb \u03b1 \u03b1 == = = \u239b \u239e \u2212 =\u239c \u239d \u2212 \u239f \u23a0 \u2211 \u2211 (29)\n1 1\nco[- sin\n- cos ( ) ]si\ns\nsin n 0\nj Zi n ij gj\niy ij ij ij ij i j b\nij gj ij i gj ij ij ij i ij j\nb b\nM M q Q\nD\nM r M Q p\nD D\n\u03bb \u03b1 \u03b1\n\u03bb \u03bb \u03b1 \u03b1 \u03c8\n==\n= =\n\u239b \u239e \u2212 +\u239c \u239f\n\u239d \u23a0 \u239b \u239e\n\u22c5 \u211c +\n+\n+ =\u239c \u239f \u239d \u23a0\n\u2211 \u2211 (30)\niy 1 1\ncos[ sin -\ncos ( )- ]cos 0sin\nj Zi n ij gj ij ij ij ij i j b\nij gj ij i gj ij ij ij i ij j\nb b\nM M q Q\nD\nM r M Q p\nD D\n\u03bb \u03b1 \u03b1\n\u03bb \u03bb \u03b1 \u03b1 \u03c8\n==\n= =\n\u239b \u239e \u2212 \u239c \u239f\n\u239d \u23a0 \u239b \u239e \u22c5 \u211c + =\u239c \u239f\u2212 \u23a0\n+\n\u239d\n\u2211 \u2211 (31)\nWhere\n1\ni n\nix i\nF =\n= \u2211\nis the non-uniform preload lined with x and i is the number of the force;\n1\ni n\niy i\nM =\n= \u2211\nis the bending moment lined with y which caused by the non-uniform preload\n1\ni n\niz i\nM =\n= \u2211\nis the bending moment lined with z axis which caused by the non-uniform preload\nWhen bearing has rotating inner raceway\n( ) '1 cos 1 cos m i\ni o\n\u03c9 \u03b3 \u03b1 \u03c9 \u03b1 \u03b1 \u2212 = + \u2212 (32)\n2 sinmR gj\nj j\nM J \u03c9\u03c9 \u03c9 \u03b2 \u03c9 \u03c9 \u239b \u239e\u239b \u239e= \u239c \u239f\u239c \u239f \u239d \u23a0 \u239d \u23a0 (33)\nx\u03b4 , y\u03b4 , z\u03b4 , y\u03b8 and z\u03b8 can be solved by iterative method until compatible values of the primary unknown quantities are obtained. Then the bearing stiffness matrix can be calculated as follows\n1\ni n\nix i\nx x\nF K\n\u03b4\n=\n== \u2211\n(34)\n1\ni n\niy i\ny y\nF K\n\u03b4\n=\n== \u2211\n(35)\n1\ni n\niz i\nz z\nF K\n\u03b4\n=\n== \u2211\n(36)\n1\ni n\niy i\ny y\nM K\u03b8 \u03b8\n=\n== \u2211\n(37)\n1\ni n\niz i\nz z\nM K\u03b8 \u03b8\n=\n== \u2211\n(38)\nIII. FINITE ELEMENT ANALYSIS MODEL Under the non-uniform preload, rolling bearings contact is a compound problem, combined with multi-body contact, friction, sliding, material nonlinearity, geometric nonlinearity. So using finite element method (FEM) to analyse bearing is one of the most efficient methods [9].\nA. Balancer and Self-balancing Spindle The ANSYS WORKBENCH\u00ae is used as the tool of FEM in this work. The bearing 3D model is built using Pro/E\u00ae, which is then imported into ANSYS. In this procedure, no model data is lost.\nSome bearing material characteristic parameters and structure parameters must be set in WORKBENCH according to the following Table .\nTABLE . BEARING MATERIAL CHARACTERISTIC PARAMETERS AND STRUCTURE PARAMETERS\nD (mm) fi fo Z dm (mm) \u03b1 ( \u00b0)\n9.525 0.52 0.52 18 77.5 15\nEball\n(106Pa) ball ball (Kg/mm3) Ering (106Pa) ring ring (Kg/mm3)\n3.2\u00d7105 0.25 3.2\u00d710-6 2.06\u00d7105 0.3 7.8\u00d710-6\nB. Contact Setting According to the Hertz contact theory, the contact form between the ball and the inner or outer rings is surface-surface. In static analysis, friction has little direct impact on the contact analysis, so the contact type is set as non-separation. This" ] }, { "image_filename": "designv11_101_0000358_amm.86.399-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000358_amm.86.399-Figure1-1.png", "caption": "Fig. 1 Schematic diagram of tooth surface Fig. 2 Coordinate change", "texts": [ " ( ) = ( + + )m\u03c6\u03c6 be sin\u03b1cos\u03c6 sin\u03b1sin\u03c6 cos\u03b1r i j k (1) When part of \u03c6 in equation (1) become ( )\u03c6+t , such as equation (2).Then, the starting point position of the formed conical logarithmic spiral changes, the other properties don\u2019t change and are as the other conical logarithmic spiral on cone . ( ) ( ) ( )( )= + +m\u03c6\u03c6+t be sin\u03b1cos \u03c6+t sin\u03b1sin \u03c6+t cos\u03b1r i j k (2) Based on space geometric modeling [2], the tooth surface construction of the logarithmic spiral bevel gear can be expressed by the tooth trace and the tooth profile curve, as shown in Figure 1.The simplified expression of tooth surface equations is equation (3). According to the theory of gear geometry [3], taking coordinate transformation theory tooth surface can be obtained, as shown in \u03c6r is radius vector of the conical logarithmic spiral at any point 1o , over point 1o we do its normal plane at this point. The intersection of this normal plane and z axis is 3o , 1 3o o = a r . Make involute by taking cosaar as radius of base circle, and Involute just passes 1o point, a is the pressure angle" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001652_ecj.10392-Figure8-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001652_ecj.10392-Figure8-1.png", "caption": "Fig. 8. Base field.", "texts": [ " To avoid self-collisions, link C must be moved in a direction different from x. The self-collision avoidance method proposed in this section finds appropriate directions in which links must be moved by defining base fields for individual manipulators. The base field Fk of manipulator Mk is a sphere of radius Rk whose center is located at Bk, the base position of Mk, and the potential function Vk(X) is defined at each point X in Fk so that Vk(X) decreases as the distance between X and Bk increases, as shown in Fig. 8. Based on the potential functions, the real-time part of the BFA calculates D(X) + V1(X) + V2(X) + \u22c5 \u22c5 \u22c5 + Vm\u22121(X) + Vm+1(X) + Vm+2(X) + \u22c5 \u22c5 \u22c5 + VM(X), as the distance at X, a movable end position of the n-th link, from initial position Sn. Here, it is assumed that the n-th link is a link of the m-th manipulator Mm, and that D(X) is the distance between Sn and X calculated by Dijkstra\u2019s shortest path algorithm. Then, the occurrence of collisions between manipulators Mm and Mk (m \u2260 k) can be decreased, because the distance between initial position Sn and point X becomes large if X is included in Fk, that is, the n-th link of Mm does not get close to Mk, as shown in Fig" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001317_amr.383-390.1524-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001317_amr.383-390.1524-Figure2-1.png", "caption": "Figure 2. Forces in top view", "texts": [ " Make the following assumptions [4] to study ground taxiing and automatic take-off characteristics of the UAV: Sample UAV is rigid,and its mass is contant; The earth axis is a inertial axis system; Treat the earth as a plane; The gravitational acceleration g is constant; All rights reserved. No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of TTP, www.ttp.net. (ID: 165.123.34.86, University of Pennsylvania Library, Philadelphia, United States of America-09/10/13,22:55:02) During ground taxiing, the forces apply to the UAV are aerodynamic forces, engine power, force of gravity, landing gear support forces, the friction forces and the lateral forces, as shown in Fig.1 and Fig.2. Compared with flight in the air, the six degree of freedom (6-DOF) motion is reduced to 3-DOF motion due to the ground restrictions. The following part will analyze these forces and moments in detail. Aerodynamic Forces In wind axis system, the total aerodynamic forces can be described as: [ ]T aR D Y L= \u2212 \u2212 (1) In body axis system, the total aerodynamic moments can be defined as: [ ]TR R M N= (2) D, Y, L is the drag force, the side force and the lift force respectively. R, M, N is the roll, pitch and yaw moment in the body axis" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001077_icl.2013.6644643-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001077_icl.2013.6644643-Figure2-1.png", "caption": "Fig. 2. Replica of a remote system for flatness evaluation.", "texts": [ " In 2003 the potential of AR in a car door assembly was reported in [19], while in 2012 AR was considered as relevant in complex assembly tasks [20]. The BMW's Augmented Reality Glasses video is an example of a new technique in which the mechanics sees the specific information he needs and can proceed step-bystep, http://www.youtube.com/watch?v=Y5ywMb6SeGc. Finally, mixed reality is a type of virtual reality combining real and virtual images. The user is immersed in a 3D space generated by the computer. In Fig. 2, the user will be able to observe from any angle, to manipulate and to interact with the virtual replica of that remote equipment, to perform a test and even to get by email the test results for later evaluation. The flatness evaluation is a surface property with interest either in mechanical engineering education or for lifelong learning activities. Another example for STEM is presented in Fig. 3, by getting immersed in a common electrical circuit for evaluating the on/off combination of on-off switches" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000024_2013.40604-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000024_2013.40604-Figure1-1.png", "caption": "Figure 1. Stryker vehicle geometry (units: meters)", "texts": [ " Differences between the off-road mobility power and duty cycle requirements for the Stryker and AAV maneuvers during the missions were identified. The variability of the mobility power characteristics from the two sets of GPS data was determined. The effect of varying the soil's cone index on the motion resistance and net mobility power of the Stryker vehicle was quantified. The vehicle analyzed was the Stryker Infantry Carrier Vehicle (ICV), and it is an 8-wheeled, 17,237 kg vehicle that is powered by a 261 kW V-8 diesel engine. General dimensions of the vehicle are shown in Figure 1. The maximum travel speed of the vehicle is 27 m/s. The vehicle is either 4 or 8-wheel drive; during maneuvers, the vehicle was operated in the 4-wheel drive mode. The frontal area, fA , was determined to be approximately 4.5 m2 based on the geometry of the Stryker. The vehicle is equipped with a Central Tire Inflation System (CTIS) that allows the operator to vary the inflation pressure of all tires simultaneously according to the terrain conditions (Ayers et al., 2009). All wheels were equipped with Michelin X tires" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000823_amr.221.165-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000823_amr.221.165-Figure1-1.png", "caption": "Fig. 1 Schematic representation of nodes and elements in one quarter section of a tire", "texts": [ " The well-known FEM software ANSYS was used to analyze the tires, thus guaranteeing that the method and simulation process are believable. All rights reserved. No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of Trans Tech Publications, www.ttp.net. (ID: 152.14.136.77, NCSU North Carolina State University, Raleigh, USA-12/05/15,20:50:39) To build a suitable FE radial tire model, a quarter tire structure was first drawn using AutoCAD software. Then the drawing was divided into gridding as shown in Fig. 1. An interface program was written in AutoLISP code to transfer the coordinate data of all the intersection points into a file, which can be read directly by ANSYS. Then the node coordinates in ANSYS were obtained from the data. A second row of coordinates was created by rotating all the nodes around the tire axis. The elements constructed by the two rows of nodes make up a sector slice of the FE tire model. A half tire model can be created by rotating the sector slice in the direction of the roll. In most cases the half tire model was used for analysis because the tire is eudipleural. If necessary, the whole tire model can be obtained by reflecting the half one. A passenger car radial tire, 205/60R15, was selected as the control tire. The black numbers in Fig. 1 represent node marks, and the green numbers in circles denote element marks. The red lines represent the steel wire belts, while the green lines are nylon cord fabrics. The dark blue diamond at bottom is the steel wire circle. 2.1 FE model of a free tire A tire possesses both axial and facial symmetry in the condition of non-touching ground; therefore, one quarter of the tire section layer can be taken as the research object. The FE model of the free tire sector slice is shown in Fig. 2. The constraints and loads are as follows: \u2022 Symmetric constraints in the tire circle (represented by small cyan crosses on the nodes in Fig" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002179_tia.2013.2270451-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002179_tia.2013.2270451-Figure3-1.png", "caption": "Fig. 3. Vectors of stator flux, rotor flux, stator current, and stator voltage of IM.", "texts": [ " The DTC can only provide effective torque control and estimation of motor parameters, and it needs a speed controller to infer the reference value of the controlled torque. The DTC is implemented by a digital signal processor combined with the ASIC through a fiber-optic cable to speed up the processing speed. The fundamental theory of the DTC is based on fieldoriented control [5] and DSC [6]. Both control schemes employ a space vector to express all quantities based on a stationary reference frame of the IM. The vectors of stator flux, rotor flux, stator current, and stator voltage of an IM are shown in Fig. 3 [7]. The electric torque, which is the product of both stator flux and stator current or the product of both stator flux and rotor flux, can be expressed as below [8], [9], i.e., Te = P 1\u2212 \u03c3 \u03c3Lm \u03d5s \u00b7 j\u03d5r = 3 2 Pis \u00b7 j\u03d5s (1) where P is the pole pair, \u03c3 is the leakage factor, Lm is the magnetization inductance, \u03d5s is the stator-flux vector, \u03d5r is the rotor-flux vector, and is is the stator-current vector of the IM. Fig. 4 shows the operating performance of the hysteresis controller, whereas the value of Te is limited to between (Tref + \u0394T1) and (Tref \u2212\u0394T1) by means of selecting one of the statorvoltage vectors" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003240_978-0-8176-8370-2-Figure9.9-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003240_978-0-8176-8370-2-Figure9.9-1.png", "caption": "Figure 9.9: The Poincar\u00e9 variables in the ecliptic plane.", "texts": [ " Clearly \u03c7ecc = \u03b7ecc = 0 correspond to circular orbits, while \u03c7inc = \u03b7inc = 0 describe orbits in the ecliptic plane. We immediately obtain G = \u039b\u2212 1 2 (\u03c72ecc + \u03b72ecc), G3 = G \u2212 1 2 (\u03c72inc + \u03b72inc). (9.2.3) We would express the position vector of a point moving along a Keplerian ellipse as a function of the Poincar\u00e9 variables, but this is not possible in a closed form: the inversion of the modified Kepler equation is required (see below). Fortunately, it is sufficient for what follows to use \u03c3 as parameter, instead of \u03bb. To this end, we proceed with the following geometric construction. Let us start (see Figure 9.9) from a circular orbit of radius a contained in the ecliptic plane q1q2 and let \u03c3 be the anomaly of the moving point P. Accordingly, \u2212\u2192 FP = (a cos\u03c3, a sin\u03c3, 0) and \u2225\u2225\u2225 \u2212\u2192FP\u2225\u2225\u2225 = a. Take an angle \u03b2, related to the numerical eccentricity E through sin\u03b2 = E, 0 \u2264 \u03b2 < \u03c0/2, and rotate about the unit vector \u2212\u2192u having the same direction as \u2212\u2192 E , then project on the ecliptic plane q1q2 and finally translate by the vector \u2212a \u2212\u2192E . Consequently, we are interested in \u2212\u2192E and in R( \u2212\u2192u,\u03b2), the 3\u00d73 288 The Multi-Body Gravitational Problem real matrix corresponding to the above rotation" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000542_s40430-013-0117-8-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000542_s40430-013-0117-8-Figure1-1.png", "caption": "Fig. 1 Shell description and basic kinematical quantities", "texts": [ " It consists of a geometrically exact formulation in which first-order shear deformation due to bending is taken into account (i.e., Reissner\u2013Mindlin kinematics), while the thickness is supposed to remain constant during the deformation (for a deformable-thickness version, see [11]). We assume a flat reference configuration for the shell mid- surface at the outset. A local orthogonal system er 1; e r 2; e r 3 with corresponding coordinates n1; n2; ff g is defined in this configuration, with vectors er a (a \u00bc 1; 2) placed on the shell middle plane and er 3 orthogonal to it (see Fig. 1). Points in this configuration are described by the vector field n \u00bc f\u00fe ar; \u00f01\u00de where f \u00bc naer a describes the position of points at the middle surface and ar \u00bc fer 3 defines the position of points at the cross section relative to the middle surface (ar is also called the shell reference director). Notice that f 2 H \u00bc hb; ht is the thickness coordinate, with h \u00bc hb \u00fe ht being the shell reference thickness. In the current configuration, another local orthonormal system e1; e2; e3f g is defined, as depicted in Fig. 1. The shell deformation is then described by a vector field x such that the position of the material points is expressed by x \u00bc z\u00fe a; \u00f02\u00de where z \u00bc z\u0302 na\u00f0 \u00de describes the position of points at the deformed middle surface and a \u00bc a\u0302 na; f\u00f0 \u00de defines the position of points at the deformed cross section with respect to the deformed middle surface (a is also called the shell current director). From Fig. 1, one finds that z \u00bc f\u00fe u, where u is the displacement vector of points of the middle surface. Vector a, in its turn, is obtained by a \u00bc Qar, in which Q is the rotation tensor of the director. Accordingly, no thickness change is allowed but first-order shear deformations are accounted for, since a is not necessarily normal to the deformed middle surface. Relation ei \u00bc Qer i holds for the local systems. The rotation tensor Q may be written in terms of the Euler rotation vector h by means of the well-known Euler\u2013 Rodrigues formula Q \u00bc I\u00fe sin h h H\u00fe 1 2 sin2\u00f0h=2\u00de \u00f0h=2\u00de2 H2; \u00f03\u00de in which h \u00bc hk k is the rotation angle of the director and H \u00bc Skew h\u00f0 \u00de is the skew-symmetric tensor whose axial vector is h" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002962_gt2013-94951-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002962_gt2013-94951-Figure1-1.png", "caption": "Figure 1. An Air Foil Bearing", "texts": [ " It is known that the speed of the AFB\u2019s ranges from 60000rpm to 150000rpm and makes the AFB suitable for high speed turbomachinery applications, especially in microturbines, whereas the main disadvantage of the AFBs is the relatively low load carrying ability. In order to increase the load capacity, better top and bump foil design should be implemented. Finally, another major characteristic of the AFBs is the lift off speed. The design procedure should be oriented in order to allow the lift off of the journal in lower rotational speed. As shown in figure 1, a typical AFB mainly consists of three parts, the housing, the bump foil and the top foil. Top foil is supported by the compliant bump foil and both of them are pinned with one of their free ends to the housing of the bearing. This setup makes the total structure compliant. As mentioned earlier, the deflection of the top and bump foil are strongly depended upon the gas film pressure magnitude and distribution and vice versa. Top foil is a thin layer of metal the role of which is to create a cylindrical geometry at the area which contains the journal" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002239_978-94-007-5006-7_3-Figure3.12-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002239_978-94-007-5006-7_3-Figure3.12-1.png", "caption": "Fig. 3.12 Representation of a spherical joint using YXZ EAJs", "texts": [ " On the other hand, YXZ EAJs is preferred if asymmetric set is chosen for the rotation representation of a 3-DOF joints. The spherical joint shown in Fig 3.3 connects the reference link #R with the moving link #M. In practice, a robotic system may have serial- or tree-type architecture with several multiple-DOF joints. This calls for a systematic numbering scheme for intersecting revolute joints and the associated imaginary or physical links. So, the use of Euler-Angle-Joints (EAJs) and the associated numbering scheme are presented here for the systematic representation of a spherical joint. For example, Fig. 3.12 shows a link, #(k 1), coupled to its neighboring link, #k, by a spherical joint, k, which has three rotational DOF. In order to represent the spherical joint using YXZ EAJs, links #(k 1) and #k are considered to be connected by three orthogonally placed revolute joints, denoted as k1, k2, and k3, connecting two virtual links (#k1 and #k2), each of zero length and mass. The link #k3 is the actual physical link #k, which is attached to the third revolute joint, i.e., k3. The corresponding frame assignment using the DH notation is shown in Fig. 3.12. The YXZ scheme of EAJs is chosen and will be used throughout this book because (1) it does not require any fixed rotation at the beginning or at the end, and (2) unlike ZYZ scheme of EAJs, it does not lead to a singularity in zeroconfiguration. Zero-configuration is defined here as the one where all the joint angles are set to zeros. Any three-parameter description of rotations including the Euler angles suffers from singularity. Singularity is encountered in EAJs when two joint axes become parallel" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001455_2012-36-0254-Figure12-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001455_2012-36-0254-Figure12-1.png", "caption": "Figure 12. Shaft displacements in cylindrical coordinates [mm]. 5 bearings a) radial and b)angular, 3 bearings c) radial and d)angular", "texts": [ " Also, according to the figure, bearing J2 for 5 bearings configuration presents no contact pressure because it is unloaded; this can also be seen in figure 11, that presents load in polar coordinates. Loading for each bearing (J1, J3 and J5) is higher when the system has 3 bearings, though, the magnitude of maximum load is maintained. Finite element analysis (FEA) has been used in association with analytical calculations (Bearinx). Shaft displacements, bearings loads and contact pressures have shown good agreement. Shat displacements in cylindrical coordinates (figure 12) for load case 100, performed with FEA, show that the radial amplitudes are higher for 3 bearings configuration than for 5 bearings layout, which means that the configuration with 3 bearings facilitates the bending motion. It can also be seen that the displacement is composed by bending and torsion and, the torsion predominates. a) 5 needle bearings, radial displacement b) 5 needle bearings, angular displacement Needle roller bearings results so far encourage its application in camshaft system in terms of structural functionality" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001605_iemdc.2011.5994935-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001605_iemdc.2011.5994935-Figure3-1.png", "caption": "Fig. 3. Permeability distribution of the core (f = 333 Hz, 190 V).", "texts": [ " Therefore, Wi Measured can be calculated from the losses obtained by the FEM as follows: NoloadHNoloadiMeasuredi WWW 2+= (5) where Wi Nolad and W2H Nolad are the total loss generated in the core and the harmonic secondary cage loss at the no-load condition. Substituting (5) into (4) yields: )()( 22 NoloadHLoadHNoloadiLoadis WWWWW \u2212+\u2212= . (6) Therefore, Ws can be calculated by the FEM from the increases in Wi and W2H with the load. One reason of these loss increases with the load is the magnetic saturation at the rotor surface. Fig. 3 shows the variation in the permeability distribution with the load. It is observed that the permeability of the rotor core on the top of the bars significantly decreases with the load. It is caused by an increase in the flux produced by the secondary currents. As a consequence, the slot harmonics increase, as shown in Fig. 2. Fig. 4 shows the calculated stray load loss Ws versus square of the bar current Ib 2. Ws is nearly proportionate to Ib 2. From the results and the discussions in the previous section, we have proposed a novel equivalent circuit of induction motors in order to take into account the harmonic losses and torques" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002110_00207721.2011.618642-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002110_00207721.2011.618642-Figure2-1.png", "caption": "Figure 2. Coordinate system for ship steering yaw-motion control.", "texts": [ " Step 1: Choose C to satisfy CDp \u00bc 0 and CBp 6\u00bc 0. Step 2: From the main theorem, get the feasible solution of control feedback gain G as (49). Step 3: Set the switching function S\u00f0t\u00de as (14). Step 4: The controller u\u00f0t\u00de is obtained from (21). In this section, a practical example considering the ship yaw-motion systems adopted from (Astrom 1980) and Kallasrom (1981) is demonstrated to verify the proposed method. The dynamic equations of the ship yawmotion system are described by using a coordinate frame fixed to the ship (Figure 2). The motion of ship is considered like a tanker, therefore, there is little coupling between different modes of motion and the steering dynamics can be described by considering modes of the surge, sway and yaw motions, separately. The dynamic equations of motion can be represented as (Astrom 1980) m\u0302 _\u0302u\u00f0t\u00de v\u0302\u00f0t\u00der\u0302\u00f0t\u00de x\u0302Gr\u0302 2\u00f0t\u00de h i \u00bc X\u0302, \u00f050a\u00de m\u0302 _\u0302v\u00f0t\u00de u\u0302\u00f0t\u00der\u0302\u00f0t\u00de \u00fe x\u0302Gr\u0302 2\u00f0t\u00de h i \u00bc Y\u0302, \u00f050b\u00de I\u0302z _r\u00f0t\u00de \u00fe m\u0302x\u0302G _\u0302v\u00f0t\u00de \u00fe r\u0302\u00f0t\u00deu\u0302\u00f0t\u00de h i \u00bc N\u0302, \u00f050c\u00de where m\u0302 is the mass of the ship with I\u0302z being its moment of inertia with respect to the z-axis and x\u0302G being the x-coordinate of the centre of mass" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002194_978-3-642-25899-2_57-Figure4-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002194_978-3-642-25899-2_57-Figure4-1.png", "caption": "Fig. 4. Velocity and turbulence of radial grooves", "texts": [ "0s, heat is absorbed by the steel plates and friction discs. A negligible amount of heat is transferred to cooling oil. The effect of oil cooling heat dissipation is mainly reflected after the clutch engagement. On the city traffic, the clutch engagement is ordinary. The temperature of friction discs was easily raised, so the efficiency of oil heat dissipation was very important. The heat transfer between friction discs and flow belongs to convection heat transfer. The efficiency is mainly affected by velocity and turbulence. Fig.4 shows the velocity of radial grooves. As we can see, with the increase of the radius, the velocity of flow also increases gradually. The velocity distribution between oil grooves is consistent. So the coefficient of heat transfer between grooves would not the velocity. This was good for the uniform heat dissipation of frictions.Fig.4 shows the turbulent of radial grooves. Serious turbulence appears at the entrance of flow. This is caused by the shock between flow and the walls. Turbulence is not good for heat dissipation. So the entrance is needed to improve. Moreover, the turbulence also appears at the export of grooves. This is because that the outlet is so small and the velocity is too high. So the shape is needed to optimize. Fig.5 shows the velocity of parallel grooves. It can be seen from the figure, the fluid velocity distribution of left and right sides vary strongly, the maximum speed of the left 17" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001382_scis-isis.2012.6505215-Figure8-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001382_scis-isis.2012.6505215-Figure8-1.png", "caption": "Fig. 8: Estimation of the door states", "texts": [ " Therefore, the system check the initial data and find the missing data. In cases where the missing data are found, the data are checked the door space or not by using the length of the missing data. Then, the system associates the data of current door position data (red line in Fig. 6) with the initial door position data (dotted line in Fig. 6). Next, the estimation of the door states are needed for a pair of the door data. For the state estimation, the system compares a pair of door data and checks a length of the straight line at the door\u2019s flat surface (Fig. 8). In case when the door is closed, the door is unified with the wall, and the length of the consecutive line (the length of line B in Fig. 8) is longer than the opened door (the length of line A in Fig. 8). Therefore, the door data which has the longer straight line is estimated as the \u201cClose\u201d. Fig. 9 shows the experimental environment. The door at the left corner of the room (the circled area in Fig. 9) is recognized. In this experiment, the door is closed at the initial state and the human opens the door. Then, it is confirmed that the current door position (the opened door) can be extracted and the initial position of the door (the closed door) can be estimated. Fig. 10 and Fig. 11 show the results" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002887_physreve.86.011927-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002887_physreve.86.011927-Figure2-1.png", "caption": "FIG. 2. (Color online) Tactical response for the Smith-Saldana model. Solutions are shown for (a) and (c) rs = 0.9 and (b) and (d) rs = 1.1, with (\u03ba1,\u03ba2) = (0.005,\u22120.005). (a) and (b) Time series of the Ca2+ concentration for y = 1/2. (c) and (d) Black and gray lines indicate z at the S and F sites, respectively. Dotted lines indicate c at the S site.", "texts": [], "surrounding_texts": [ "decreases concomitantly. (ii) The sol is transported from the rear part to the leading edge by the pressure gradient generated in step (i). (iii) When the thickness of the sol exceeds a threshold level, the organism begins to migrate.\nIn light of step (i), the Ca2+ concentration is a critical factor in deciding the direction that the organism migrates. In this article we focus on steps (i) and (ii). In particular, we investigate how the frequency of the chemical oscillation controls the concentration of Ca2+.\nOur model is based on that reported previously in Refs. [19,21] and is illustrated in Fig. 1. Our model system comprises two parts: an outer gel composed of actomyosin filaments surrounding an inner flowing sol. The streaming of the protoplasmic sol is driven by a pressure gradient generated by contraction of the actomyosin filaments in the gel. The equations for the dynamics of the sol and gel are the same as those derived in Ref. [21]:\nst = \u2207 \u00b7 (sM\u2207p) + \u03b3 (z \u2212 s), (1)\nzt = Dz\u22072z + \u03b3 (s \u2212 z), (2)\nx = (x,y) \u2208 [0,1] \u00d7 [0,1], t > 0,\nwhere \u2207 = (\u2202/\u2202x,\u2202/\u2202y); s(x,t) is the thickness (or density) of the sol; z(x,t) is the thickness of the gel; and M , \u03b3 , and Dz are coefficients. The sol flows according to Darcy\u2019s law and its conversion to the gel is described by \u03b3 (z \u2212 s), which reduces the contrast with the gel thickness. The gel diffuses slowly and its conversion to the sol is described by \u03b3 (s \u2212 z). The total mass of the organism (s + z) is conserved over the entire space.\nThe basal thickness of the sol for the organism is described by sb(x,t), which is the volume capacity of the cavity bounded by the gel cortex. We assume that the basal thickness acts like the natural length of a Hookean spring and thus the elastic force is governed by the displacement of contraction from the basal thickness sb. Thus the pressure of the sol is governed by p = \u03b2(s \u2212 sb).\nThe basal thickness sb varies periodically in time due to the actomyosin contraction. The periodic contraction is generated by periodic variations of chemical concentration. It has been\nexperimentally observed that the contraction takes place at low concentrations of Ca2+ oscillation [5]. In addition, the amplitude of the contraction increases as the thickness of the gel (z) increases. Thus we assume that the dynamics of sb(x,t) are described as follows:\nsb = s\u0304 + az(c \u2212 c\u0304), (3)\nwhere c = c(x,t) is the concentration of Ca2+, s\u0304(x,t) is the mean thickness of the sol layer, and a and c\u0304 are constants. The time-averaged thickness s\u0304 is described by s\u0304t = \u03b3s(s \u2212 s\u0304), where \u03b3s is a constant.\nEquations (1) and (2) indicate that z increases with time due to the sol-gel transformation. Therefore, the tactical responses shown in the experiments might be established by disrupting the balance between the inflow and outflow rates of the sol in the stimulated area. Our hypothesis is that the modulation of the chemical oscillator (the actomyosin contraction) by the accumulation of the sol (rheological deformation of the organism) plays an important role in disrupting this balance. This hypothesis is based on the finding that the presence of the interaction between the chemical period and the amplitude of contraction oscillation is modulated locally at the place where the sol gathers and scatters [22].\nIn Refs. [19,21] actomyosin contraction generated by chemical oscillations of Ca2+, ATP, and so on was simply described by sin(\u03c9t), which is the independent function of x. Here we employ a more realistic model equation, namely, the Smith-Saldana (SS) model [23,24], based on a biochemical analysis of the time period of Ca2+ fluctuations. We include the diffusion of Ca2+ in the gel layer and the chemical oscillator is assumed to be affected by the amount of sol as follows:\nct = f (c,\u03d5)/\u03c4 + \u03ba1s + Dc c, \u03d5t = g(c,\u03d5)/\u03c4 + \u03ba2s, (4)\nwhere = \u22022/\u2202x2 + \u22022/\u2202y2, c(x,t) is the concentration of Ca2+, \u03d5(x,t) is the probability of a myosin lightchain kinase being phosphorylated, f (c,\u03d5) = \u2212kV nc(c,\u03d5) + kL(Nc \u2212 c), and g(c,\u03d5) = KQ(nc(c,\u03d5))(1 \u2212 \u03d5) \u2212 kE\u03d5, with KQ(nc) = kQ{K\u2217nc(c,\u03d5)/[1 + K\u2217nc(c,\u03d5)]} and nc(c,\u03d5) = [1 + 0.17(c \u2212 7.5)](a0 + a1\u03d5 + a2\u03d5 2 + a3\u03d5 3). The same parameter values as those in Ref. [24] are set: a0 = 0.349 353, a1 = \u22120.045 456 7, a2 = 1.15905, a3 = 1.823 858, kV = 0.12, kL = 0.004, kQ = 1.0, kE = 0.1, K\u2217 = 1.75, and Nc = 25.0. The first terms of the c and \u03d5 equations generate a limit cycle and \u03ba1 and \u03ba2 are the coupling coefficients between the amount of the sol and the chemical oscillator. A weak coupling is assumed such that the values of |\u03baj | are small. Equation (4) is used for actomyosin contraction unless otherwise stated.\nThe length of the plasmodium and the period of oscillation are nondimensionalized by the characteristic length and period of actomyosin contraction in a fashion similar to that reported in Ref. [19]. The pressure p and stiffness \u03b2 are nondimensionalized by the typical pressure estimated at approximately 1.5 \u00d7 104 N/m2. The parameters are set as a = 0.1, \u03b2 = 20.0, \u03b3s = 10.0, Dc = 2.0 \u00d7 10\u22123, Dz = 0.01, and c\u0304 = 8.7. The Neumann boundary condition is imposed. The time increment is t = 0.1 and the grid size is 50 \u00d7 50.\nThe experiments conducted by Matsumoto et al. [1] showed that the tactical response could be controlled by the local modulation of the frequency of actomyosin contractions; specifically, the organism was attracted to and repelled from\n011927-2", "the positions where the frequency was higher and lower, respectively. In numerical simulations, the frequency could be controlled by changing \u03c4 in Eq. (4). Thus the frequency control is represented by taking \u03c4 as a function of space variables:\n\u03c4 (x) = { \u03c40rs, x \u2208 s, t tp\n\u03c40, otherwise, (5)\nwhere \u03c40 = 0.1, s is the position where the frequency is controlled, and tp is the time when perturbation is first applied. We set tp = 2.0 and s : (x,y) \u2208 [0,\u03b4] \u00d7 [(1 \u2212 \u03b4)/2,(1 + \u03b4)/2]}, where \u03b4 = 0.2. As \u03c4 corresponds to the time constant of the kinetics part of Eq. (4), a small rs means a high frequency at s . For convenience, we refer to s as the stimulated site (S site) and define the free site (F site) f : (x,y) \u2208 [1 \u2212 \u03b4,1] \u00d7 [(1 \u2212 \u03b4)/2,(1 + \u03b4)/2] as a reference (Fig. 1).\nIII. RESULTS\nWe now consider how the interaction between the chemical oscillator and the sol amount affects the migration direction. Periodic change in the thickness of the organism due to periodic shuttle streaming is observed when rs = 1, irrespective of whether the interaction is absent or present. When rs < 1 (rs > 1), the phase of the contraction is advanced (delayed) and the amplitude of the thickness oscillation at the S site becomes larger than that at the F site [Figs. 2(a)\u20132(d)]. In response to the change in frequency of the contraction, the time-averaged thickness of the organism (or the amount of the sol) either decreases or increases depending on \u03ba1 and \u03ba2.\nBy taking \u03ba1 and \u03ba2 as control parameters, the balance between the inflow and outflow rates of the amount of the\nsol at the S site can be changed. Figure 3 shows a phase diagram of the time-averaged thickness at the S site, where the time-averaged thickness is defined as Zi = \u222b ti+1\nti z dt/T\nand ti denotes the time when u attains the local maximum. The parameters \u03ba1 and \u03ba2 are changed in the region (\u03ba1,\u03ba2) \u2208 [\u22120.005,0.005] \u00d7 [\u22120.005,0.005]. The changes in (\u03ba1,\u03ba2) from the top left-hand corner to the bottom right-hand corner in the diagram with rs = 0.9 and rs = 1.1 cause gradual increases and decreases, respectively, in the averaged thickness Z8 [Fig. 3]. For both cases, rs = 0.9 and 1.1, Z8 show relatively few changes when the interaction between the chemical oscillator and the sol is absent, that is, (\u03ba1,\u03ba2) = (0,0). These numerical results indicate the following: (i) The interaction is crucial to break the balance of the sol flow rates at the S site and (ii) the actual responses observed by Matsumoto\n011927-3", "et al. [1] are reproduced when \u03ba1 and \u03ba2 are taken from the bottom right-hand corner in the diagram. Hereinafter, we assume (\u03ba1,\u03ba2) = (0.005,\u22120.005) in order to investigate the mechanism.\nAs shown in the experiments [1], the baseline of z at the S site increases and decreases as the phase of the chemical oscillator is advanced and delayed, respectively [dashed lines in Figs. 2(c) and 2(d); see also Figs. 4(a) and 4(b)]. In addition, the baseline of c also increases and decreases for t > ts when rs = 0.9 and 1.1, respectively. Such responses of c are crucial for the change in Zi . According to the p and sb equations, s increases (decreases) as the time-averaged value of c over the unit time interval increases (decreases), suggesting that z increases (decreases) through the sol-gel conversion [see Eqs. (1) and (2)]. We should therefore consider the mechanism by which the baseline of c increases and decreases through the interaction of the dynamics of the sol for rs < 1 and rs > 1, respectively.\nThe influence of the interaction on the dynamics of c can be clearly shown by drawing the solution orbit of c and \u03d5. In the presence of the interaction, the position of the circle of c in the phase plane moves to the right for rs = 0.9 [gray arrows in Fig. 4(a)]. As a result, the baseline of c increases as the time increases. By contrast, the position of the circle of c moves to the left for rs = 1.1 and the baseline of c at the S site decreases [gray arrows in Fig. 4(b)]. Thus we can find\nSmith-Saldana model\nresonance between the periodic oscillation of (c,\u03d5) and the rhythmic sol flow, generated by the spatial graduation of the phase of the actomyosin contraction around the S site.\nIt should be noted that the periodic motion of (c,\u03d5) and the interaction of the sol are not sufficient to account for the tactical responses. One representative model showing periodic oscillations is the Stuart-Landau (SL) equation, which was used in Refs. [19,25]: f (c,\u03d5) = c \u2212 \u03d5 \u2212 c(c2 + \u03d52) and g(c,\u03d5) = c + \u03d5 \u2212 \u03d5(c2 + \u03d52). We now compare the dynamics of the SS model and the SL equation. To ensure that the experimental conditions (such as the oscillation frequency and the minimum and maximum p) are as similar as possible to those of the SS model, the parameters \u03c4 , a, and c\u0304 are set for the SL system such that \u03c40 = 0.2336, a = 0.111, and c\u0304 = 0.70 and the parameters \u03ba1 and \u03ba2 are changed in the region (\u03ba1,\u03ba2) \u2208 [\u22120.005,0.005] \u00d7 [\u22120.005,0.005].\nIt was found that Z8 increases as the frequency of the S site increases for rs \u2208 [0.9,1.1]. However, the difference in Z8 between rs = 0.9 and 1.1 is much smaller than that for the SS model (Z8 = 1.06 for rs = 0.9 and Z8 = 0.96 for rs = 1.1) because the resonance does not occur; the increasing and decreasing rates of the thickness saturate quickly in comparison with the SS model (Fig. 5). These numerical results do not capture the actual observation.\nThe solution orbit of the SL equation in the c-\u03d5 plane for (\u03ba1,\u03ba2) = (0.005,\u22120.005) is shown in Figs. 4(c) and 4(d). For rs = 0.9 (rs = 1.1), the shape of the limit cycles becomes elliptical; the top right-hand corner and the bottom left-hand corner expand (shrink) in the direction of the arrows as shown in Figs. 4(c) and 4(d), where (\u03ba1,\u03ba2) = (0.005,\u22120.005) for rs = 0.9 and 1.1, respectively. However, in contrast to the case of SS, the position of the rotating center shows negligible changes, as does the mean thickness. This suggests that the resonance, which can change the position of the rotating center of c, is required for the mechanism.\nIV. DISCUSSION\nThe effect of the interaction between the fluid and chemical reactions has been studied, for example, in Refs. [15,16]. In the case of Physarum, the numerical results in Fig. 5 suggested that the existence of the interaction is not a sufficient condition to reproduce the actual observation shown by Matsumoto et al. [1]. The mathematical model presented here showed that the Ca2+ concentration change due to the resonance through the interaction between the rhythmic streaming of the sol and the chemical oscillators is a key mechanism for tactical responses. We remark that such a cooperative response between z and c has also been observed experimentally [26].\nStich and Mikhailov have studied the dispersion relation of the complex Ginzburg-Landau (GL) equation [6]:\nut = \u22072(u \u2212 \u03b2\u0303v) + (u + \u03c9\u0303v) \u2212 (u \u2212 \u03b1\u0303v)(u2 + v2), (6) vt = \u22072(\u03b2\u0303u + v) + (\u2212\u03c9\u0303u + v) \u2212 (\u03b1\u0303u + v)(u2 + v2),\nwhere \u03c9\u0303(x,y) \u2261 + \u03b4\u03c9(x,y) and is a constant. The heterogeneity function \u03b4\u03c9(x,y) is defined as follows:\n\u03b4\u03c9(x,y) = { \u03c9, x \u2208 s\n0, otherwise. (7)\n011927-4" ] }, { "image_filename": "designv11_101_0003424_pesa.2013.6828258-Figure4-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003424_pesa.2013.6828258-Figure4-1.png", "caption": "Fig. 4: The cross section of a coaxial cable", "texts": [ " If there is only excitation in the i th stranded winding, the voltage of winding i can be obtained from (21) as follows )( )()( )( 1 d iw iwiw p ww iw I CjG Ajd SN pal U c W W )( )()( )( )( )( )( 1 ReIm iw iwiw iw iw iw iw I CjG I I dA S l jI I dA S l ii W W WW (30) The total voltage of the winding i is: W W W W N j jw jw N j jw jw ioiw I I dA S l jI I dA S l EU ii 1 )( )(1 )( )( )()( ReIm )( )()( 1 iw iwiw I CjG )( )()(1 )()( 1 )()()( 1 iw iwiw N j jwjjw N j jwijwio I CjG ILjIRE (31) From (31) a precise equivalent circuit can be obtained. Its equivalent circuit is shown in Fig. 3. Because the dc conductance Gw(i) and capacitance Cw(i) can be computed easily. Rw(ii) and Lw(ii) can also be isolated if necessary. The computation of the impedance between the two conductors of a coaxial cable as shown in Fig. 4 is used as an example. The space between the cylindrical conductors is filled with three layers of dielectrics. r1 = 11.9 and r2 = 2.25. The dimensions are : d1 = 6.858 mm, d2 = 7.874 mm, d3 = 19.177 mm, d4 = 25.396 mm, and d5 = 25.65 mm. The computed results of resistances and inductances versus frequency are shown in Table I, which shows that the solutions using the proposed method are the same as those obtained by traditional method. The analytical solution of the dc resistance is 0.0021599W/m, which is the same as that obtained by the numerical method at low frequency; the analytical solution of the inductance between the two conductors when the currents are only located on the surfaces of the conductors is 0" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002250_978-3-7091-1768-2_2-Figure8-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002250_978-3-7091-1768-2_2-Figure8-1.png", "caption": "Fig. 8. The Shear Compression Disk. The specimen consists of a disk into which a pair of grooves are machined on each opposite surface at a given angle (b). Figure (a) describes the initial pre-load step to achieve confinement through insertion to depth D into a hard steel die. Figure (b) describes the actual experimental step for which pressure is applied only to the central part of the disk. Figure (c) shows details of the specimen, spacer and confining steel die.", "texts": [ " In that respect, it is felt that the present technique can be considered as a useful extension of the thick elastic jacket confinement which puts virtually no limitations on the confining pressure, albeit variable during the test. The shear compression disk (SCD) was designed to allow for investigations of the influence of confining pressure and stress triaxiality on the mechanical response of various materials (Dorogoy et al., 2011; Karp et al., 2013). The specimen and experimental setup are shown in Fig. 8. The basic idea is quite simple and consists of confining a disk-like specimen by means of a thick steel die in which the conical geometry of the disk and the die create a state of confinement that is directly proportional to the depth of insertion of the disk. The latter has a pair of circular grooves of slightly different diameters on its upper and lower face respectively. Therefore, the first step of a test (preset) consists of inserting the disk to a predetermined depth in its conical guide to create a predefined state of hydrostatic stress (and triaxiality)" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002962_gt2013-94951-Figure17-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002962_gt2013-94951-Figure17-1.png", "caption": "Figure 17.Von Misses Stresses for eccentricity ratio 0.6", "texts": [ " The curve is a 5 th degree polynomial which is derived by interpolation of the discrete points for each eccentricity. The 5 th degree polynomial that describes the fitting of Figure 15 is given by equation (22). 33.961.575.1946.2574.2352.104 0 2 0 3 0 4 0 5 0 totf (N) (22) Figure 16 illustrates the contour plot of the pressure for 0.6 eccentricity ratio. Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 05/04/2014 Terms of Use: http://asme.org/terms 8 Copyright \u00a9 2013 by ASME As for the structural problem, figure 17 illustrates the contour plot of the Von Misses stress developed in the top foil. As seen from Figure 18, as the eccentricity increases, the velocity vector along the circumferential direction increases in magnitude due to larger top foil deflection. This phenomenon affects the load carrying capacity by increasing it. Changing the bump foil and the top foil parameters of the AFB has a significant effect on the structural stiffness of the bearing. Simulations are made for bump stiffness being 1000 times lower than the initial stiffness" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003001_detc2013-12231-Figure8-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003001_detc2013-12231-Figure8-1.png", "caption": "Fig. 8 The instantaneous contact load distribution", "texts": [ "org/about-asme/terms-of-use The geometry accuracy level of the beveloid gear pair is assumed to be level 7 and the assembly errors are designed as follows, 1 2,e e 0.029mm\uff0c 3e 0.025\u00b0. The quasi-static loaded tooth contact analysis for the intersected gear model considering assembly errors is performed by Abaqus. In the simulation, the torque applied to the pinion is 3730Nm while the gear is fixed totally. The relative position of the gear pairs is adjusted each time to simulate the meshing process. In this way the instantaneous contact load distributions in a whole mesh cycle are obtained as shown in Fig. 8. Then, the boundary condition of gears is interchanged and the loaded tooth contact analysis is conducted again. \u9f7f1 \u9f7f2 \u9f7f3 \u9f7f1 \u9f7f2 \u9f7f3 (a) 0 \u00b0 (b) 2 \u00b0 From the results, the simultaneous meshing gear pairs are alternatively two or three during one mesh cycle. The maximum contact pressure, 3969 MPa, is located in the tip region when a tooth meshes out. It is caused by stress concentration phenomenon. From the local enlarged view in Fig. 9, it can be seen that the edge contact and stress concentration appeared during meshing and the maximum contact pressure mainly locates in the edge region of the toe" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001402_s13369-012-0232-3-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001402_s13369-012-0232-3-Figure1-1.png", "caption": "Fig. 1 Pinion and gear mechanism with concave teeth and convex teeth", "texts": [ " 1 Denotes convex gear, 2 denotes concave gear \u03c6c Pressure angles \u03c61 Rotate angle about xr -axis \u03c62 Rotate angle about xn-axis \u03b1 A rotary angle and its region is 0 \u223c 2\u03c0 1 Convex tooth of the gear 2 Concave tooth of the gear Most gear manufacturers and designers want to develop more compact gear pairs with higher capacity. Generally, designs for tooth-shaped gears are involute gears. An involute gear has the advantages of allowing the use of linear rack cutters, leading to high production efficiency, line-contact meshing, constant pressure angle, insensitivity to center-distance variation, and simplicity. Some types of involute gears, such as helical gears and spur gears, are used in industry for power transmission between parallel axes. In this paper, another kind of involute gear is presented, as shown in Fig. 1. This paper also presents the method for determining a mathematical model of a rack cutter based on inverse envelope concept. Yang [1] presented a rack cutter with ring-involute tooth that was used to generate a pinion and a gear based on an envelope of a one-parameter family of the ring surfaces. The undercutting condition of the proposed gear was also presented. Significant advantages have been found in the shape of involute teeth, such as kinematic errors being independent of the central distance between gear centers [1]" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003409_cp.2013.2376-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003409_cp.2013.2376-Figure2-1.png", "caption": "Figure 2. Magnetic antenna pattern", "texts": [ " 1, e is the induced voltage of the magnetic antenna and can expressed as [4]: 2 cos cos cos cos A A nSEe t E t \u03c0 \u03bc \u03d5 \u03c9 \u03bb \u03d5 \u03c9 = = (1) where n is number of turns in the coil, S is the area of the coil plane, \u03bb is the signal wavelength, cosE t\u03c9 is the electric field at the receiving site, A\u03bc is the ferrite material permeability, \u03d5 is the included angle between the coil plane and the propagation direction of EM wave, and 2AE nSE\u03c0 \u03bb= . Once a magnetic antenna is manufactured, S, n, , E and A\u03bc are fixed. As shown in (1), e is only related to \u03d5 , so a magnetic antenna has a directive pattern, and cos \u03d5 is a normalized direction function. If only the amplitude characteristic is considered, a normalization magnetic antenna pattern similar to \u201c8\u201d is obtained [5], as shown in Fig. 2. When the propagation direction of the EM wave is parallel to the coil plane, i.e. \u03d5 = 0\u00b0 or \u03d5 =180\u00b0, the induced voltage of the magnetic antenna is maximum, which is the peak in the magnetic antenna pattern. When the propagation direction of EM wave is vertical to the coil plane, i.e., \u03d5 =90\u00b0 or \u03d5 =270\u00b0, the induced voltage of the magnetic antenna is minimum, which is the zero in the magnetic antenna pattern. III. ROTATING THE MAGNETIC ANTENNA PATTERN Once a magnetic antenna is installed and fixed on the mobile platform, its directivity is always constant to that of the mobile platform" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002290_978-3-642-33457-3_6-Figure6.6-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002290_978-3-642-33457-3_6-Figure6.6-1.png", "caption": "Fig. 6.6 A homopolar magnetic bearing designed by Kimman [34]", "texts": [ " A magnetically levitated body may have certain degrees of freedom (DOF) passively controlled, thus, by means of permanent magnets only; however, for stable levitation at least one DOF must be actively controlled, according to the theorem of Earnshaw [32]. As a rule, rotors are supported by pairs of AMB in differential mode that enables linear control of the magnetic force and, in turn, rotor position. Quite often, a permanent magnet is utilized to create the bias flux in the bearings and current of the coils is then used to control the rotor position. The most suitable for support of high speed rotors is the homopolar bearing structure in which the rotor (ideally) does not experience changes of the bias flux while rotating (Fig. 6.6). For a long time magnetic bearings were considered too complex and expensive to be commercially appealing [35]. However, in course of the last two decades AMB have proven their effectiveness and reliability and shown great potential for an increasing number of applications [33, 35]. AMB have several important advantages with respect to other types of bearings. They provide purely contact- and frictionless operation with no contamination and no need for lubricants. Beside frictionless operation, AMB offer possibility of creating practically arbitrary damping or stiffness, the property which can be greatly utilized to adjust the dynamical properties of the system as the rotational speed changes" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001734_iros.2011.6094411-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001734_iros.2011.6094411-Figure1-1.png", "caption": "Fig. 1: Mechanism of the Proposed CVT", "texts": [ " The purpose of this study was to develop a linear load-sensitive continuously variable transmission mechanism to automatically change the gear reduction ratio depending on the load with a large range of motion, and this article reports on the basic principle as well as the basic characteristics of the actual prototype machine. T 978-1-61284-456-5/11/$26.00 \u00a92011 IEEE 4048 CONTINUOUSLY VARIABLE TRANSMISSION MECHANISMREVIEW STAGE The continuously variable transmission (CVT) mechanism proposed in this study comprise of (a) spherical drive, (b) drive axis to rotate it, (c) variable motor housing with the passive rotational axis to vary the inclination angle of the active rotational axis, (d) fixed bracket to support these and (e) linear sliding plate as shown in Figure 1. The passive rotational axis for the variable motor housing goes through the center of the spherical drive. Rotary motion is converted into linear motion as the spherical drive rotates and pushes out the linear sliding plate. Next, the principle for CVT is described. Figure 2 shows the cross-section drawing of a transmitter. Defining that the radius of the spherical drive is r and the inclination angle for the active rotational axis is \u03b8, suppose that the \u03b8 value when the active rotational axis is in horizontal position against the linear sliding plate is 0 [deg]" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002960_detc2011-47688-Figure4-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002960_detc2011-47688-Figure4-1.png", "caption": "Figure 4. Schematic representation of Model 1", "texts": [ " To achieve this motion, a translational joint is placed between the axle seat and the modified hangers. The ADAMS/Car leaf spring model measures the displacement of the spring and sends this to an elasto-plastic leaf spring model [10] which is implemented in SIMULINK as an embedded MATLAB function. The elasto- plastic leaf spring model then solves for the spring force (Fs). Before the spring force is send back to ADAMS the vertical forces acting at the front (FzR) and rear (FzF) hanger interface points, or attachment points, (see Fig. 4) are calculated. FzR and FzF are calculated by Eqn.(3) and Eqn.(4). Eqn.(3) and Eqn.(4) are obtained from simultaneously solving the equation of the sum of forces in the z-direction and the sum of moments taken about the axle seat. fr sr zF ll Fl F + = (3) fr sf zR ll Fl F + = (4) The ADAMS model uses two point-point actuators which is placed between the axle seat and the front and rear hanger. The two point-point actuators are controlled by the forces FzR and FzF, respectively. The resulting force on the axle seat is the spring force Fs" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003242_9781400840601-016-FigureB.1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003242_9781400840601-016-FigureB.1-1.png", "caption": "Figure B.1 Rotation represented by a unit quaternion. The aircraft on the left is shown with the body axes aligned with the inertial frame axes. The aircraft on the left has been rotated about the vector v by = 86 deg. This particular rotation corresponds to the Euler sequence \u03c8 = \u221290 deg, \u03b8 = 15 deg, \u03c6 = \u221230 deg.", "texts": [], "surrounding_texts": [ "APPENDIX B\nQuaternions\nB.1 Quaternion Rotations Quaternions provide an alternative way to represent the attitude of an aircraft.While it could be argued that it is more difficult to visualize the angular motion of a vehicle specified by quaternions instead of Euler angles, there are mathematical advantages to the quaternion representation that make it the method of choice for many aircraft simulations. Most significantly, the Euler angle representation has a singularity when the pitch angle \u03b8 is \u00b190 deg. Physically, when the pitch angle is 90 deg, the roll and yaw angles are indistinguishable. Mathematically, the attitude kinematics specified by equation (3.3) are indeterminate since cos \u03b8 = 0 when \u03b8 = 90 deg. The quaternion representation of attitude has no such singularity. While this singularity is not an issue for the vast majority of flight conditions, it is an issue for simulating aerobatic flight and other extreme maneuvers, some of which may not be intentional. The other advantage that the quaternion formulation provides is that it is more computationally efficient. The Euler angle formulationof the aircraft kinematics involves nonlinear trigonometric functions, whereas the quaternion formulation results inmuch simpler linear and algebraic equations. A thorough introduction to quaternions and rotation sequences is given by Kuipers [127]. An in-depth treatment to the use of quaternions specific to aircraft applications is given by Phillips [25].\nIn its most general form, a quaternion is an ordered list of four real numbers. We can represent the quaternion e as a vector inR4 as\ne = \ne0 e1 e2 e3\n ,\nwhere e0, e1, e2, and e3 are scalars. When a quaternion is used to represent a rotation, we require that it be a unit quaternion, or in other words, \u2016e\u2016 = 1.\nIt is common to refer e0 as the scalar part of the unit quaternion and the vector defined by\ne = e1ii + e2ji + e3ki\nBrought to you by | New York University Bobst Library Technical Services Authenticated\nDownload Date | 7/16/15 4:34 PM", "as the vector part. The unit quaternion can be interpreted as a single rotation about an axis in three-dimensional space. For a rotation through the angle about the axis specified by the unit vector v, the scalar part of the unit quaternion is related to the magnitude of the rotation by\ne0 = cos (\n2\n) .\nThe vector part of the unit quaternion is related to the axis of rotation by\nv sin (\n2 ) = e1 e2 e3 .\nWith this brief description of the quaternion, we can see how the attitude of a MAV can be represented with a unit quaternion. The rotation from the inertial frame to the body frame is simply specified as a single rotation about a specified axis, instead of a sequence of three rotations as required by the Euler angle representation.\nB.2 Aircraft Kinematic and Dynamic Equations Using a unit quaternion to represent the aircraft attitude, equations (3.14) through (3.17), which describe the MAV kinematics and\nBrought to you by | New York University Bobst Library Technical Services Authenticated\nDownload Date | 7/16/15 4:34 PM", "dynamics, can be reformulated as\n .pn .pe\n.pd\n = e21 + e20 \u2212 e22 \u2212 e23 2(e1e2 \u2212 e3e0) 2(e1e3 + e2e0) 2(e1e2 + e3e0) e22 + e20 \u2212 e21 \u2212 e23 2(e2e3 \u2212 e1e0)\n2(e1e3 \u2212 e2e0) 2(e2e3 + e1e0) e23 + e20 \u2212 e21 \u2212 e22\n u v\nw\n \n(B.1) .u\n. v . w\n = rv \u2212 qw pw \u2212 r u\nqu\u2212 pv\n + 1\nm\n fx fy\nfz\n , (B.2)\n .e0 .e1 .e2\n.e3\n = 1 2 0 \u2212p \u2212q \u2212r p 0 r \u2212q q \u2212r 0 p\nr q \u2212p 0\n e0 e1 e2\ne3\n \n(B.3)\n .p .q .r = \n1 pq \u2212 2qr\n5 pr \u2212 6(p2 \u2212 r 2)\n7 pq \u2212 1qr\n + 3l + 4n 1 J y m\n4l + 8n\n . (B.4)\nNote that the dynamic equations given by equations (B.2) and (B.4) are unchanged from equations (3.15) and (3.17) presented in the summary of chapter 3. However, care must be taken when propagating equation (B.3) to ensure that e remains a unit quaternion. If the dynamics are implemented using a Simulink s-function, then one possibility for maintaining \u2016e\u2016 = 1 is to modify equation (B.3) so that in addition to the normal dynamics, there is also a term that seeks to minimize the cost function J = 1\n8 (1\u2212\u2016e\u20162)2. Since J is quadratic, we can use gradient descent to minimize J , and equation (B.3) becomes\n .e0 .e1 .e2\n.e3\n = 1 2 0 \u2212p \u2212q \u2212r p 0 r \u2212q q \u2212r 0 p\nr q \u2212p 0\n \ne0 e1 e2 e3\n \u2212 \u03bb\u2202 J \u2202e\n= 1 2 \u03bb(1\u2212 \u2016e\u20162) \u2212p \u2212q \u2212r p \u03bb(1\u2212 \u2016e\u20162) r \u2212q q \u2212r \u03bb(1\u2212 \u2016e\u20162) p\nr q \u2212p \u03bb(1\u2212 \u2016e\u20162)\n e0 e1 e2 e3 ,\nBrought to you by | New York University Bobst Library Technical Services Authenticated\nDownload Date | 7/16/15 4:34 PM" ] }, { "image_filename": "designv11_101_0003407_icepe-st.2013.6804316-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003407_icepe-st.2013.6804316-Figure2-1.png", "caption": "Fig. 2. Results of correction on spacer Images", "texts": [ " While coordinate transformation the coordinates may be non-integer, that is to say the point does not exist, if we use the values directly, it will go wrong. To solve this problem, this article uses bilinear gradation interpolation operation shown in the formula (10). f(i + u,j + v) = (I-u)(I-v)f(i,j) + u(l-v)f(i + l,j) + v(l-u)fU,j + I) +uvf(i + I,j + I) (10) According to formula (9), we can achieve the perspective distortion correction of spacer image. The determination of unknown parameter f, a is very important. In this paper, we determine which by the way of selecting reference point. Fig.2( c) and Fig.2( d) are the results of geometric distortion correction for Fig.2(a) and Fig.2(b) based on the above method, from which we can find that the shape distortion due to perspective principle has been improved significantly. IV. SPACERS FAULT RECOGNITION A. Extraction of spacer skeleton Now, the skeleton algorithm is divided into two categories, one is based on topological thinning, the other one is based on the distance transform. By comparison, this paper adopts a simple and effective method to extract the skeleton, the experiment proves that it has good effect such as less time-consuming" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001766_icmtma.2011.670-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001766_icmtma.2011.670-Figure2-1.png", "caption": "Fig. 2 Woking condition of a suspended wheeled mobile 2-link flexible manipulator G1,2,3-weight of component1,2,3; Fp-loader; F1-level driving force of the mobile manipulator", "texts": [ " The dynamic model, which takes into account the dynamic couplings of components of the mobile manipulator, is presented base on the multi-flexible-body dynamics. Based on the system dynamic model, The dynamic transient stability criterion is constructed. At last, Computation simulation results are presented to illustrate the study. A 2-link wheeled-suspended mobile flexible manipulator consists of 3 components as shown in Fig.1. Component1, component2 and component3, represents mobile base, flexible link1 of manipulator, and flexible link2, respectively. Fig.2 is the abstract drawing coming from Fig.1. O-x0y0,Oi-xiyi is ,respectively, the global coordinate system and body reference of body i(1,2,3). This paper makes the following assumptions: the wheels are weightless; the mobile base moves at the speed of constant linear velocity; the mobile base moves on the uneven ground defined by a simple harmonic function with the wheels and the ground remaining in contact at all times. Fig.1 A wheeled-suspended mobile flexible manipulator which can move on a rough terrain iK , represents,respectively, stiffness coefficients of spring and damping iC Much of the current researches in multi-body dynamics is devoted to the selection of system coordinates and system degrees of freedom that can be efficiently used to describe the system configurations [22]", " Therefore, the independent variable approach is applied to describe the system dynamic equations[22]: iT iq i eQ ii i i M q Q (2) Where represents the dependent variable vector of system configuration, represents, respectively, system mass matrix, vector of general force including inertia forces corresponding velocity. iq iiM iQ The analysis of tip-over stability of mobile manipulator with high speed is based on the system dynamic model. In order to match the high specified performance requirements, especially such as accurately positioning, the mobile manipulator should keep all wheels contacting with the ground at all times as well as not skidding. Therefore these c o n s t r a i n t r e l a t i o n s h i p s c a n b e d e s c r ib ed a s : 2 3>0, >0, >0fF F F ( 3 ) Where are presented in Fig.2. Skidding stability index is defined as 2 3,F F 2 1= -fF F F where is friction coefficient between ground and wheels. The level driving force of mobile base and reactions can be solved f r om func t ions2 and be desc r ibed a s fo l lows : 1F 2 3,F F 1 3 4 1 3 4 1 3 4 1 1 3 4 3 4 3 4 1 1 1 2 2 2 ( , , , , , , , , , , , , , , , , , )fi fi fi fi fi fi F F r r r q q q q q q ( 4 ) 1 1 2 1 2 1 0 1 1 01 1 2 1 0 = ( - - (-sin(2 / ) 1) - _ )+ ( - 2 / cos(2 / )) F K r d H vt wa K L C r d H v wa vt wa ( 5 ) 1 1 3 2 2 2 0 1 2 02 1 1 2 2 2 0 1 2 ( - (-sin(2 / 2 ( ) / ) 1)- _ ) ( 2 / cos(2 2 ( ) / )) F K r d H vt wa d d wa K L C r d H v wa vt wa d d wa / ( 6 ) Where parameters are shown in Fig.1 and Fig.2. With respect to a specified performance, the dynamic transient tip-over stability can be evaluated based on the functions 3. The system transient stable stages are defined by equations 3. According to equations 3, muti-goal optimization is established as follows: 1 3 4 1 2 3 2 3 min( , , , , , , ) >0 >0 >0f m m m L L L v F F F (7) In this paper, the state vector[22] is applied to describe the functions 2 as follows: ( , t) Y f Y (8) Functions 8 are ordinary differential equations. Gear\u2019s method is applied to solve functions 8 by the software Matlab2006" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001240_amm.143-144.422-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001240_amm.143-144.422-Figure3-1.png", "caption": "Fig 3. Open the walking mechanism in SolidWorks", "texts": [ " The transfer standard table between two software file IGES Parasolid STEP Render ACIS VRML SolidWorks have have have no have have ADAMS have have have have no no IGES Parasolid STEP Geometry appearance reserves reserves lost Quality characteristics lost reserves lost Color information lost reserves lost Error message have no have Between Solidworks and ADAMS of the Data Convert. Therefore, the subject of the software using Solidworks Parasolid standards will data file of the success of the import the ADAMS software. The details are as follows: Through the SolidWorks software will IGES format graphics files conversion for Parasolid formats: Open from SolidWorks with CATIA saved IGES. (*.igs) format file, was shown in figure 3. Choose a drop-down menu \"file-save as\", select Parasolid (*. x_t) format of the preservation of type. Input filename, is about to model file save for needed Parasolid (*. x_t) format (figure 4). Fig.4. Save the walking mechanism of SolidWorks Selection menu \"File-who\", or into the ADAMS/View to choose a Flie who. In the File, click the right column by triangle, the choice Parasolid as CAD format File. In the File To Read the column in the empty boxes, right click and select the File and Browse, or manual input filename, needed In the File by choice column, from the Parasolid ASC\u2161, Binary or Neutral choice in the input File types required" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001740_9781782421702.12.749-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001740_9781782421702.12.749-Figure1-1.png", "caption": "Figure 1 \u2013 The RV compressor design concept (3)", "texts": [ " NOMENCLATURE u specific internal energy (J/kg) v velocity (m/s) V volume (m3) w width (m) z height (m) Greek symbols \u03b1 angular acceleration (rad/s2) \u03b4 gap (m) \u03b7 friction coefficient (-) \u03d5 angle (rad) \u03bc viscosity (Pa\u00b7s) \u03c1 density (kg/m3) \u03a3 summation \u03b8 angle (rad) \u03c9 angular velocity (rad/s2) Subscripts 0 at center of vane or vane slot DCV discharge chamber disc discharge tank f friction imp impact m motor n normal r rotor P pressure R rotor center SCV suction chamber 749 s isentropic S point of interest at vane slot suct suction reservoir v vane V point of interest at vane ve exposed vane VS vane slot chamber VS0 vane slot without vane Ever since its invention, rotary machines have rapidly become more popular due to their superior vibration characteristics, simplicity and compactness. However, frictions between the rubbing components have been one of the limitations (1). In response, Teh and Ooi (2) introduced a novel mechanism called the Revolving Vane (RV) machine (Figure 1). Unlike the conventional rotary machines, the cylinder of the RV machine rotates together with the rotor, which reduces the relative velocities and frictional losses between the rubbing components. Many works have been carried out to model various aspects of the machine (2-7). The models have also been shown to be in good agreement with the experimental results (8). However, the existing simulation models have a lot of limitations, mainly because they assume that the dynamics of the components are completely geometrically defined" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002278_978-1-4471-4628-5_3-Figure3.1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002278_978-1-4471-4628-5_3-Figure3.1-1.png", "caption": "Fig. 3.1 Schematic diagram of slider crank mechanism", "texts": [ " However, we need mechanical actuators for tasks, such as force/torque amplification, change of speed by gears, transfer of motion from one axis to other by timing belt. We may also require a particular type of motion such as quick return motion where we may require mechanical actuation. Design of many mechanisms is based on two basic forms of kinematic chains (i) Slider crank mechanism (ii) Four bar mechanism. Let us begin our study with slider crank mechanism. Slider Crank Mechanism The schematic diagram of the typical slider crank mechanism is shown in Fig. 3.1. Note that this system has a single degree of freedom. Thus, all velocities are related to the crank rotation. The bond graph of the system is shown in Fig. 3.2. The input crank is driven by an effort source \u03c4 . The slider is moving against a spring of stiffness K and frictional damper Rs . The angular motion of the crank is represented by the 1\u03b8\u0307 junction with the I element attached to junction. This I element represents the rotary inertia (J1) of crank about its axis. The I element here models the complete dynamics of the crank" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001414_cobep.2013.6785219-Figure10-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001414_cobep.2013.6785219-Figure10-1.png", "caption": "Fig. 10. 3D Magnetic flux density with DC excitation of 1A.", "texts": [ " It is possible to conclude, from the figure 8, when the electrical current increases in the coil, it also increases the effect of armature reaction, but this effect does not cause a significant difference in the magnetic flux density average, making their average value in 0.332T. Also, by means FEM, it is possible to analyze the magnetic flux distribution in the area presented in the Figure 4b; it is important to notice that this analysis is really important to verify if the ferromagnetic material is magnetically saturated. Figure 9 and 10 present these results as function of the coil excitation. From the figures 9 and 10, it was realized that the magnetic flux density is not significantly altered, which had been verified in the analysis of figure 8. Figure 10 shows that the region where is concentered more magnetic flux density is in PM, because it is the magnetic flux source. Another important analyzis is about the behavior and distribution of the magnetic flux density in the area closer to the permanent magnetic, in the 3D view. For it is monitored the quantity in front on the permanent magnetic, in the same region than the sampling line, but now considering the area. Figure 11 shows these results as function of coil excitation. (a) The linear traction force, responsible for operation of the electromagnetic device is here analyzed as function of current applied in the coil by means FEM and measurements for two situations: when the permanent magnets are completely aligned with the coil (figure 12) and when the PM is 50% aligned with the coil (figure 13)" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001202_icecs.2013.6815485-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001202_icecs.2013.6815485-Figure1-1.png", "caption": "Figure 1. Khalifa University UAV (KUAV)", "texts": [ " A proportional-derivative (PD) law is used by integrating the new attitude reference for the attitude control of quad-rotor UAVs. Simulation results prove the efficiency of the new method which provides a new model with intuitive physical meaning for quadrotor UAVs. I. INTRODUCTION Unmanned aerial vehicles (UAVs) have been prevalently employed in numerous defense- and civil-related applications over the last two to three decades. Among various types of UAVs, miniature quad-rotor UAV (an example, constructed at the Robotics Institute at Khalifa University, is shown in Fig. 1) has gained particular popularity due to its unique features such as compact size, good agility, high maneuverability and vertical take-off and landing (VTOL). For now they are widely studied and used in surveillance, first responder for fire and civil accidents, inspection, photography and mapping [1]. Quad-rotor UAVs generally have a cross shape with four rotors arranged at the four ends of the cross. By controlling speeds of the four rotors in variable combinations, under controlled strategy is applied to control four degrees of freedom (DOFs) in realizing the 6-DOF translations and orientations", " This was used to be a reference for attitude control between the actual and the desired motion. The new method avoided the singularity problem of Euler angle model by applying the body angular velocity directly which could also improve the processing speed considering the calculation in real time. A proportional-derivative (PD) control law was used in the simulation which verified the effectiveness of the new modeling and control method. Future work will focus on the experiment implementation with the KUAV platform built at the Robotics Institute at Khalifa University as shown in Fig. 1. [1] Minh-Duc Hua, Tarek Hamel, Pascal Morin, Claude Samson, Introduction to Feedback Control of Underactuated VTOL Vehicles: A Review of Basic Control Design Ideas and Principles, IEEE Control Systems, 33(1), pp. 61-75, 2013. [2] S. Bouabdallah, A. Noth, and R. Siegwart. PID vs LQ Control Techniques Applied to an Indoor Micro Quad-rotor. In Intelligent Robots and Systems, pages 2451\u20132456, 2004. [3] P. Castillo, R. Lozano, and A. Dzul. Stabilization of a Mini Rotorcraft with Four Rotors. IEEE Control Systems Magazine, pages 45\u201355, 2005" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001450_0954406213497899-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001450_0954406213497899-Figure3-1.png", "caption": "Figure 3. The experimental setup: (a) schematic diagram; (b) real photo of the system.", "texts": [ " The CMAC performance is measured when it is connected in parallel form with a conventional servo controller of a PV type. CMAC is implemented on MATLAB/SIMULINK R2007b software package. The system, which is the subject under control, is an electrohydraulic servo system that is equipped to give a linear simple harmonic motion using a hydraulic cylinder. The desired and the actual motion are represented by the piston-rod displacements (xd and xa, respectively). The overall structure of the system can be simplified by the block diagram representation as shown in Figure 2, and the experimental setup is shown in Figure 3. The hydraulic circuit is shown in Figure 3(a) ((1) variable displacement vane pump, (2) proportional valve, (3) pressure transducer, (4) potentiometer, and (5) data acquisition card), and the test rig setup is shown in Figure 3(b) ((1) DC power supply 24 V, (2) power amplifier card, (3) on/off switch, (4) DC power supply 10V, (5) proportional control valve, (6) pressure transducers, (7) hydraulic cylinder, (8) digital multimeter, (9) linear potentiometer, and (10) DAQ I/O connection board). The root mean square (RMS) is a statistical measure of the magnitude of a varying quantity. It is especially useful when the variations are positive and negative, such as sinusoidal functions. The RMS value for a continuous function e(t) is defined over the interval T14 t4T2 by equation (14)" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000024_2013.40604-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000024_2013.40604-Figure2-1.png", "caption": "Figure 2. Michelin X tire geometry at during Stryker maneuvers at PTA (units: meters)", "texts": [ " The frontal area, fA , was determined to be approximately 4.5 m2 based on the geometry of the Stryker. The vehicle is equipped with a Central Tire Inflation System (CTIS) that allows the operator to vary the inflation pressure of all tires simultaneously according to the terrain conditions (Ayers et al., 2009). All wheels were equipped with Michelin X tires. The inflation pressure of the tires remained a constant 483 kPa during the 2009 Stryker maneuvers. The tire parameters necessary for applying Equation (2) and (5) are given in Figure 2 while the tire deflection represents in Figure 2 was at a 483 kPa inflation pressure. The steering angles of wheels on the steered axles of the Stryker were determined from the calculated turning radius of the vehicle by assuming the slip angle of each tire was negligible. The turning radius was calculated from the vehicle's GPS data. For the steered angle calculations, it was assumed the vehicle's center of gravity was at the geometric center of the vehicle, and the vehicle turned about this point. The normal loads on each tire were assumed to be equal and constant during maneuvers with minimal effects due to weight transfer" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000900_amr.562-564.654-Figure5-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000900_amr.562-564.654-Figure5-1.png", "caption": "Fig. 5 Cam mechanism with Swing link", "texts": [ " Link 2 can translate along x- and y-axis, but cannot rotate about z-axis. The constraint of the general pair 2.5 IG is 2,5 2,5 ( ,0 0 ) 4I Im m z\u03b1 \u03b2 \u03b3= = . And the common constraint ( ,0 0 ) 4gz gz II IIm m z\u03b1 \u03b2 \u03b3= = ,so ( ) ( ) ( )2,5 ,0 0 0 0 ,0 0 4X gz II I IIm m z m z z\u03b1 \u03b2 \u03b3 \u03b1 \u03b2 \u03b3 \u03b1 \u03b2 \u03b3= + = =\uff0c .The mobility of the mechanism is 6 3 2 [(3 3) (4 3)] 1F = \u2212 \u00d7 + \u2212 + \u2212 = . The mobility of link group EF is 2 6 4 0IIF = \u2212 + = , which means link EF is the Assure group. It is equivalent to the 2R link group in example 1. Example 3. Fig. 5 is Cam mechanism with swing link, joint C with two freedom is a higher pair and the relative velocity of point C is always changing. Roller 2 has the relative displacement both along x-axis and y-axis. The mobility of the mechanism is 5 3 1 (3 3) 2F = \u2212 \u00d7 + \u2212 = . Example 4. Fig.6 shows a Stephenson\u2019s six bar mechanism. In loop I, there is 3X I Im m= = . The constraint of generalized pair 2,4 IG in loop II is ( )2.4 ,0 0Im z\u03b1 \u03b2 \u03b3= , the common constraint of link group EF is ( )0,0 0gz IIm z\u03b1 \u03b2= , so the virtual constraint is ( ) (2,4 0,0 0 0 I II X gz IIm m z m\u03b1 \u03b2 \u03b1 \u03b2= + ) ( ),0 0 0,0 0 3z z\u03b1 \u03b2= = , the mobility of loop II is 2 6 3 1IIF = \u2212 + = \u2212 " ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000164_amm.105-107.62-Figure8-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000164_amm.105-107.62-Figure8-1.png", "caption": "Fig. 8 Quasi-periodic Oscillation\uff08 1.231\u2126 = \uff09", "texts": [ " So power spectrum is distributed at separate point, which is equal to / 2Bh\u03c9 (where B\u03c9 is basic frequency and h is positive integer). In Fig. 7, \u2126 is equal to 1.104. Monocycle anharmonic response appeared in the system. In the time history, there is monocycle movement in the system. The period is equal to that of inspirit function( 2 / 5.69\u03c0 \u2126 = ). And this is not harmonic vibration. The corresponding phase diagram is noncircular closed curve. And the Poincare section includes one separate point. The power spectrum is distributed at separate point, which is equal to Bh\u03c9 . In Fig. 8, \u2126 is equal to 1.231. There is quasi-periodic oscillation in the system. The corresponding phase diagram is a closed curve strap with specifically width. Poincare section includes three closed curve. In Fig. 9, \u2126 is equal to 1.250. And the response is subharmonic solution with treble period. And the period is treble period of inspirit function(3 2 / 15.072\u03c0\u00d7 \u2126 = ). And the Poincare section includes three separate point. So power spectrum is distributed at separate point, which is equal to / 3Bh\u03c9 " ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003703_1.5062968-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003703_1.5062968-Figure1-1.png", "caption": "Figure 1. Laser scanning approaches for the post-processing of LAM-produced parts. Top: Ablation of a single material layer at a constant z-offset to the design shape. Bottom: Processing of multiple layers with varying step \u0394z.", "texts": [ "5 \u2013 5 MHz, maximum average power Paver = 5 W) was used for the laser postprocessing studies. Laser radiation was focused on the surface of the samples using a 5\u00d7/NA = 0.1 microscope objective and a CCD-based imaging system to control the beam position. The samples were translated by a 3D linear positioning stage Aerotech ALS130 at a constant speed v = 10 mm/s. Different scanning approaches were applied to maintain the design consistency, improve the resulting surface finish and processing efficiency (Fig. 1). Curved surfaces were processed in multiple overlapping tracks resulting in layers that followed the final design shape at a constant vertical z-offset. This procedure was repeated for different offset values with varying step \u0394z = 10 \u2013 50 \u00b5m until the final design shape (z = 0) was revealed. The lateral separation between the tracks was \u0394x = 25 \u00b5m in all irradiation experiments. The resulting surface profile of the processed areas was investigated by optical microscopy and white-light interferometry (Zygo NewView 6300)" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000075_amr.457-458.237-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000075_amr.457-458.237-Figure2-1.png", "caption": "Fig. 2 Frames and forces assignment", "texts": [ " Ducted fans are mounted on the two sides of rear end of the cabin, which can carry out vector thrust. There are elevators and rudders on the surface of the airship. All rights reserved. No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of TTP, www.ttp.net. (ID: 128.193.164.203, Oregon State University, CORVALLIS, United States of America-26/06/14,13:10:44) To study forces and motion law of the airship, and determinate its position and attitude, we choose the reference frames which are as shown in Fig. 2, including horizontal frame ( e e e e e s o x y z\u2212 ), velocity frame ( a a a a s ox y z\u2212 ) and body-fixed frame ( b b b b s ox y z\u2212 ). For small area and short period motion and mechanics process, horizontal frame is considered as an inertial frame. The origin of velocity frame and body-fixed frame is chosen as the volume center of the airship, not the center of gravity, because the center of gravity varies with altitude, surrounding and running station, while the volume center varies little. Parameters of airship motion state include position ( )Te e e ex y z=P , velocity in horizontal frame ( )Te e e ex y z= P , velocity in body-fixed frame ( )Tu v w=v , angular velocity in body frame ( )Tp q r=\u2126 , attitude angle [ ]T\u03d5 \u03b8 \u03c8\u03a6 = , angle of attack \u03b1 and angle of sideslip \u03b2 " ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000461_icmree.2011.5930965-Figure8-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000461_icmree.2011.5930965-Figure8-1.png", "caption": "Figure 8. The pressure distribution of central region towards ground", "texts": [ " The solution of eon verge nee eurve (3) The load imposition and load solving The pressure of 830kPa (standard inflation pressure), 730kPa, 630kPa and 530kPa were imposed respectively in the inner surface of the finite element model and the load model is shown in Figure 6. In this paper, the large deformation nonlinear method is used to solve the load and its solution of convergence curve is shown in Figure 7.And also the number of iteration is 70. IV. ANALYSIS OF THE FINITE ELEMENT SIMULATION (1) The pressure analysis of central region towards ground Figure 8 shows the ground pressure curve of retreaded tire. It is analyzed from Figure 8 that, the pressure of the central region towards ground is 138478Pa and the pressure around the center towards ground is 348360Pa and 625316Pa, which suggests that the pressure of the central region towards ground is the least, and that the pressure around the center along the width and the tire rolling direction towards ground increase. Figure 9 describes the pressure curve of the central region towards ground under the different working conditions. It is seen from Figure 9 that, if the vertical load is lower, the pressure of the central region towards ground is larger" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001396_amm.121-126.3477-Figure6-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001396_amm.121-126.3477-Figure6-1.png", "caption": "Fig. 6 Temperature field with rotating velocity 48rpm Fig. 7 Temperature field with rotating velocity 64rpm", "texts": [ " The flow of lubricant through the hole is equal to the bearing input flow. In order to study the rotating velocity impacting on the film temperature field, when the inlet flow capacity is 100 kg/s, external pressure is 0.1Mpa, and choose viscosity of lubricating oil \u00b5 equals 0.0288 Pa\u00b7s, density \u03c1 equals 900 kg/m 3 , the worktable rotating velocity assign respectively 8rpm, 16rpm, 24rpm, 32rpm, 40rpm, 48rpm, 56rpm, 64rpm, 72 rpm and 80rpm, the three-dimensional temperature fields are obtained in CFX to be as Fig.4, Fig.5, Fig.6,Fig.7, Fig.8. The curve of concerning the rotating velocity and temperature is shown in Fig.9. According to the results obtained, the conclusions can be drawn as follows. Numerical calculation of the three-dimensional temperature field of annlar cavity multi-pad hydrostatic thrust bearing is achieved through Computational Fluid Dynamics and the Finite Volume Method. It solves the problem that temperature field is not directly measured in the heavy type hydrostatic thrust bearing because oil film is very thin in the fact project" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001420_jot.78.000066-Figure6-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001420_jot.78.000066-Figure6-1.png", "caption": "FIG. 6. Method based on the loss of total internal reflection.", "texts": [ " The signal level from the photodetector falls off when the end of the needle lies in the epidural space (53 mm), because the tissue of the epidural space is a substrate that possesses a higher absorption coefficient for the radiation. Experiments are currently being done on a prototype of the device, based on a compact USB 650 spectrometer from the Ocean Optics Co. The effect of the breakdown of the total internal reflection angle from the polished end of the lightguide can be used to verify liquid and gelatinous biological tissues (substrates). The total-internal-reflection angle \u03c90 is determined for the optical radiation incident on the end of the lightguide (Fig. 6) by the well-known relationship sin\u03c90 = n1/n2, where n1 is the refractive index of the external medium, and n2 is the refractive index of the lightguide. When the end of the lightguide lies in air, n1 = 1, and the total-internal-reflection angle for a lightguide made from a polymer with n2 = 1.49 will equal \u03c90 \u2248 42\u25e6. The biological tissue in the epidural space includes connective tissue, fat, and vessels and has a refractive index in the range n1 \u2248 1.3\u20131.4; therefore, for a lightguide made from a polymer with n2 = 1", ", arcsin (1/n2) < \u03b1 < arcsin (n1/n2) , then, because of the breakdown of the total-internal-reflection angle at the instant the end of the needle (the lightguide) breaks through into this tissue (liquid), a significant part of the light will propagate into this tissue, the fraction of reflected radiation will sharply decrease, and the signal amplitude from the photodetector will fall to the minimum value. The biological tissue along the path of the needle lying in front of the epidural space (muscle, supraspinous, interspinous, ligamentum flavum) has a denser consistency, wets the polished end of the lightguide with, for instance, blood to a smaller extent, and is virtually opaque for the visible region. Therefore (Fig. 6), as long as the end of the needle lies, for instance, in the tissues of the interspinous and ligamentum flavum 2, light ray AB, incident on the end of the lightguide, which is made at an angle of \u03b1, undergoes total internal reflection and is deflected along direction BC. A significant part of it, after multiple reflections from the walls and end of the lightguide, comes back and is incident on the photodetector. When the end of the needle reaches the epidural space, total internal reflection does not occur, and the signal from the photodetector sharply decreases" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001652_ecj.10392-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001652_ecj.10392-Figure3-1.png", "caption": "Fig. 3. Procedure of BFA.Fig. 2. Copies of point P.", "texts": [ " Assumption 2 is not serious, that is, collisions among links can be removed easily. Usually, and especially in three-dimensional workspaces, there are enough free spaces around attitudes that do not collide with obstacles, and therefore self-collisions can be removed by locally adjusting the attitudes of relevant links. Section 3 discusses the method. 2.3 The path planning algorithm and its computation volume Based on the above proof procedure, a backtrack-free path planning algorithm BFA can be constructed easily. Figure 3 shows the overall structure of the algorithm, where the link numbers are defined so that 1 is assigned to the base link. The algorithm consists of two parts, the off-line and the real-time parts. The off-line part is executed only when the locations of the manipulator or obstacles are changed and the real-time part is executed every time a goal attitude is given to the manipulator. For each n, from n = N to n = 1, the off-line part finds feasible attitude sets of the n-th link at individual points, and based on them, it determines the (n \u2013 1)-connectivity of individual neighboring point pairs and calculates (the n \u2013 1)-reachable set R(n \u2013 1)" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001488_amc.2012.6197112-Figure5-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001488_amc.2012.6197112-Figure5-1.png", "caption": "Fig. 5. Experimental setup.", "texts": [ " A log likelihood is applied to eq. (9), the probability of \ud835\udc57th window \ud835\udc43\ud835\udc57(\ud835\udc9a\ud835\udc61\u2223\ud835\udf06\ud835\udc5b) is described as follows \ud835\udc43\ud835\udc57(\ud835\udc9a\ud835\udc61\u2223\ud835\udf06\ud835\udc5b) = 1 \ud835\udc5a \ud835\udc5a\u2211 \ud835\udc56=1 log ( \ud835\udc4e\ud835\udc5b\ud835\udc56\ud835\udc61,\ud835\udc56\ud835\udc61+1 \ud835\udc4f\ud835\udc5b\ud835\udc56\ud835\udc61,\ud835\udc56\ud835\udc61+1 (\ud835\udc9a\ud835\udc61)\ud835\udc64(\ud835\udc61\u2212 \ud835\udefc) ) (11) where \ud835\udc5a denotes sampling number. The nearer log likelihood approaches 0, the nearer the motion approaches the model of ones. The probability of window which has biggest probability is the probability of model. And, the model whose log likelihood of probability is highest is assumed recognition result. In this paper, hand type robot is used.It is shown in Fig. 5. This system is contructed with linear motor as master system and hand robot as slave system. Hand robot\u2019s motor is rotary geared motor and resolving power of the encoder is 1600 pulse/rev. Motor of master system is linear motors and position encoders whose resolutions are 100 nm. And C++ language is used for control and it is mounted by RTAI 3.7. A soft ball is used as a envrionment. Parameters of experimental system are shown in Table I In experiment, width of window is 10 s, and distance of next window is 1 s" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001579_elan.201300010-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001579_elan.201300010-Figure1-1.png", "caption": "Fig. 1. The scheme in cross section of the composite s electrode structure.", "texts": [ "6 ratio; this composition has been used in several previous works [9, 25,29] that mention the 62% composition rendering the best electrochemical properties displayed by the graphite-epoxy resin composite electrodes. A tubular container piece served to integrate the composite mix into a relatively small PVC tube, fitted with a metal electric contact on one side, while on the other, the composite graphite-epoxy resin was shoved in aiming to fill approximately half the volume available, as shown in Figure 1. After, the newly manufactured device is put to consolidate for 24 h in a laboratory stove at 60 8C. Several films were electrosynthesized over a high purity graphite electrode, changing the main potentiostatic electropolymerization parameters, namely: \u2013 Applied potential: considering a working range from 600 to 1100 mV according to the results of the cyclic voltamperograms. \u2013 Benzoate concentration: preliminary experiments demonstrated that at concentrations lower than 1 10 3 M there is no modification of the electrode s surface, while concentrations above 0" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002119_icmech.2011.5971277-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002119_icmech.2011.5971277-Figure1-1.png", "caption": "Fig. 1. Diagram of a micropart feeder with a sawtoothed surface and symmetric vibrations", "texts": [ " To move microparts in one direction, the driving force applied to each micropart must vary according to its direction of motion. The motion of sub millimeter sized mi crop arts can be affected not only by inertia but by adhesion due to electrostatic, van der Waal's, and intermolecular forces, and by surface tension [1]. The effects of adhesion are dependent on the two materials in contact with each other. We have developed a novel microparts feeder by utilizing an asymmetric fabricated surface, such as a sawtoothed surface, as a feeder table (Fig. 1) [2]. This asymmetric fabricated 978-1-61284-985-0/11/$26.00 \u00a92011 IEEE surface can feed microparts, such as, in one direction using horizontal and symmetric vibrations because contact between the micropart and the asymmetric fabricated surface varies according to the direction of motion. To accurately formulate the dynamics of microparts, it is necessary to analyze both inertia caused by vibration of the feeder surface and contact dependent adhesion. In this paper, we assessed the affects of feeder surface materials on the feeding of microparts" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003409_cp.2013.2376-Figure4-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003409_cp.2013.2376-Figure4-1.png", "caption": "Figure 4. Orthogonal antenna pattern rotating", "texts": [ " According to the antenna output voltage expressed in (1), antennas A1 and A2 can be obtained using the output voltage u1 and u2 as: 1 cos cosAu E t\u03d5 \u03c9= (2) 2 cos(90 )cos sin cos A A u E t E t \u03d5 \u03c9 \u03d5 \u03c9 = \u2212 = (3) After u1 and u2 are weighted, u3 and u4 are obtained as: 3 1 cos cos cos cosA u u E t \u03b2 \u03d5 \u03b2 \u03c9 = = (4) 4 2 sin sin sin cosA u u E t \u03b2 \u03d5 \u03b2 \u03c9 = = (5) The synthetic output voltage u is. 3 4 A cos cos cos sin sin cos cos( ) cos A A u u u E t E t E t \u03d5 \u03b2 \u03c9 \u03d5 \u03b2 \u03c9 \u03d5 \u03b2 \u03c9 = + = + = \u2212 (6) In (2), cos( )\u03d5 \u03b2\u2212 is a normalized direction function of the synthetic output voltage u. If the rotation angle \u03b2 is set to be the benchmark, namely 0\u03b2 = , the normalized pattern of the orthogonal magnetic antenna is shown in Fig. 4(a), which also has a \u201c8-shape\u201d. If the rotation angle is changed, the antenna pattern is rotated as shown in Fig. 4(b). According to the output maximum SNR or the output maximum voltage of the receiver, the antenna pattern rotation system generates the rotation angle , so that it realizes rotation of the antenna pattern. When \u03b2 \u03d5= or 180\u03b2 \u03d5= \u00b1 , the antenna pattern rotation system outputs the maximum voltage maxu , which is the peak in the antenna pattern. When 90\u03b2 \u03d5= \u00b1 , the antenna pattern rotation system outputs minimum voltage maxu , which is the zero in the antenna pattern. According to specific requirements, the rotation angle may be input manually" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000312_amr.291-294.2970-Figure5-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000312_amr.291-294.2970-Figure5-1.png", "caption": "Figure 5. Slice processing", "texts": [ " After the generation of three-dimensional entity model of casting, it needs to translate solid model into STL file by computer interface module, as shown in Figure 4. STL file are universal standardized input format in rapid prototyping manufacturing system, which is mainly used to connect three-dimensional entity model and rapid prototyping manufacturing system. This file is composed of many data representing a series of small triangle which used to express the shapes and sizes of three-dimensional entity model. Magics6.2 software is used to slice processing on the entity model, converting three-dimensional graph to two-dimensional graph, as shown in Figure 5. According to a certain precise requirements and STL file casting model, Magics6.2 software can slice cross section outline along the height direction of model at regular intervals, and then the requisite casting model may be rapid generated by repeating sintering of laser. The slice thickness of casting model shown in Figure 5 is 0.1 mm. After slice processing, AFI machining codes were generated by ARPS. After the generation of AFI machining codes, it is sent to AFS-3000 rapid prototyping system which is comprised of computer, laser scanning system, raw material storage, feed mechanism, hot-press arrangement,lifting table and NC System.The main parameter of forming process is shown in Table 1. Table.1 The main parameter of forming process Laser power Temperature of starting sintering Temperature of processing sintering powder thickness powder speed scanning speed 50W 120\u2103\uff5e130\u2103 100\u2103\u00b12\u2103 0" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001425_ijptech.2011.038108-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001425_ijptech.2011.038108-Figure1-1.png", "caption": "Figure 1 Benchmark dimensions", "texts": [ " Following objectives have been set for present experimental study: 1 to study feasibility of decreasing the shell thickness from recommended one (12 mm) for statistically controlled RC solution of aluminium alloy in order to reduce the production cost and time 2 to evaluate the dimensional accuracy of the castings obtained and to check the consistency of the tolerance grades of the castings (IT grades) as per allowed IS standards for casting process 3 proof of concept, to present the concept in physical form with minimum cost by avoiding the cost of making dies and other fixtures for a new concept. In order to accomplish the above objectives, \u2018aluminium casting\u2019 has been chosen as a benchmark (Figure 1), representative of manufacturing field, where the application of RT and RC technologies is particularly relevant. The experimental procedure started with drafting/ model creation using AutoCAD software (Figure 2). For the process of RC process based on 3DP, following phases have been planned: 1 After the selection of the benchmark, the component to be built was modelled using a CAD (Figure 3). The CAD software used for the modelling was UNIGRAPHICS Ver. NX 5. 2 The upper and lower shells of the split pattern were made for different values of the thickness" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002960_detc2011-47688-Figure6-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002960_detc2011-47688-Figure6-1.png", "caption": "Figure 6. Forces at point of contact", "texts": [ " Downloaded From: http://proceedings.asmedigitalcollection.asme.org on 10/10/2018 Terms of Use: http://www.asme.org/about-asme/terms-of-use 4 Copyright \u00a9 2011 by ASME The longitudinal forces FxR and FxF are simply calculated by taking the vertical forces (FzR and FzF) and relating it to the longitudinal forces (FxR and FxF) via the slope of the leaf spring at the point of contact between the leaf spring and the bearings. The assumption is made that the contact between the leaf spring and the bearing consists of a thin line. Figure 6 shows the forces acting at the point of contact between the leaf spring and the bearing. When we know the vertical forces (FzR and FzF) and the angle of the slope (\u03b1) we will be able to calculate the longitudinal forces (FxR and FxF). We obtain the relationship between the deflection of the leaf spring (z) and the angle of the slope of the leaf spring at the point of contact with the bearing as follows. Table 1. Angle of slope at different deflections z [m] \u03b1r [deg] 0 19.3 -0.016 14 -0.032 8.5 Table 1 shows the experimentally measured angle of the slope for three deflections of the leaf spring" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001044_imece2011-62617-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001044_imece2011-62617-Figure2-1.png", "caption": "Figure 2. MOMENTS AND SHEAR FORCES ON PLATE", "texts": [ " Using the assumptions of Mindlin-Reissner plate theory and small displacement bending theory, the displacements of a plate at any arbitrary location (x,y,z) may be obtained as: ( yxzu x . )\u03b2= (1) ( yxzv y . )\u03b2= (2) (3) ( yxww .= ) where u, v, and w are the displacement components in the x, y, and z-directions respectively, and \u03b2x and \u03b2y denote the rotations of the normal to the plate middle plane in the xz and yz planes respectively. The notations for rotations, shear forces, and moments are shown in Fig. 2 and Fig. 3. The bending strain is given by: 2 Copyright \u00a9 2011 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/conferences/imece2011/70896/ on 06/07/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use \u23aa \u23aa \u23aa \u23aa \u23ad \u23aa\u23aa \u23aa \u23aa \u23ac \u23ab \u23aa \u23aa \u23aa \u23aa \u23a9 \u23aa\u23aa \u23aa \u23aa \u23a8 \u23a7 \u2202 \u2202 + \u2202 \u2202 \u2202 \u2202 \u2202 \u2202 = \u23aa \u23ad \u23aa \u23ac \u23ab \u23aa \u23a9 \u23aa \u23a8 \u23a7 xy y x z yx y x xy y x \u03b2\u03b2 \u03b2 \u03b2 \u03b5 \u03b5 \u03b5 (4) The domain is divided into control volumes or plate elements. The governing equations for an individual control volume are obtained by balancing the forces in the z-direction and the moments along the x and y axes: wIFz &&1=\u2211 (5) (6) xx IM \u03b2&&3=\u2211 (7) yy IM \u03b2&&3=\u2211 where I1 and I3 are the inertia moments" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002239_978-94-007-5006-7_3-Figure3.9-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002239_978-94-007-5006-7_3-Figure3.9-1.png", "caption": "Fig. 3.9 Representation of DH frames for ZXZ EAJs", "texts": [ "20), Q\u02d9X is the rotation matrix corresponding to the twist angle, and Q k and Q\u02d9Z correspond to the joint angles. Based on Eq. (3.20), the DH parameters are shown in Table 3.3. Like the previous section, one can relate the rotation matrix by using ZXZ Euler angles as a combination of three intersecting revolute joints. The representations of DH frames, however, differ from what has been done for ZYZ EAJs. Assignment of the DH frames for the three intersecting revolute joints representing ZXZ EAJs is shown in Fig. 3.9. The corresponding rotation matrices between F1 and FR, F2 and F1, and FM and F2 are given below: Q1 2 64 S 1 C 1 0 C 1 S 1 0 0 0 1 3 75 ; Q2 2 64 C 2 S 2 0 0 0 1 S 2 C 2 0 3 75 ; and Q3 2 64 S 3 C 3 0 0 0 1 C 3 S 3 0 3 75 (3.21) The overall rotation matrix, QZXZ, between the frames FM and FR is then given by QZXZ D 2 4 S 1C 2S 3 C C 1C 3 S 1C 2C 3 C 1S 3 S 1S 2 C 1C 2S 3 C S 1C 3 C 1C 2C 3 S 1S 3 C 1S 2 S 2S 3 S 2C 3 C 2 3 5 (3.22) Here too, it can be seen that the orientation matrix QZXZ of Eq" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000105_978-3-642-20222-3_3-Figure3.82-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000105_978-3-642-20222-3_3-Figure3.82-1.png", "caption": "Fig. 3.82 (continued)", "texts": [ " The illustrations that follow present the start-up curves of a medium-power induction motor (S2L) with the application of CSI current inverter controlled by means of PWM modulation. The induction motor supplied from CSI inverter with small inertia and low load torque as well as from capacitors with low value of capacity tends to operate in an unstable way. An illustration of this is found in Fig. 3.81. The higher the inertia, capacity of a filter capacitor Cf and the negative feedback in relation to the speed, the more stabilized is the operation of the drive after start-up \u2013 see Fig. 3.82. a) b) Fig. 3.82 Starting of the medium power (3.96) CSI PWM drive with J = 3Js, T/Tn = 0.2, Cf = 150 [\u03bcF]: a) DC current b) electromagnetic torque c) stator currents d) angular speed e) torque-speed trajectory The examples presented in Figs. 3.82 and 3.83 indicate that the appropriate selection of the motor\u2019s parameters and feedback lead to the stabilization of the drive even for relatively small load Tl /Tn = 0.25. The current IDCr = 98 [A] is the input value for stabilization of the inverter supply current after start-up" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003800_1.5062238-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003800_1.5062238-Figure1-1.png", "caption": "Figure 1 Deposition tool with the laser scanner. Left: deposition mode. Right: scanning mode.", "texts": [ " At the currently chosen measurement distance (between the scanner and the part) the length of the laser line is around 35 mm. Here, 256 points/line are used. Hence, with a 35 mm wide laser line, a spatial resolution of 140 \u00b5m in the yaxis is obtained. During deposition it is important to protect the scanner from the high power laser reflections and the heat radiation from the built part. For this purpose, a linear drive unit, on which the scanner is mounted, is utilized such that it lifts the scanner away from the melt pool during the actual deposition, and lowers it down at the time of scanning, see Figure 1. The repeatability of the linear drive unit in the z-direction is measured to \u226430\u00b5m. User Interface A user interface is designed to provide an operator with all the necessary information needed to run the process remotely from outside the laser room. This is an important feature due to the potential risks associated with the use of high power lasers. If necessary, the interface allows the user to make on-line changes of the laser power, the deposition speed, the wire feed rate, and the robot's height" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000663_icssem.2012.6340798-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000663_icssem.2012.6340798-Figure1-1.png", "caption": "Figure 1 aerial work platform", "texts": [ "eywords-component; aerial work platform angle sensor steering system closed loop control I. INTRODUCTION Aerial work platform is the equipment which can lift works and load to the required height. It structure is composed of three parts: Part 1,that is the basic parts the beam, walking wheels, steering structure and so on. Part 2, Work platform, a basket, guardrail and so on; Part 3: scissor part: it consists of links between chassis assembly and the platform assembly. Shown in Figure 1, the scissor arms though the hydraulic cylinder complete lifting operating. Due to the pure electric aerial work platform can self-propelled, electric start, Simple operation, and large work surface, has been widely used in municipal maintenance such as airports, docks, cargo transportation logistics center, large factories etc. II. INTRODUCTION TO ORIGINAL STEERING SYSTEM STRUCTURE The original steering system structure is composed mainly of the steering wheel, the link system, the drive system, and other components" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003607_978-3-642-23026-4_10-Figure10.6-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003607_978-3-642-23026-4_10-Figure10.6-1.png", "caption": "Fig. 10.6 A pressure device inside a fluid-filled vessel. The pressure is measured by the relative position of a movable piston in the device", "texts": [ " The reason for lumping liquids and gases together and calling them fluids is because neither liquids nor gases (such as liquid water and steam, for example) have a fairly rigid three-dimensional array of atoms/molecules as compared to solids (such as ice, for example). In contrast to solids, fluids can flow and conform to the boundaries of any container in which they are placed. This is because a fluid cannot sustain a force that is tangent to its surface. In the language of the previous section, a fluid flows because it cannot withstand a shearing stress. On the other hand, a fluid can exert a force in a direction perpendicular to its surface. Figure 10.6 shows a pressure device inside a fluid-filled vessel. The device consists of a light piston of area A fitting in a vacuumed cylinder and resting on a light spring. As we insert the device into the fluid, the fluid will compress the piston due to the effect of a normal force of magnitude F\u22a5. Using Eq. 10.11, after replacing F\u22a5 by F, we define the average pressure exerted by the fluid on the piston by the following relation: The pressure at any point in the fluid is the limit of the above ratio as A of the piston, centered on that point, approaches zero" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001545_s1068366613040077-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001545_s1068366613040077-Figure2-1.png", "caption": "Fig. 2. Temperature field FT at uniform density qFP.", "texts": [ "T x f x dx F x x = \u2212 \u222b After the dimensionless parameter is intro duced, the average temperature is determined by the shape factor of the source as follows: (5) If the total density of heat generation (6) and the heat partition factor \u03b5* are known, then the density of the heat flow that propagates into the body is as follows: (7) Here, is the friction force, \u03bc is the coefficient of friction, P is the load, v is the velocity of the FMHS, is the area of contact with the length l and the width b of the source, is the pressure, and \u03c30 is the maximum value of normal stresses that have the distribution function . With allowance for formulas (5), (6), and (7), from (3), we obtain the following formula for the aver age contact temperature during friction that involves the shape factor of the FMHS: (8) For example, if the density of the source has the uniform distribution (Fig. 2), then = and x l \u03c8 = \u0422S av 1 0 ( ) .K F F d= = \u03c8 \u03c8\u222b c0q F S p\u00b5= = \u00b5v v ( )FP * 01 .q q= \u2212 \u03b5 F P\u00b5 = \u00b5 cS lb= c 1 0 0 ( )Pp f d S = = \u03c3 \u03c8 \u03c8\u222b ( )f \u03c8 S FP FP . K q alT = \u03bb \u03c0v const( )f \u03c8 = 1 0 1 T d F \u03c8 \u03c8 = \u03c8 \u2212 \u03c8 \u222b du u 0 \u03c8 \u222b 2 \u03c8,= S 1 0 42 . 3 K d= \u03c8 \u03c8 =\u222b 304 JOURNAL OF FRICTION AND WEAR Vol. 34 No. 4 2013 KRAVCHENKO et al. Values of the function FT (4) and the shape factor KS (5) for the basic shapes of the FMHS are presented in Table 1. The integral (4) was calculated using the substi tution . The study of the adequacy of for mula (4) shows that it does not take piecewise con tinuous functions and negative values of the current variable" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003029_dscc2013-3998-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003029_dscc2013-3998-Figure1-1.png", "caption": "Figure 1. REMOTE-CONTROLLED HELICOPTER TESTBED", "texts": [ " Helicopters can transport heavy loads attached by a suspension cable. This technique is extremely useful in construction, forestry, human rescue, and supply transport operations. However, the load often swings back and forth under the helicopter. This effect can significantly degrade control of the helicopter, and makes it difficult to deposit the load in a desired location. To study the detrimental effects of load swing, and to investigate mitigating control techniques, the small-scale helicopter testbed shown in Figure 1 has been constructed [1]. Many important tasks for helicopters with suspended loads involve horizontal position control. For instance, during the Depart/Abort task [2], the helicopter begins in a stable hover, moves forward a specified distance, and resumes hovering over a target. This task requires control over the helicopter\u2019s longitudinal position. To control the simulated model helicopter, a feedback controller that mimics the behavior of human pilots can be implemented. The attributes of this \u201cvirtual pilot\u201d should be adjusted to match the dynamics of the model helicopter", " To mitigate the cost, time, and difficulty of using free-flying helicopters, many researchers have built mechanisms to eliminate helicopter degrees of freedom that are not required for the specific research 1 Copyright \u00a9 2013 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 01/29/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use question being investigated. For basic investigations of the swing of suspended loads, the helicopter must be allowed to translate in a horizontal direction, and rotate in a vertical plane along this direction. Figure 1 shows a testbed that was constructed to fulfill the requirements for suspended load research. The helicopter is attached to a pivoting base suspended between two carts. The carts translate horizontally along two guide rails and allow the helicopter to pitch forward and backward. A load with adjustable weight and cable length is suspended from the helicopter. The load swings in the vertical plane, which adds a third degree of freedom to the system. The testbed uses a computer and microcontroller to read sensor inputs, perform control calculations, and generate commands" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001162_amm.380-384.82-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001162_amm.380-384.82-Figure1-1.png", "caption": "Fig. 1 Pressure and temperature distribution of journal bearing", "texts": [ " The traditional hydrodynamic lubrication model generally assumed constant temperature and ignored the viscosity change. The pressure distribution solved most easily in this case, so a lot of achievements of the bearing performance obtained on isothermal conditions [4]. However, during the practical operation, because of the compression and shear of the lubricating oil, the bearing generates friction power. And the higher is the speed, the larger is the friction power and heat. Finally, it forms a stable, uneven pressure and temperature fields, as shown in Fig. 1 [4]. Therefore, in high-speed conditions, the results of traditional model are not accurate enough, and can\u2019t meet practical needs. All rights reserved. No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of Trans Tech Publications, www.ttp.net. (ID: 132.174.254.159, Pennsylvania State University, University Park, USA-22/05/15,12:12:11) Based on hydrodynamic lubrication theory, considered the influence of temperature on the viscosity of the lubricant internal the bearing and combined generalized Reynolds equation, energy equation and viscosity-temperature equation, a high-speed journal bearing theoretical model is established and the oil film temperature and pressure distributions are obtained by using the finite difference method" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001430_s12206-012-1266-x-Figure6-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001430_s12206-012-1266-x-Figure6-1.png", "caption": "Fig. 6. Loosening process [1].", "texts": [ " The contact stresses on the lock nut were also roughly less than 1 MPa. The stresses were also too small to initiate yielding. Since the wheel shaft was supported by the not-preloaded bearings, it rotated severely in plane against the carrier housing. The lock nut slid by 0.44 mm on the carrier face, as shown in Fig. 5(c). Compared to the first case with an initial gap of 1.0 mm, the slippage increased to be 2.4 times larger. The slippage is seen to be related to the loosening. The loosening process in Fig. 6 is very similar to the present loosening failure. The lowered annular surfaces were observed with a lowresolution USB digital microscope. The height, i.e. the radial thickness, is 3.5 mm. Many circular scratches were easily observed. It seems as if they were stacked one after the other along the circumferential direction. The diameter of the arc circle was measured by drawing a circle on the photograph. It was about 1.0 mm, as shown in Fig. 7. The circular scratch formation process was investigated" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001255_icrms.2011.5979461-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001255_icrms.2011.5979461-Figure1-1.png", "caption": "Figure 1. Diagram of planetary gear assembly", "texts": [ " The results showed that the adhesive wear was the main wear type for thrust washer while the fatigue wear was a small part; a great axial force existed in the working mechanism, which lead to big contact stress and speed up wear; the high temperature on the friction surface could reach about 700 and cause the tempering process in the material; Inadequate cooling effect due to lack of lubrication oil may be another main reason for the high temperature caused by wear. Keywords- bearing outer ring; wear failure; axial force; high temperature; lubrication I. INTRODUCTION The power shift steering transmission equipped on tracked vehicles can integrate gearshift and steering together based on power split principle, the two power flow of which are integrated to the planetary gear train and then output to the wheels. The wear between the thrust washer and spur planet gear (as shown in Fig.1) is a common failure mode, which can accelerate the failure process of bearings, gears and other components, especially under high-speed conditions. Therefore, planetary gear bearings' wear characteristics along with the reason should be investigated. Through failure analysis of parts, the type of its failure can be determined, the reasons for its failure can be identified, and then preventive measures could be taken to reduce or prevent similar failures occurred in the product design life. In China, very little publication [1,2,3] about the roller bearing in spur planetary gear can be found. The needle roller bearing of planetary gear in power-shift steering transmission has been studied [4,5,6], but there is not a comprehensive analysis of the wear characteristics and failure reasons. The analysis of the failure mechanism and reasons based on the change of the material of bearing was presented in this paper. II. WORKING ENVIRONMENT As shown in Fig.1, Planet gear is supported by the needle roller bearing with outer ring, which is installed on the axle of the planet gear. The outer ring of the needle roller bearing is assembled in the hole of planet gear with interference fit. The 978-1-61284-666-8/11$26.00 2011 IEEE lubrication oil pressured (oil pressure 0.10 0.15MPa) flows into the inner cavity of the planet gear axle along the lubrication pipe shown in Fig.1, and then flows through two holes to lubricate the needle roller bearing. The lubrication oil flows from the center of the bearing to the two ends, most of which flows out through the gap between the planet gear and thrust washer. Only a little oil will flow out through the gap between the thrust washer and the planet carrier. Material composition of bearing is GCr15. The working environment of the supporting components for the planet gear is serious. The relative speed range of the planet gear is 0 ~ 5000r/min, and the radial load of the planet gear axle is in a range of 0~ 33000N [4]" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001712_0954406212438142-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001712_0954406212438142-Figure2-1.png", "caption": "Figure 2. (a) A typical turbocharger\u2013\u2013assembly of compressor disc, turbine disc and two identical floating ring bearings. Note that the distances la and lb are not shown to scale for clarity. Since the turbine is much heavier than the compressor, the centre of gravity is close to the turbine bearing centre. (b) A typical floating ring bearing with oil holes which supply oil from the outer-film to the innerfilm. The outer-film is between the bearing and the ring, and the inner-film is between the ring and the rotor (Cummins turbotechnologies Ltd).", "texts": [ " In Figure 1(b), the waterfall plot shows only the second mode for the same speed range. Tondl10 suggested that gyroscopic action could reduce the amplitude of self-excited vibration and for a fairly large gyroscopic moment, it may even suppress it completely. The aim of this article, therefore, is to investigate the effect of the gyroscopic moment on the conical whirl mode instability in a turbocharger by developing the simple model described by Holmes, et al.9 Gyroscopic effect in a symmetric rotor-bearing system Figure 2 shows a typical automotive turbocharger assembly. Since the steel turbine disc is much heavier than the aluminium compressor disc of a turbocharger, the centre of gravity of the turbocharger is near to the turbine bearing centre. Hence the turbocharger is modelled as an asymmetric rotor. In a floating ring bearing such as that shown in Figure 2, a ring separates the oilfilm of the bearing into an inner-film and an outer-film. The ring usually rotates but in some configurations, the ring is pinned to the bearing and not allowed to rotate. In such cases, the ring experiences a precession. This is the case considered in this article. The inner-film of the floating ring bearing can be modelled as a hydrodynamic bearing and the outer-film can be modelled as an external damper in series with the bearing.2,9 The stiffness of the outer-film is assumed to be negligible compared to the damping in the squeezing action of the ring on the outer-film", " \u00f0r1a ra\u00de \u00fe \u00f0r1b rb\u00de\u00bd \u00bc 0 \u00f021\u00de The constraint forces at the interface of the ring and the outer-film (external damper) are given by f \u00bc r\u0302 \u00bc _r \u00f022\u00de where r\u0302 \u00bc Rr,Rsf gT; Rr, s are the constraining forces applied by the outer-film that is the external damper and is the damping coefficient of the outer-film of the floating ring bearings. Combining equations (16) and (17) with equation (22) results in _ra 2A\u00f0_r1a _ra\u00de A!!\u00f0r1a ra\u00de \u00bc 0 \u00f023\u00de _rb 2A\u00f0_r1b _rb\u00de A!!\u00f0r1b rb\u00de \u00bc 0 \u00f024\u00de Combining equations (21), (23) and (24) and assuming solutions of the form r \u00bc Re t, gives the non-dimensional matrix equation where, La, b \u00bc la, b=l. Figure 9(a) shows the real part of the root of the characteristic equation of the determinant of equation (25). For a typical turbocharger shown in Figure 2(a) using La \u00bc 0:01,Lb \u00bc 0:99, for 5 1=2, the system has an unstable conical mode and an unstable cylindrical mode with positive Re sf g. However, the real part of the conical mode becomes zero for \u00bc 1=2. For 4 1=2, the system has a stable conical mode, with the Re sf g being negative. The stability of the in-phase whirl mode remains unaffected since does not directly affect the transverse motion. The imaginary part of the unstable roots for different values of is shown in Figure 9(b). This plot represents an overlay of theoretical waterfall plots for three different values of " ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000261_amr.317-319.1896-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000261_amr.317-319.1896-Figure1-1.png", "caption": "Fig. 1 The basic components of BT40 tooling system and its real object", "texts": [ " Secondly, based this FEM model of tooling system, the paper investigated the changing law which length and diameter of tooling system influence to structural modal parameters. After that, by means of vibration testing of machine raising speed idle running, further verified the FEM model and its accuracy, the researched results can provide a theoretical basis to evaluate the dynamic performance other type\u2019s tooling system. Analysis of the dynamic performance of tooling system with combining FEM and EMA Based on FEM the analysis of the dynamic performance of tooling system. As illustrated in Fig.1\uff0cwith BT40 tooling system as a research object, jointing cutter clamp and tool holder together, the cutter is simplified to be a cutter bar with the diameter at 12 mm, clamping length at 60 mm, and the contact length between tool holder and cutter bar is at 20 mm. Set up 3D parameterized model of this tooling system firstly, and then set up its FEM assembly model which is treated as follows: All rights reserved. No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of Trans Tech Publications, www", " Below is only calculated various structural modal parameters without constraints when the structural dimensions of BT40 tooling system are changed. The changing rule of tooling system structural modal parameter with various clamping lengths. The clamping length of cutter bar is a key factor which influences tooling system dynamic performance. When cutting on some deep-cavity part surfaces, the lengthened cutter bar must be used\uff0cso it is very necessary to research the changing law that the length of cutter bar influences on structural modal parameter. On the basic model of BT40 tooling system which showed in Fig.1, not changing the diameter of cutter bar, but in turn change its effective clamping lengths to be 80 mm, 100 mm, 120 mm and 150 mm, and calculate their first 3-order structural modal parameters and modal shapes. The above results can be illustrated in Fig.4, and it is concluded logically as follows\uff1a (1) Along with the length of cutter bar increasing in a certain range, the overall trend of each order natural frequency of tooling system becomes smaller, and the modal shape of each order remains unchanged", " (3) Affirmatively, when the length of cutter bar increased to a certain degree, its first order natural frequency is already very low, so it is easy be close to the rotary frequency of spindle system\uff0cat this time, it is very difficult to adapt to the normal machining needs, in particular, to the high speed cutting needs. The changing rule of tooling system structural modal parameter with various tool diameters. The diameter of cutter bar is also an important factor of tooling system designing, and the smaller the diameter is, the smaller material mass and rotary inertia are. But the influence of diameter of cutter bar to the tooling system dynamic performance needs a further research. On the basis model of BT40 tooling system which showed in Fig.1, not changing the length of cutter bar, but in turn change the diameters to be 5 mm, 8.5 mm, 16 mm, 20 mm and calculate their first 3-order structural modal parameters and modal shapes. The above results can be illustrated in Fig.5, and it is concluded logically as follows: (1) Along with the diameter of cutter bar increasing, the overall of trend of the mass, inertia moment and the first order natural frequencies of tooling system is increasing, but not in a linearly proportional relation\uff0cand when the diameter increased to 15 mm, the increase of each order natural frequency is moderate" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003240_978-0-8176-8370-2-Figure8.7-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003240_978-0-8176-8370-2-Figure8.7-1.png", "caption": "Figure 8.7: As in Figure 8.3, with G3 = 0.2 and B2 = 0.04, but with a non null electric field E3 = 0.01.", "texts": [ "1), we get the Stark\u2013Quadratic\u2013Zeeman-parallel (SQZp) problem: HSQZp = 1 2 p2 \u2212 1 q +B(q1p2 \u2212 q2p1)+ 1 2 B2(q21 + q22)+E3q3. Clearly, the two angles \u03d1S and \u03d1D are still null and the system has axial symmetry. The problem has been investigated in Deprit, Lanchares, I\u00f1arrea, Salas & Sierra (1996), Salas, Deprit, Ferrer, Lanchares & Palaci\u00e1n (1998), and Salas & Lanchares (1998) but with an analytical method. The main difference with respect to the previous QZ case is that the c coefficient in (8.1.3) is now present, so the intersection parabolas with the plane \u03be1\u03be3 are shifted along the \u03be1 axis. In Figure 8.7 the G-case is shown. In Figure 8.8 (left) we show the distribution of the FMI in the presence of an electric field, with E = 0.06 and B2 = 0.04. The computation has been performed for the full range, i.e., \u22121 \u2264 G3 \u2264 1, for the asymmetry of the problem. The rather small values of the perturbation show the power of the FMI method. The resonances are clearly recognized, and magnifying further details, as in Figure 8.8 (right), allows us to penetrate the structure 246 Some Perturbed Keplerian Systems very deeply" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000373_1.4806816-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000373_1.4806816-Figure3-1.png", "caption": "FIGURE 3. Simulation of Extrusion Process of the Microchannel Tube", "texts": [ " Number of the total elements of the simulation model is 150 thousand. The computation parameters are set as follows. Time step of the computation is 0.01s and 0.003s at the dividing and welding stage respectively. Material of the billet is aluminum alloy AA3003. Extrusion ratio is 140. Extrusion speed is 5mm/s. Temperature of the billet and the die are 450 C and 420 C, respectively. Billet is defined as a thermal-plastic model and the others are defined as rigid. The simulation of extrusion process of the micro-channel tube is shown in Fig.3. For the porthole die in this study, the welding plane of the micro-channel tube is at the horizontal section of the tube. Fig. 4 shows the bonding processes of billet inside the welding chamber. As depicted in the pictures from (a) to (d), the separated metals gradually contact at the welding plane and bonding to each other at certain temperature and pressure. It can be seen that the welding chambers of the channels at both sides are filled first, while the middle welding chambers are hard to be filled because of the close-packed arrangement of the mandrels" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000558_amr.472-475.74-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000558_amr.472-475.74-Figure3-1.png", "caption": "Fig. 3 Formation of the imaginary rack-cutter surface.", "texts": [ " No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of Trans Tech Publications, www.ttp.net. (ID: 141.211.4.224, University of Michigan Library, Media Union Library, Ann Arbor, USA-05/07/15,04:54:14) Derivation of face-gear tooth surfaces Applied coordinate systems. Coordinate systems applied for generation of face-gear is shown in Fig. 2. Fixed coordinate systemsSm1, Sq are rigidly connected to the cutting machine. Movable coordinate systemsS1 and St1 are rigidly connected to the face-gear and the head-cutter, respectively. Mathematical model for head-cutter surface. Fig.3 depicts the formation of the imaginary rack-cutter surface and the relations between coordinate system Sg and St0. The mathematical model of the generating cone for gear straight-line head-cutter is proposed in [6]. Equations of generated face-gear tooth surface. The face-gear surface is determined as the envelope to the family of head-cutter surfaces as follows: ),(),,( 1111 \u03b4\u03d5\u03b4 SrMMMSr cctqcq= \uff081\uff09 0),,( 1 )1( 1 ==\u22c5 \u03d5\u03b4Sfvn c c \uff082\uff09 Here ),(1 \u03b4Src is the vector function that determines the surface of the head-cutter in coordinate system St1; M1q, Mqc and Mct1 describe the coordinate transformation from Sq to S1, Sc to Sq, St1 to Sc; )1(cv are the relative (sliding) velocity" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001045_icma.2013.6618057-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001045_icma.2013.6618057-Figure1-1.png", "caption": "Fig. 1. Human demonstration system.", "texts": [ " In section III, how to obtain impedance parameters and equilibrium trajectories is presented. Finally, an application of the proposed method is shown in section IV. II. TASK SKILL TRANSFER MET HOD In this section, the task skill transfer method is described. The method is divided into two processes; human demonstra tion and task skill acquisition. In the following, each process is explained in detail. A. Human Demonstration As mentioned earlier, a bilateral teleoperation system is employed to acquire reusable task skills. The overview of the system is illustrated in Fig. 1. The master and slave arms have 6-DOF at the tips, and both the arms are able to measure the tip forces and poses (position and orientation) . The control laws of both the arms are represented by Xm = Xs = A (f m + fs) , (1) where xm E \ufffd6 and Xs E \ufffd6 are the linear and angular velocities at the tips of the master and slave arms, respec tively, A E \ufffd6 x 6 is a positive-definite admittance matrix, and f m E \ufffd6 and f s E \ufffd6 are the force and moment applied to the tips of the master and slave arms, respectively", " The target task is the nut attachment task. A screw nut held by the slave arm is fastened to the bolt fixed to the ground, as shown in Fig. 3. The base coordinate system \ufffdb and nut coordinate system \ufffdn are placed as shown in the figure. The initial position of the origin of \ufffdn was set to [0 , 0 , 0.02f (m) with reference to \ufffdb. The initial orientation of \ufffdn relative to \ufffdb was set to [0 , 0.15 , of (rad) , where the orientation is represented by roll-pitch-yaw angles. We developed a bilateral teleoperation system illustrated in Fig. 1. A compact 6-DOF haptic interface, developed at Tohoku University [18][19], was utilized as the master arm, and a 7-DOF robot manipulator PA-1O was used as the slave arm, as shown in Fig. 4. The control frequency of the system was 300 Hz. The ad mittance matrix A in (1) was determined based on preliminary experiments, which was where AL = (7.5 x 10-4) 13 [mJ(Ns) ] AA = (7.5 x 10-2) 13 [rad/(Nms) ] where 13 E \ufffd3 x 3 is the identity matrix. A human operator performed the nut attachment task. Dur ing the task execution, the position and orientation of the nut with reference to the base coordinate system plus the force and moment applied to the nut were recorded" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001457_kem.480-481.306-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001457_kem.480-481.306-Figure1-1.png", "caption": "Fig. 1 A grade B cast steel wheel [13]", "texts": [ " In the contrast, most of researches used elastic or monotonic constitutive data. Reasonability can not be met unless a material has a same stress-strain relation for both monotonic and cyclic loads. China is developing high-speed and heavy hauled railway industry. Reliability and safety are a primary task. It is a basic work to investigate wheel material cyclic deformation behaviour to ensure reasonability of fatigue analysis. Present work investigates on China railway grade B cast steel wheel. As shown as in Fig. 1, a grade B cast steel wheel was fabricated by a metallurgical process first smelted and modeled in electric furnace with a control of temperature gradients to eliminate residual stress field formation, and then, moved into auto-controlled cooling furnace with a temperature decrease gradually and following a sequence of heat or machining treatments including rim quenching and tempering and web shot peening. Previous chemical composition and physics property inspections have verified that [13]: the present wheel shows a slight higher chemical composition for All rights reserved" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000823_amr.221.165-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000823_amr.221.165-Figure3-1.png", "caption": "Fig. 3 Constraints and loads of radial tire model touching ground", "texts": [ " Only the elements of carcass cord fabric were frozen to minus 100 centigrade (marked in blue), while other parts were maintained at zero centigrade (marked in red). ANSYS\u2019s coupling of thermal and structural analysis shows that the reduced temperature only made the material shrink from cold, and that other functions were not influenced by controlling the parameters in the simulation. In real applications, tires are facially symmetric when at rest or moving in a straight line. Therefore, a half-tire FE model was established as shown in Fig. 3. The model can simulate the deformation, stress and stain, distribution of pressure and force of friction in the contact area when the tire bears an upright load and level traction force. The constraints and loads are the same as those in section 2.1 with the exception of the vertical load. In the area touching ground, a plane was used to simulate the road (as shown at the bottom of Fig. 3). The plane was pushed upward 20mm to simulate the effect of car weight. The parameters for each tire part should be defined before performing the FEM simulation. According to the previous experiment [12], the tire parts were separated into two types: isotropic and anisotropic. The name of each part is shown in Fig. 4. The test value of each part used in the simulation is shown in Tables 1, 2 and 3. Conclusion In this part, the three-dimensional radial tire models, FE models of free and ground-touching tires, were built and showed in detail" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001402_s13369-012-0232-3-Figure4-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001402_s13369-012-0232-3-Figure4-1.png", "caption": "Fig. 4 The determination of the gear with convex teeth", "texts": [], "surrounding_texts": [ "Figures 3 and 4 define the coordinate systems for obtaining the generating surface. Two fixed reference coordinate systems Sn(On, xn, yn, zn) and Sr(Or, xr, yr, zr) are rigidly connected to the frames of the concave gear and convex gear, respectively. A fixed coordinate system Sf(Of , xf , yf , zf) is rigidly connected to the frame. Coordinate system Sc(Oc, xc, yc, zc) is attached to the conical surface. The surface translates over the linear displacement S along the yf -axis. The coordinate system S2(O2, x2, y2, z2) rotates at angle \u03c62, and S1(O1, x1, y1, z1) rotates at \u03c61 and about the xn-axis and xr-axis, respectively. The angle \u03b1 is a rotary angle. Its region is 0 \u223c 2\u03c0 . To generate an envelope to the family of the ring surfaces, a coordinate transformation matrix from the Sc coordinate system to the S1 coordinate system can also be used. By applying homogeneous coordinates matrix [7] for coordinate transformation, the matrix M1c, can be obtained as follows: M1c(\u03b1, \u03c61) = \u23a1 \u23a2\u23a3 cos \u03b1 sin \u03b1 cos \u03c61 sin \u03b1 sin \u03c61 \u2212rp1\u03c61 sin \u03b1 cos \u03c61 + rp1 sin \u03b1 sin \u03c61 \u2212 sin \u03b1 cos \u03b1 cos \u03c61 cos \u03b1 sin \u03c61 \u2212rp1 cos \u03b1 cos \u03c61 + rp1 cos \u03b1 sin \u03c61 0 \u2212 sin \u03c61 cos \u03c61 rp1 sin \u03c61 + rp1 cos \u03c61 0 0 0 1 \u23a4 \u23a5\u23a6 (4) where rp1 represents the standard pitch radii of the convex gear. Similarly, matrix M2c is the coordinate transformation matrix from Sc to S2. Thus, matrix M2c can be written as follows: M2c(\u03b1, \u03c62) = \u23a1 \u23a2\u23a3 cos \u03b1 sin \u03b1 cos \u03c62 \u2212 sin \u03b1 sin \u03c62 \u2212rp2\u03c62 sin \u03b1 cos \u03c62 + rp2 sin \u03b1 sin \u03c62 \u2212 sin \u03b1 cos \u03b1 cos \u03c62 \u2212 cos \u03b1 sin \u03c62 \u2212rp2\u03c62 cos \u03b1 cos \u03c62 + rp2 cos \u03b1 sin \u03c62 0 sin \u03c62 cos \u03c62 \u2212rp2\u03c62 sin \u03c62 \u2212 rp2 cos \u03c62 0 0 0 1 \u23a4 \u23a5\u23a6 (5) where rp2 represents the standard pitch radii of the concave gear. Using Equations (1)\u2013(5) and the coordinate transformation matrix Mic(\u03b1, \u03c6i ), the family of the ring surfaces can be expressed by: Rg i (\u03b2, j , \u03b1, \u03c6i ) = Mic(\u03b1, \u03c6i )R g c(\u03b2, j ) (6) where \u03b1 and \u03c6i are independent parameters calling the parameters of motion. The variable \u03b2 and j are the design parameters as represented in Sect. 2 ( j = c, d, h). Vector Rg i (\u03b2, j , \u03b1, \u03c6i ) is the family of the ring surfaces. Superscript g is ab, bc, and cd . Subscript i is 1 and 2. The position vector of the ring surface is Rg c, as indicated in Equations (1)\u2013(3). If the family of ring surfaces is regular surface, one can find an envelope to a two-parameter family of surfaces [7]. The envelope to a two-parameter family of the ring surfaces should satisfy the equations: ( \u2202Rg i \u2202 j \u00d7 \u2202Rg i \u2202\u03b2 ) \u00b7 \u2202Rg i \u2202\u03c6i = 0 (7) ( \u2202Rg i \u2202 j \u00d7 \u2202Rg i \u2202\u03b2 ) \u00b7 \u2202Rg i \u2202\u03b1 = 0 (8) Substituting Equations (1), (2) and (6) into Equations (7) and (8) with i = 1, an envelope 1 to the family of the ring surfaces is obtained. This envelope 1 is called the convex tooth of the gear. Substituting Equations (2), (3) and (6) into Equations (7) and (8) with i = 2, an envelope 2 to the family of the ring surfaces is obtained. The envelope 2 is called concave the tooth of the pinion. Substituting Equations (1)\u2013(6) into Equations (7) and (8), the equations of meshing can be written as: \u03c6i = (yi c yi c\u03b2 + zi czi c\u03b2)xi c j \u2212 (yi c yi c j + zi czi c j )xi c\u03b2 rpi (xi c j yi c\u03b2 \u2212 yi c j x i c\u03b2) (9) \u03b2 = 0 (10) The dimensional parameters of the gear mechanism are listed in Table 1. A computer program was employed to draw the complete profile of the gear with convex tooth and the pinion with concave tooth. Using Equations (1), (2), (6), (9) and (10) with i = 1, the first tooth of the gear was created. The other teeth were copied from the first tooth and rotated at its axis by 2cy\u03c0/N1, cy = 1, . . . , N1, where N1 is the number of teeth of the gear. The complete contour of the gear is shown in Fig. 5a. Using Equations (2), (3), (6), (9) and (10) with i = 2, the first tooth of the pinion was created. The other teeth were copied from the first tooth and rotated at its axis by 2cx\u03c0/N2, cx = 1, . . . , N2, where N2 is number of teeth of the pinion. The complete contour of the pinion is shown in Fig. 5b. The assembly model of the pinion and the gear is shown in Fig. 5c. Using the proposed mathematical model and rapid prototyping technology, the pinion and the gear created by a ring surface were obtained and shown in Fig. 5d." ] }, { "image_filename": "designv11_101_0002753_s0885715613000857-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002753_s0885715613000857-Figure3-1.png", "caption": "Figure 3. (a) 3D image rendering of the inner channel structure of Ni Blade sample; (b), (c) Filtered Backprojection tomographic reconstructions of transverse sections across the blade.", "texts": [ "S2, September 2013. the sample in steps of 0.1\u00ba. Parallel monochromatic beam was used, and 2.75s exposure time per image was chosen to match the 16-bit dynamic range of the detector. The 3D volume reconstruction was performed using filtered back-projection algorithm (iradon) implemented in Matlab \u00ae . The resulting reconstruction was rendered by introducing an intensity threshold, applying Sobel filter, median filter, and masking of the blade using edge detection approaches. The results are illustrated in Figure 3. The inner structure of DMLS-built blade revealed in Figure 3 provides improved insight into the dimensional accuracy and surface quality of the objects produced by the DMLS process. Figure 3(c) reveals the presence of internal pores (dark areas), particularly in the area close to the blade trailing edge. Furthermore, the rough nature of the internal surfaces produced by DMLS can be resolved. B. Diffraction Setup The flexible setup of I12 JEEP allows smooth changing between imaging and diffraction experiments, while staying in the monochromatic mode. The imaging and diffraction detectors are Nikolaos Baimpas et al. S440S440 Vol. 28, No.S2, September 2013. mounted on separate moving translation stages", " Calibration diffraction patterns were acquired at 15\u00ba increments from 0\u00ba to 180\u00ba. The movement of the centre of the (Obrist A. ) Debye-Scherrer ring, together with the pattern scaling due to the changing calibrant position together provided calibrated a means of obtaining the information about the circular orbit of the calibrant on the rotation stage. The complex shape of the blade cross-section posed a number of previously unaddressed problems with respect to the data acquisition, as well as calibration and analysis. The lade cross-sectional shape illustrated in Fig.3(b,c) corresponded to the approximate envelope of ~22mm(width) \u00d7 ~4mm(depth). The sample absorption at angles close to 90\u00ba rotation caused the diffraction signal to become extremely weak. The intense absorption phenomenon is also apparent in the absorption tomography sinogram (Fig.4a) in the angular region 60\u00ba-100\u00ba. In the integrated diffracted intensity sinogram in Fig.4(b) this phenomenon manifests itself by the blue \u201cblade-shaped\u201d blue region of low intensity. The diffraction patterns (Debye-Sherrer rings) registered on the detector were analysed using Fit2D (2D crystallography diffraction data analysis software, ESRF) by \u201ccaking\u201d and binning", " There is little evidence of residual strain variation in the slices at 0\u00ba and 180\u00ba, so that after correction the two lines are almost flat and lie on top of each other. This also provides the opportunity to assess the level of noise in the strain measurement (~2\u00d710-4); and to conclude that, within the accuracy of the measurement, the first strategy results in a correction procedure that overcomes fully the geometric aberrations caused by the experimental setup (absence of secondary collimation). The second correction strategy adopted was to follow the tip of the blade located at the right hand side in Fig. 3b through the sequence of angular positions where it is not obscured by the blade bulk (from 0\u00ba to 60\u00ba and 120\u00ba to 180\u00ba). Based on the assumption that the measurement always interrogates the same gauge volume at different rotation increments, the strain deduced after the correction should remain constant. Any remaining errors in the SDD correction are bound to have an acute effect on The application of geometry corrections for Diffraction Strain Tomography (DST) S445S445 Vol. 28, No.S2, September 2013" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000176_j.proeng.2013.07.006-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000176_j.proeng.2013.07.006-Figure3-1.png", "caption": "Fig. 3. Experimental setup: (a) schematic; (b) inside wind tunnel test section", "texts": [ " The diameters of these two balls are approximately 90 mm and 98 mm respectively. The DeMarini softball has slightly larger diameter than the Thunder SY softball. However, these two balls possess the same 88 pair of stitches. The DeMarini ball has slightly higher and wider seams than the Thunder SY ball. The frontal views and their seam orientations of two softballs are shown in Fig 2. In order to investigate the aerodynamic properties experimentally, a support system made of a sting with an angle adjustment mechanism (as shown in Fig 3) was developed to hold the baseball and softball on a force sensor in the wind tunnel. The distance between the bottom edge of the ball and the tunnel floor was 400 mm, which was well above the tunnel boundary layer and out of the ground effect completely. An aerofoil (fairing) with the same height as the sting was positioned around it (see Fig 3). The aerofoil shaped fairing would protect the sting from the incoming wind load and greatly enhance the accuracy of the results since the previous experimental setup [4] overestimated the magnitude of the drag force when compared to published data. The ball was connected through a mounting sting with a JR3 multi-axis load cell (also commonly known as a 6-degree of freedom force-torque sensor made by JR3, Inc., Woodland, USA). A purpose made computer software was used to digitize and record all three forces (drag, side, and lift forces) and three moments (yaw, pitch and roll moments) simultaneously", " It may be noted that the transition to fully turbulent flow for Rawlings Major League ball (with lower seam height) occurs at slightly higher speeds compared to Rawlings NCAA Champion ball with higher seam height. The minimal difference in CD values for higher and lower seam height baseballs after transition indicates that the local flow separation due to seams is minimised or fully eliminated. The effect of seam and stitches are highly evident at low speeds as the local flow separation is present due to seams, stitches and their complex orientations. The DeMarini softball displays the lowest CD value compared to Thunder SY softball as well as baseballs (see Fig 3). The flow transition for both softballs starts later at 65 km/h compared to 40 km/h for baseballs and becomes fully turbulent at 120 km/h. The variations of CD values among all four seam positions (1, 2, 3 and 4) as shown in Fig 2 and Fig 3 are evident at all Reynolds numbers tested for all 3 baseballs and 2 softballs, however, the variations between position 1 and 2, and position 3 and 4 are minimal as these two positions are considered to be the mirror image. Additionally, the CD variations among four positions for each ball are evident at low Reynolds number (below 40 km/h), however, these variations are minimal at high Reynolds numbers (Re = 1.6 \u00d7 105 or above) which is believed to be due to the elimination or minimization of local flow separations from seams" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000363_msf.770.141-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000363_msf.770.141-Figure2-1.png", "caption": "Figure 2 Parameters of displacement matrix of rigid body", "texts": [ " Principle and procedures of variable forming point grinding method In this method, as shown in Figure 1, select a point A on end face of cup wheel as Forming point, make the end face of cup wheel maintain tangency contact with workpiece at Forming point through adjusting the axis of grinding wheel. The most important of this method is that with time and cutting amount varying, Forming point changes from point A to 'A gradually at un-uniform speed, which can always keep Forming point un-worn and increase suface processing quality As shown in Figure 2, the axis and displacement of grinding wheel are adjusted following the steps below: (1) install work-piece and grinding wheel on a three-axis machine tool that has three degrees of freedom in X-Z plane, 2 for pure translation and 1 for pure rotation, and then ascertain tool setting point that work-piece vertex p1 corresponds to . (2) the vertex p1 is the first cutter contact point of tool path, the other points as well as its corresponding angles between on tool path can be determined by constant scallop height method" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002004_iros.2011.6094766-Figure6-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002004_iros.2011.6094766-Figure6-1.png", "caption": "Fig. 6. Cylindrical objects used in the real experiment.", "texts": [ " At this time, the controller calculates the contact point at that position. After that, a tracking algorithm was applied to make the fingertip slide and roll on the object. During the movement, the contact force was still in a constant range. At the final contact point, the contact force was regulated again to estimate the final point of contact. To check the effect of the trajectory on the accuracy of the method, we examined various trajectories. Also, several cylinders with different radii were tested. The size of these cylinders can be seen in Fig. 6. Using the soft finger contact model of Bicchi, we calculated the contact position on the object [14]. From that, we could calculate the change in the contact points during the tracking process. The rotational angle of the fingertip was 150, and the sliding motion was estimated by Eqs. (10) and (22). The trajectory s can be seen in Fig. 7, the left curve is the raw trajectory, and the right curve is the smooth trajectory. The length of the raw trajectory was about 14.6288mm. The radius of the cylinder was calculated to be 207" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000035_amm.373-375.1406-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000035_amm.373-375.1406-Figure1-1.png", "caption": "Fig. 1, The quadrotor schematic.", "texts": [ " Some micro-UAVs spring up in a variety of civilian domains, e.g., traffic monitoring, natural risk management, environmental protection, etc. [3] . As a part of the UAVs, the quadrotor has attracted considerable from institutions and universities throughout the world. The quadrotor is a mechatronic system with four rotors in a cross configuration. Similar to the conventional helicopter, the quadrotor is powered by propellers. To combat the torques produced by the four propellers, their rotation directions are distributed skillfully. As shown in Fig. 1, while the front and the rear motor rotate clockwise, the left and the right motor rotate counterclockwise which nearly counteracts gyroscopic effects and aerodynamic torques in trimmed flight [4] . The quadrotor flight control mechanism is shown in Table 1. The vertical and lateral motions can be fulfilled only by collectively increasing or decreasing the speed of the single propeller. When all the propellers spin at the same speed, the angular momentum is in balance in all the axes and the total thrust is the sum of the thrusts generated by the four single propellers [5] ", " As the most papers [11,12] , aerodynamic effects and rotor dynamics are ignored in this paper, which is justified by the fact that the quadrotor has a small airframe, fly at relatively low speeds, and has small propellers [13] . To further consideration of modeling, some other papers [14, 15] take gyroscopic effects into account. Before the derivation of the EOM (Equations of Motion), two reference frames need to be defined. Let I Oxyz denote an earth-fixed inertial frame and b b bB Ox y z a body-fixed frame whose origin O is at the center of mass of the quadrotor as shown in Fig. 1. The rotation matrix from B to I is expressed in Eq. (1): R R R R 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 (1) Where [ , , ]T denotes the vector of three Euler angles and s and c are abbreviations for sin(.) and cos(.). The derivation of the quadrotor dynamics model is performed in the body-fixed frame B . The equations of motion for a rigid body of mass 1 1m R and inertia 3 3J R subject to external torque 3 1M R and force 3 1F R are given by the Newton-Euler equations: J J M (2) mV mV F (3) Where [ , , ]Tp q r and [ , , ]TV u v w are, respectively, the angular and linear velocities expressed in the body-fixed reference frame" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000690_icems.2011.6073983-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000690_icems.2011.6073983-Figure3-1.png", "caption": "Fig. 3. No-load field distributions of three DSEGs", "texts": [ " FEA SIMULATION In order to verify the feasibility of the new DSEG elements above, 2-D Finite Element Analysis (FEA) models were developed by ANSYS Parameter Design Language (APDL). The static characteristics of 24/16, 12/16, and 6/16-pole DESGs are considered here, and they are combined by 6/4- pole DSEG element, 6/8-pole DSEG element and 6/16-pole DSEG element respectively. For comparison, the three DSEGs have the same main parameters, such as rotational speed, phase winding turns and main dimensions, which are shown in Tab. 1 in detail. The non-load field distributions from the FEA models are also shown in Fig. 3. Although each field winding element of DSEG has the same turns, the field winding elements equipped in 24/16-pole DSEG are two times than that of 12/16-pole DSEG, and four times that of 6/16-pole DSEG. So the theoretical value of field winding self-inductance of 24/16-pole DSEG is two times that of 12/16-pole DSEG, and four times that of 6/16-pole DSEG. The simulation results shown in Fig. 4 indicate that the average value of field winding self-inductance of 24/16-pole DSEG is 2.02 times that of 12/16-pole DSEG, and 3", " Due to the edge effects between the stator pole and rotor pole, the field winding self-inductance fluctuates a little. And the fluctuation value of the 6/16-pole DSEG field winding self-inductance is the most small. Fig. 4. Field winding self-inductances curves Fig. 5. Cogging torque curves Fig. 5 illustrates the cogging torques of the three DSEGs from simulation. The cogging torque peak value of 6/16-pole DSEG is 0.64N m (accounts for 1.67% of rated torque), which is approximately half of 12/16-pole DSEG, and a quarter of 24/16-pole DSEG. As shown in Fig. 3, the DSEG with less stator poles has the longer magnetic circuit of field winding in the stator and rotor yoke under the condition that the stator inner diameter keeps invariant. So the reluctance of 6/16-pole DSEG field winding is the maximum of the three, and the air gap flux density is the minimum. In order to acquire the same phase EMF, the lengths of the stator core are unequal. The stator core lengths of 6/16-pole DSEG, 12/16-pole DSEG and 24/16- pole DSEG are 136.4mm, 124.3mm and 119mm respectively" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000494_amr.301-303.1618-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000494_amr.301-303.1618-Figure2-1.png", "caption": "Fig. 2. The winding system of wire rope", "texts": [ "4 Experiment equipment The experiment equipment of multi-layer winding wire rope and pulley is used in this paper; this equipment can simulate the actual operation condition of crane's lifting and amplitude system. The equipment includes test stand and electric control system, which is centrally manipulated and controlled by operating board fixed in console cabinet. As shown in Fig. 1, the test stand consists of main winding device and auxiliary winding device, these two devices all include driving motors, reducer, double dogleg multi-layer winding drum, electric eddy dynamometer used to provide artificial load, etc. As shown in Fig. 2, wire rope winds around one side of drum, then winds through pulley block, and winds in another drum. In the process of test, wire rope's driving force is supplied by driving motors of winding equipment, and the load is provided by electric eddy dynamometer of winding equipment. 2.5 The setting of wire rope\u2019s single working cycle Set the wire rope\u2019s single working cycle as: a progress that the wire rope on the main drum begins to get involved from the start until the rope around the main drum reaches the end position of the volume around the drum, and then release back the wire rope to the starting position. The number of middle pulley is 7, as in Fig. 2. That is, in a single working cycle, wire rope bends 14 times through the pulley. 2.6 The setting of scrapping standard of wire rope Research [7] has shown that it\u2019s not scientific to replace wire rope by the same standard, since failure forms of wire rope vary with use conditions. In the experiment, considering the particularity of multi-layer winding, the replacement standard of wire rope is: 1\uff09More than 12.5% of the number of strand is broken in one-twist; or 2\uff09The rope of which diameter has been decreased by more than 7% of the normal diameter; or 3\uff09The rope which has been kinked; or 4\uff09In the drum section, when thrust occur by the upper lay to the lower lay " ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001680_ccdc.2013.6561389-Figure9-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001680_ccdc.2013.6561389-Figure9-1.png", "caption": "Fig. 9: Collision interval calculation diagram The collision range between the mechanical arm and the line-shaped obstacle is called the collision interval. Their computer method is same no matter whether P(z, x) has solution for the modeling of the line-shaped obstacle. Assuming [Pt, Pb] is a collision interval, the upper and lower critical collision angles of the point Pt and Pb are computed firstly. For PtPb , the head and tail values of the upper and lower critical collision angles are only computed. The compute results are shown in Fig.10. P(z, x) has the unique solution in Fig.10(a)(P1(z1=0.8m, x1=1m), P2(z2=1.2m, x2=0.6m)). P(z, x) has the double solutions in Fig.10(b)(P1(z1=1m, x1=1.38m), P2(z2=0.55m, x2=0.7m)). P(z, x) has no solution in Fig.10(c)(P1(z1=0.8m, x1=1m), P2(z2=1m, x2=0.6m)).", "texts": [ "8(b)) or P\u201d (Fig.8(c)) on the line-shaped obstacle. P1P2 (except for the cross points) in Fig.8(a), P\u2019P2 (except for the cross points) in Fig.8(b), P1P\u2019 (except for the cross points) in Fig.8(c) and P\u2019P\u201d (except for the cross points) in Fig.8(d) meet the formula (1), which are processed as collision between the mechanical arm and the line-shaped obstacle. The case is that the circular arc passed by the straightened mechanical arm and the line-shaped obstacle have not the cross point, as shown in Fig.9. They meet the formula (1) and can be processed as collision between the mechanical arm and the line-shaped obstacle. (a) 2013 25th Chinese Control and Decision Conference (CCDC) 2655 C-space By above knowable, the obstacles in C-space are expressed as the boundary function, based on which the obstacle avoidance path planning can\u2019t be done directly, thus the information of the boundary function need be described. At present, the information description method is of two main classes, the description based on skeleton and the description based on unit segmentation" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000258_icma.2012.6285740-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000258_icma.2012.6285740-Figure2-1.png", "caption": "Fig. 2 SJQ12 forearm two degrees of freedom myoelectric controlled", "texts": [], "surrounding_texts": [ "Speculated by the vibration mechanism of prostheses shows that it is very important to analyze the natural frequency of the prosthetic body and rotating parts, Modal analysis method to figure out the structure in a sensitive frequency range order modal characteristics can predict the structural vibration response of this band intrinsic external or internal variety of local oscillator role, The use of vibration during the same time in order to obtain the EMG intelligent prosthetics, it must be modal analysis, artificial limbs natural frequencies and mode shapes form. And it can be used to verify the vibration test data, and provide the appropriate basis for the results of the analysis, and to guide the selection of the acceleration sensor and determine the installation location. [3] EMG Intelligence prosthetic materials used in aluminum alloy, the material parameters: Young's modulus E=72GPa, density =2700kg/m3, Poisson's ratio = 0.3. The establishment of various components of the model in ANSYS using solid modeling method, to the natural frequency of the system by meshing and solving load, After the modal analysis in ANSYS artificial limb as a whole, as well as the six-order natural frequency of the rotating parts are as follows: TABLE I THE NATURAL FREQUENCY OF THE MOTOR Modal 1 2 3 4 5 6 Natural frequency 25.656 25.996 38.272 101.03 395.95 709.81 Fig. 4 Motor modes It can be seen from the above calculation, using the ANSYS modal analysis of mechanical structures, the method is simple, get the modal visual image, fully meet the modal analysis needs. And derive the corresponding graph of vibration and natural frequency, the natural frequency of the use of the rotating parts of the ANSYS analysis can guide the installation of the sensors and the analysis of the test signal from the data point of view, fourth-order natural frequency of the motor, especially close to its transfer frequency, it is easy to produce a resonance phenomenon, and to strengthen the vibration of the motor. And the natural frequency of the reduction gear is relatively low, it is easy to enlarge the vibration of the motor. Above the natural frequency of low frequency natural frequency of the basic 10000HZ the following, and then select ULT20 series piezoelectric acceleration sensor, the frequency range :0.7-10000HZ fully meet the test frequency range, taking into account the installation of the sensor, due to motor the natural frequency is very close to the transfer frequency. Therefore, the sensor is installed in the vicinity of the motor can get a strong signal of frequency, data analysis, and because of the failure of rotating parts in different directions have different characteristics of the phenomenon, so select the motor near the vertical two-way installation. .INTELLIGENT PROSTHETIC VIBRATION TEST PLATFORM TO BUILD" ] }, { "image_filename": "designv11_101_0001262_s00542-012-1580-3-Figure8-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001262_s00542-012-1580-3-Figure8-1.png", "caption": "Fig. 8 Stress wave propagation from bottom-left corner. a 1 ls, b 5 ls, c 10 ls, d 15 ls, e 20 ls, f 25 ls, g 30 ls", "texts": [], "surrounding_texts": [ "The measurement specimen is shown in Fig. 2. Seventeen shock sensors were attached to the top cover of an HDD. The HDD was put inside the thermostatic chamber shown in Fig. 3. The output of the shock sensors and the position error signal were measured with an oscilloscope. Data collection was triggered when the position error exceeded 20 nm. The chamber operated on a temperature cycle of 0\u201370 C every 180 min. A total of 31 cycles were run. We detected 492 shocks caused by the thermal stick\u2013slip phenomenon. The time histories of the responses of the 17 shock sensors for one of the shocks are shown in Fig. 4. The dominant oscillation was caused by the shock sensors resonating at around 90 kHz. The first responding shock sensor was S8, which was located at the top-left corner. Therefore, the stick\u2013slip hypocenter was assumed to be at the top-left screw point. The responding sensor order could not be immediately identified. we used response envelopes to observe the stress wave propagation. The envelope A(t) of the shock sensor response x(t) was calculated as follows A\u00f0t\u00de \u00bc jx\u00f0t\u00de \u00fe j~x\u00f0t\u00dej; ~x\u00f0t\u00de \u00bc Z1 1 x\u00f0u\u00de p\u00f0t u\u00de du \u00f01\u00de where ~x is the Hilbert transform of x, and j is an imaginary unit. The envelopes were normalized by the maximum value and were represented graphically with contour plots. The values between sensors were interpolated. The contour plots of the normalized envelopes changing with time are shown in Fig. 5. These figures show the stress wave propagation from the top-left corner. The time scale starts at the occurrence of the thermal stick\u2013slip phenomenon. From these figures, we found that the stress wave travels from the top-left corner to the bottom-right corner in 30 ls. This time was also obtained when the thermal stick\u2013slip phenomenon occurred at the other hypocenters. The stress wave propagations for those other hypocenters are shown in Figs. 6, 7, 8. From these figures, we conclude that the stress wave propagation can be measured by this method. The velocities of the stress wave were calculated from the distance between the hypocenter of the thermal stick\u2013slip phenomenon and shock sensors divided by the propagation time for each shock sensor. The velocity distribution is shown in Fig. 9. We found two peaks centered at 4.66 and 2.97 km/s. These velocities correspond to those of longitudinal and distortional waves in a plate. The theoretical velocities cL and c2 for stainless steel (the top cover of the HDD) are described as follows (Kolsky et al. 1963) cL \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi E q\u00f01 m2\u00de s \u00bc 5:31 km/s; c2 \u00bc ffiffiffiffi G q s \u00bc 3:14 km/s \u00f02\u00de where E is Young\u2019s modulus, G is shear modulus, m is Poisson\u2019s ratio, and q is density. The two measured velocities were deviated by 12 and 5 %, respectively, from the theoretical values; the error in the longitudinal wave was especially large. One reason for the error in the longitudinal wave could be the dependency of the velocity on the wavelength: the smaller the wavelength, the slower the velocity. This property of the velocity of the longitudinal wave shifted the measured mean value down. The position error caused by the stick\u2013slip phenomenon that occurred at the top-left corner at time 0 is shown in Fig. 10. The 176 measured position errors are superposed in Fig. 10a. The amplitudes were different but the phases 1 2 3 4 5 6 0 20 40 60 80 100 120 Velocity of stress wave [km/s] Fr eq ue nc y Fig. 9 Velocity distribution of stress wave 0 20 40 60 80 100 120 140 \u2212 40 \u2212 30 \u2212 20 \u2212 10 0 10 20 30 40 Time [\u00b5s] Po si tio n er ro r [n m ] (a) coincided. This suggests similar shocks occur due to the thermal stick\u2013slip phenomenon. A smooth response curve was obtained by averaging the results, as shown in Fig. 10b. There was dead time of around 40 ls. This is considered as the time in which the stress wave was traveling. The limit of position error to avoid overwriting on the adjacent tracks is 20 nm, therefore the first local extremum at 93 ls had the potential to overwrite. By defining the arrival time of the stress wave as the time at which the position error exceeded the extremum by 10 %, we found that the arrival time of the stress wave was 55 ls." ] }, { "image_filename": "designv11_101_0002256_978-1-4471-2330-9_10-Figure10.1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002256_978-1-4471-2330-9_10-Figure10.1-1.png", "caption": "Fig. 10.1 A sketch for the crank-slider mechanism", "texts": [ " The position of the ball is controlled on the platform by means of two independent actuators. The problem will include developing the equations of motion of the ball, developing the two PID controllers for the platform, and animating the ball and the platform. The electronic version of the Simulink \u00ae models, VRML models, in addition to the recorded movies for the animated crank-slider mechanism and the controlled motion of the ball on the plate can be downloaded from Springer\u2019s web site http:// extras.springer.com/ . Figure 10.1 shows a sketch for the crank-slider mechanism. Point O is the center of the crank shaft. ON is the crank shaft. Its length is = 10 (cm)r . NP is the connecting rod. Its length is = 20 (cm)l . Since P is constrained to move in the vertical direction Chapter 10 Animation of Crank-Slider Mechanism of a Piston Using Simulink \u00ae Matlab\u00ae is a registered trademark of The Mathworks, Inc. 124 10 Animation of Crank-Slider Mechanism of a Piston Using Simulink\u00ae and the rest of the components are coplanar, this system has one degree of freedom which will be considered to be the crankshaft angle b " ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000027_ijmr.2012.050101-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000027_ijmr.2012.050101-Figure1-1.png", "caption": "Figure 1 Tooth geometry (a) involute (b) trochoidal (c) involute-trochoidal in common coordinate system", "texts": [ " In the rest of this paper, these issues are dealt with either explicitely or implicitly in the case of spur gears based on appropriate discretisation of their generating involute and trochoidal curves. In particular, in Section 2 tooth geometry is defined, in Section 3 possible machining strategies are discussed, in Section 4 the parametric tool path is derived both for roughing and for finishing, in Section 5 machining of a sample gear is presented and in Section 6 conclusions are summarised and future work is outlined. Gear geometry is analysed in detail in Litvin and Fuentes (2004). The main characteristics of a single spur gear tooth can be seen in Figure 1. The largest part of the tooth profile consists of an involute curve starting on the base circle with radius rg, see Figure 1(a). The minor part of the tooth profile consists of the trochoidal curve, see Figure 1(b). The module (m) taking standard values and the number of teeth (n), which is typically a designer\u2019s choice, define the pitch circle, i.e., ro = m\u00b7n/2. The circular pitchis t0 = 2\u03c0ro/n. The parametrical equations of involute curve, as a function of the rolling angle \u03c9, are: ( ) [sin( ) ( ) cos( )]gx \u03c9 r \u03c9 \u03c9 \u03c9= \u2212 (1) ( ) [cos( ) ( )sin( )]gy \u03c9 r \u03c9 \u03c9 \u03c9= + (2) Base circle radius is calculated as rg = ro cos(a0). The involute angle \u03b10is characteristic of the gear-family and typically takes values: 20\u00ba, 25\u00ba and 14", " The same iteration principle as in involute point computation was used to ensure equidistance of consecutive points. A gear tooth space profile and the addendum circle arc corresponding to a single tooth fully define the pattern needed. Moreover, tooth space profile is symmetrical reducing by half the computation through mirroring. An important issue is to move both involute and trochoidal points to the same coordinate system X0OY0, whose vertical axis OY0 coincides with the symmetry axis of the gear tooth space, see Figure 1(c). Involute curve which is generated in \u03a7\u039f\u03a5 coordinate system should be rotated by an angle \u03c6inv\u2013 = \u03c6t/2\u2013 \u03b1g, whereas trochoidal curve, which is generated in XTO\u03a5T coordinate system, should be rotated by an angle v = \u03c6t/2\u2013 w, see Figure 4, through the use of rotation matrices. The angle between two sequential teeth is given as: \u03c6t = 2\u03c0/n and the angle between the tooth symmetry-axis and trochoidal O\u03a5Taxis is given as: 1 02 ,tw \u03c6 d r= \u2212 where ( ) ( )1 0 0 02 tan cos .od l B r= \u2212 \u2212\u03b1 \u03b1 \u0391ngle \u03b1g between the involute curve starting point on base circle and the tooth symmetry axis is calculated as: ( )0 0 0 02 tan " ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003706_1.5062214-Figure9-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003706_1.5062214-Figure9-1.png", "caption": "Fig. 9. Schematic diagram showing formation of porosity with different laser power", "texts": [ " Because the bubbles can float out of the surface of weld pool before it solidified, large solidification duration benefit the floating of bubbles. As shown in Fig. 8, when the welding speed is low, during the same time, more part of the molten pool is existed. More bubbles floats out of the molten pool and less pores remains in the weld bead. In other word, with the decrease in welding speed, the solidification duration of molten pool increase and this make the time for bubbles floating out of the weld pool increase. As shown in Fig. 9, the penetration depth decrease with the decrease in laser power (H0 b KC (\u222b fC dt + a ) for \u222b fC dt < \u2212a 0, otherwise (3.17) Similarly constitutive relation for RC is given as e = \u23a7\u23a8 \u23a9 RC fC for \u222b fC dt > b or \u222b fC dt < \u2212a 0, otherwise (3.18) Further discussion on modeling of backlash will be presented in subsequent chapters" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000264_2012-36-0457-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000264_2012-36-0457-Figure1-1.png", "caption": "Figure 1 \u2013 Charging Air Tube Support.", "texts": [ " The diesel engine charging air tube support has a important structural function and it failure affects parts that precludes the engine operation. The diesel engine charging air tube support, produced from hot rolled steel, has the important structural function of supporting the charging air tube, fixed on the turbine and the intercooler. The tube or turbine failure precludes the engine operation. The support also has the function of reducing the displacements imposed on the turbine by engine vibrations that may damage the structure of the mentioned assembly. The Figure 1 illustrates the studied support. The Figure 2 presents the support position in the engine. The material used to manufacture the support should follow the DIN EN 10149-3 1.0971 [1] (S260NC) standard chemical composition, which was confirmed by testing. The Table 1 shows the chemical composition obtained. The material mechanical properties are shown in Tab le 2. E is the elasticity modulus, is the Poisson Ratio, \u03c1 is the density, Rup is the rupture limit , Yield is the yield limit and Fat is the fatigue limit" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000002_amm.52-54.1560-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000002_amm.52-54.1560-Figure3-1.png", "caption": "Fig. 3 Specimen and layout of the measuring points", "texts": [ " The specimens used in the experiment were divided into four kinds, as is shown in table 1. The material of the specimens were T10 and GCr15; the surface manufacturing conditions of the specimens were milling and grinding and each kind contains two specimens. Because the temperature difference can\u2019t be measured directly, we surveyed the specific measuring points in the experiment and obtained the temperature difference by calculating. The specimen and layout of the measuring points are shown in figure 3. The Impact of Material on Thermal Contact Resistance. The curves of thermal contact resistance changing according to materials are shown in figure 4. From curves A and B, we can see that under the same forces and surface manufacturing conditions, the thermal contact resistances are different because of different materials. From curves B and C, we can see that the thermal contact resistances are different because of different surface manufacturing conditions even though the materials are the same" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001050_coase.2012.6386411-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001050_coase.2012.6386411-Figure1-1.png", "caption": "Fig. 1. 2-points defect.", "texts": [ " \u210e\ud835\udc3f\ud835\udc4d \ud835\udc61 is the component which is responsible for the variation in the impulse amplitudes and is assumed to be the hamming window [9]. As the impact of the load zone repeats after every revolution, \ud835\udc47\ud835\udc46 represents the time per revolution. The expression in (4) presents the vibrations induced by the single point defect. \ud835\udc63\ud835\udc56 ,1 \ud835\udc61 = \u210e\ud835\udc40 \ud835\udc61 \u2217 { \ud835\udeff \ud835\udc61 \u2212 \ud835\udc58\ud835\udc47\ud835\udc3c\ud835\udc45 \u00d7 ( \u210e\ud835\udc3f\ud835\udc4d \ud835\udc61 \u2217\u221e \ud835\udc59= \u2212\u221e \u221e \ud835\udc58= \u2212\u221e \ud835\udeff \ud835\udc61 \u2212 \ud835\udc59\ud835\udc47\ud835\udc46 )}. (4) Before we develop analytical expressions for multiple points defect model (MPDM), the intuitive understanding of the problem is provided. Figure 1 presents a case with two defect positions marked as \u2018a\u2019 and \u2018b\u2019. Assuming that the inner-race is rotating in the \u2018anticlockwise\u2019 direction, the rolling element \u2018A\u2019 first passes the defect position \u2018a\u2019 and then \u2018b\u2019. The vibrations produced by defect point \u2018b\u2019 are delayed version of the vibrations produced by defect point \u2018a\u2019 due to relative defect positions. In case of two point faults, the overall vibrations induced by two points fault can be represented as in (5): \ud835\udc63\ud835\udc56 ,2 \ud835\udc61 = \u210e\ud835\udc40 \ud835\udc61 \u2217 \ud835\udeff \ud835\udc61 \u2212 \ud835\udc58\ud835\udc47\ud835\udc3c\ud835\udc45\u2212 \ud835\udc61\ud835\udc511 \u00d7\u221e \ud835\udc58= \u2212\u221e \u210e\ud835\udc3f\ud835\udc4d \ud835\udc61 \u2217 \ud835\udeff \ud835\udc61 \u2212 \ud835\udc59\ud835\udc47\ud835\udc3c\ud835\udc46\u2212 \ud835\udc61\ud835\udc511 \u221e \ud835\udc59= \u2212\u221e + \ud835\udeff \ud835\udc61 \u2212 \ud835\udc58\ud835\udc47\ud835\udc3c\ud835\udc45 \u2212 \ud835\udc61\ud835\udc512 \u00d7\u221e \ud835\udc58= \u2212\u221e \u210e\ud835\udc3f\ud835\udc4d \ud835\udc61 \u2217 \ud835\udeff \ud835\udc61 \u2212 \ud835\udc59\ud835\udc47\ud835\udc3c\ud835\udc46\u2212 \ud835\udc61\ud835\udc512 \u221e \ud835\udc59= \u2212\u221e . (5) In (5), \ud835\udc61\ud835\udc511 and \ud835\udc61\ud835\udc512 refer to the instantaneous positions of the first and the second fault respectively, i.e., incorporate the time shift (delay) between the impulses generated by two different faults as shown in Fig. 1. Equation (5) can be rearranged as in (6): \ud835\udc63\ud835\udc56 ,2 \ud835\udc61 = \u210e\ud835\udc40 \ud835\udc61 \u2217 \ud835\udc3c\ud835\udc56 \ud835\udc61\u2212 \ud835\udc61\ud835\udc511 + \ud835\udc3c\ud835\udc56(\ud835\udc61 \u2212 \ud835\udc61\ud835\udc512) , (6) where \ud835\udc3c\ud835\udc56(\ud835\udc61) represents the impulse train generated by single point inner race defect. In (6), the two impulse trains are generated with instantaneous positions as \ud835\udc61\ud835\udc511 and \ud835\udc61\ud835\udc512. Let us assume that the relative positions of the defect points with respect to predefined reference are \ud835\udc61\ud835\udc511, \ud835\udc61\ud835\udc512, \ud835\udc61\ud835\udc513, \u2026 , \ud835\udc61\ud835\udc51\ud835\udc5b , then an n-points defect can be generalized as in (7): \ud835\udc63\ud835\udc56 ,\ud835\udc5b \ud835\udc61 = \u210e\ud835\udc40 \ud835\udc61 \u2217 \ud835\udc3c\ud835\udc56(\ud835\udc61 \u2212 \ud835\udc61\ud835\udc51\ud835\udc57) \ud835\udc5b \ud835\udc57=1 , (7) Equation (7) shows that the n-points defect is the summation of the impulse trains generated by each fault point with a particular reference position" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000264_2012-36-0457-Figure9-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000264_2012-36-0457-Figure9-1.png", "caption": "Figure 9: Static analysis results - Pressure at the Tube Socket.", "texts": [ "00 [Hz], which is the 4 \u00bd order frequency of the 6 cylinder engine, for a maximum of 2200 [rpm] with a 10% safety factor, according to Campbell Diagram [5]. In the pressure at the tube socket and thermoelastic static analyses and in the dynamic analysis, the stresses in the support must not exceed the material fat igue limit [6]. The engine assembly presented their first natural frequencies at 157.21 [Hz]. Figure 8 shows the modal analysis results. 1st Mode f1 = 157.21 [Hz] Maximum displacement versor x = -0.62; y = 0.49; z = 0,61 The pressure at the air tube socket and thermoelastic static analyses and dynamic analysis results are shown in Figure 9, 10 and 11 respectively. Values within a red frame are above material fat igue limit. Tab le 4 shows the static and dynamic analyses summary results. The von Mises stresses are presented as a percentage of the materials fatigue limits. The values presented in red are above the material fatigue limit. Table 4 - Static and dynamic analyses summary results . Analyses Support fat = 127 [N/mm\u00b2] Pressure 143% Thermoeslastic 55% Dynamic 98% Figure 10: Static analysis results \u2013 Thermoelastic. Figure 11: Dynamic analysis results" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000326_amr.472-475.2096-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000326_amr.472-475.2096-Figure1-1.png", "caption": "Fig. 1 Stewart Mechanism Fig. 2 Structure diagram", "texts": [ " Based on the spatial mechanism parallel characteristics, this paper presents an ideal for its kinetics modeling in parallel, and takes a typical mechanism kinetics analysis modeling as an example, analyzes its parallel modeling process. The rest of the paper is organized as follows: section 2 introduces the parallel analysis for parallel robot mechanism kinetics modeling; Section 3 introduces the parallel modeling process; Section 4 takes a typical spatial mechanism kinetics analysis modeling as an parallel modeling example ;Section 5 presents the conclusion. Taking the typical Stewart mechanism (Fig. 1) as an example, static platform \uff08 1 2 3 4 5 6B B B B B B \uff09and movable platform\uff08 1 2 3 4 5 6 A A A A A A\uff09uses the circular layout, the radius are R and r respectively, and the static platform and movable platform are connected through 6 driving legs \uff081, 2, 3, 4, 5, 6\uff09, which length are expressed as \uff08 1L , L2 , 3L , 4L , 5L , 6L \uff09. 1A , 2 A , 3 A , 4A , 5 A , 6A , 1 B , 2B , 3 B , 4B , 5 B , 6 B are spherical pairs, 1P , 2P , 3P , 4P , 5P , 6P are moving pairs and active pairs through which change movable platform\u2019s position and orientation", " The Stewart mechanism has 6 same branches which are SPS structured, according to the ordered SOC theory, the Stewart mechanism can be decomposed into 5 ordered SOC. All rights reserved. No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of Trans Tech Publications, www.ttp.net. (ID: 130.88.90.140, The University of Manchester, Manchester, United Kingdom-29/04/15,14:48:03) The structure decomposition of the Stewart mechanism (Fig. 1) can be as following: { }1 1 [6,5,3] 1 1 1 4 4 4 ( 3)KC SOC B P A A P B= \u2212 \u2212 \u2212 \u2212 \u2212 \u2212 \u2212 \u2206 = + { }2 22 2 2 ( 0)PSOC B P A O\u2295 \u2212 \u2212 \u2212 \u2212 \u2212 \u2206 = { }3 3 3 3 3 ( 1)SOC B P A\u2295 \u2212 \u2212 \u2212 \u2212 \u2206 = \u2212 { }4 4 5 5 5 ( 1)SOC B P A\u2295 \u2212 \u2212 \u2212 \u2212 \u2206 = \u2212 { }5 56 6 6 ( 1)SOC B P A\u2295 \u2212 \u2212 \u2212 \u2212 \u2206 = \u2212 The first SOC ( { }1 1 1 1 4 4 4SOC B P A A P B\u2212 \u2212 \u2212 \u2212 \u2212 \u2212 \u2212 ) is composed by L1 and L4 with 3 local degrees of freedom, the structure factor of this branch is 14-2-6-3=3\uff1b The 2nd SOC ( { }2 2 2 2 P SOC B P A O\u2212 \u2212 \u2212 \u2212 \u2212 ) consists L2, the structure factor of this SOC is 0" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000075_amr.457-458.237-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000075_amr.457-458.237-Figure1-1.png", "caption": "Fig. 1 Stratospheric airship sketch", "texts": [ " On the basis of previous work, this article emphasizes the differences between airship and other aircrafts, seriously deduces the dynamic model using Newton method, and makes a simulation investigation for the corresponding aspects. Brief Introduction of the Stratospheric Airship The structural design and dimension of the airship make a difference according diverse mission demands, so its forces and motion characteristics are different. In this article, the airship is a non-rigid airship. Its general layout is as shown in Fig. 1. The airship is 250 meters long, and is made up of ballonet and gasbag. The gasbag has four gas chambers which are filled with helium and the ballonet has three gas chambers which are filled with air. The rear of the airship is equipped with propeller to supply thrust force. The bottom of the airship is a cabin. Remote sensing system, telemetering system, control system, the payload and other equipments are fixed in the cabin. Ducted fans are mounted on the two sides of rear end of the cabin, which can carry out vector thrust" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000031_amm.162.11-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000031_amm.162.11-Figure2-1.png", "caption": "Figure 2 Belt mechanism with circular eccentric output link", "texts": [ " The condition of pure rolling of line l on circle of radius r is: ( )\u03b1\u03b3 +\u22c5= rAT . (1) All rights reserved. No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of Trans Tech Publications, www.ttp.net. (ID: 136.186.1.81, Swinburne University, Hawthorn, Australia-16/05/15,18:30:11) The equation of the involute (i) in polar parametric co-ordinates (Fig. 1) is: =\u2212=\u2212= = .tan , cos \u03b1\u03b1\u03b1\u03b1\u03b3 \u03b1 inv r AT r R (2) The involute\u2019s parameter is the angle \u03b1 In Fig. 2 the considered mechanism is presented with its geometrical characteristic parameters. The kinematic analysis consists in determining the transmission function ( )\u03d5\u03c8\u03c8 = and its derivative i.e. the velocities and accelerations functions respectively (\u03c8 and \u03c8 ). The starting position is considered to be an extreme one, usually 0min \u03c8\u03c8\u03c8 == (Fig.3) and is supposed to be known, because it represents the mechanism\u2019s initialization position [10]. The considered belt mechanism has two mechanisms that are instantaneous isokinetic with it [7, 8, 9]: a \u201cfour bar\u201d mechanism OABD (with 2 variable links length AB and BD, Fig. 2) and a fictive/ imaginary \u201ccam mechanism\u201d with involute of point A as \u201ccam\u201d and link 2 (OA) as follower, respectively (Fig. 3). The involute is univocally defined knowing its initial point K on the circle of radius 4r . The condition of pure rolling of the flexible element (the line l ) on the circle can be written in the form: ( )00400 )( \u03b1\u03b1 invrKBarcl +\u22c5== . (3) The initial position of the circle\u2019s centre C in respect with the origin O can be expressed in the triangle OCD and 0OCB as being: ( ) 2 1 041 2 4 2 1 cos2 \u03c8\u22c5++= llllOC (3) and ( )( ) 2 1 2 4 2 02 rllOC ++= , (3\u2019) respectively, from which results the initial free length of the flexible-unextensible element: ( ) 2 2 1 2 4041 2 4 2 10 cos2 lrlllll \u2212\u2212\u22c5++= \u03c8 " ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002030_ijhvs.2013.053008-Figure16-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002030_ijhvs.2013.053008-Figure16-1.png", "caption": "Figure 16 Shape of the natural vibrating mode at 20.86 Hz (II mode)", "texts": [], "surrounding_texts": [ "The performed experimental analysis suggested the possibility of reproducing the experimentally observed behaviour using a simple lumped parameters model of the driveline. In particular a linear torsional model of the driveline with 17 degrees of freedom (dofs) has been set up accounting for the torsional deformability of the transmission shafts and of the wheelset axles (see Figure 6). The dofs of the model are the rotation of each wheel, of the cardan shafts extremities and of the gearboxes. Moments of inertia of shafts and gearboxes have been evaluated through 3D solid models, while torsional stiffness and damping have been identified by minimising the difference between the oscillation frequency obtained from the experimental data and from the model. In order to evaluate the effect of creep force-creep curve shape on torsional vibrations in the locomotive driveline, the three different curves shown in Figure 7 have been considered. Curve 1 (dashed line) represents the SHE model with a constant friction coefficient, while Curves 2 and 3 refer to the model proposed by Polach. As already mentioned (Section 2), a friction coefficient dependent on the creep velocity and implicitly from the vehicle speed (see equations (1) and (2)) is assumed in this model. In particular, Curve 2 (solid line) is concerned with a constant vehicle speed of 3 m/s, while Curve 3 (dashed-dotted line) refers to a constant vehicle speed equal to 25 m/s. It is worth remembering that the main effect of contact forces in longitudinal motion is to introduce damping into the system (Wickens, 2003). The equilibrium at rotation of the i-th wheel is in fact given by: J y ,i\u2126i + Fx ,i Ri = 0 (3) where Jy,i represents the moment of inertia of the i-th wheel, \u2126i is its angular speed, Fx,i and Ri are the longitudinal contact force and the rolling radius, respectively. As known, the longitudinal force Fx acting on a wheel is a function of the vertical load (N) and of the longitudinal (\u03b5x) and lateral (\u03b5y) creepage, its linearisation leads to: Fx = Fx0 + \u2202Fx \u2202\u03b5x 0 \u03b5x \u2212\u03b5x0( ) + \u2202Fx \u2202\u03b5 y 0 \u03b5 y \u2212\u03b5 y0( ) + \u2202Fx \u2202N 0 N \u2212 N0( ) (4) Assuming small oscillations in the neighbourhood of tangent track condition (small lateral creep) and neglecting the load transfers, the linearised expression of the longitudinal contact force can be reduced to: Fx = Fx0 + \u2202Fx \u2202\u03b5x 0 \u03b5x \u2212\u03b5x0( ) (5) Considering that the longitudinal creepage can be expressed as 1x R V \u03b5 \u2126 = \u2212 (6) where V is the longitudinal speed of the vehicle and the equivalent damping introduced by longitudinal contact force of each wheel is given by: req = \u2202Fx \u2202\u03b5x 0 R2 V = f0 R2 V (7) Since f0 represents the slope of the creep force-creep curve calculated in correspondence of the longitudinal creep \u03b5x0, the equivalent damping introduced by the contact forces can become equal to zero (SHE model) or negative (Polach model) for high creepages (see Figure 7), this may lead to unstable vibrations. From Figures 9\u201314 the effect of a different shape of the creep force-creep curve at high creepages is investigated. The damping factor (h) and the natural frequency \u03c9 of the first four vibrating modes are represented. The shape of the vibrating modes is instead reported in Figures 15\u201318. The root loci have been obtained by linearising the contact forces of all the wheels in correspondence of the same longitudinal creep. The longitudinal creep range 0.009\u22120.3 has been investigated. In particular, in all the figures, crosses are referred to a longitudinal creep of 0.009 and circles are associated to a longitudinal creep equal to 0.3. A structural damping factor equal to 1\u2030 has been considered. Figures 9 and 10 refer to the SHE model (Curve 1 in Figure 7), Figures 11 and 12 are concerned with the model proposed by Polach considering a constant vehicle speed of 3 m/s (Curve 2 in Figure 7), while Figures 13 and 14 are associated to the Polach model when a constant vehicle speed of 25 m/s is assumed (Curve 3 in Figure 7). As it can be seen, all the contact models lead to a non-oscillatory response of the first two vibrating modes at low creepages. This avoids torsional oscillations during standard operating conditions. As the longitudinal creepage increases the damping factor (h) of all the vibrating modes decreases and this produces oscillations in the driveline response. A significant effect of creep force-creep curve slope at high creepages can also be observed. If the SHE contact model is considered (Curve 1 in Figure 7), the real part of the vibrating modes tends to be zero, but the system remains always stable (see Figure 10). The slope of the creep force-creep curve at high creepage predicted by the SHE model in fact is equal to zero and so as the equivalent damping introduced by the contact forces req. If a negative slope is instead considered, some of the vibrating modes may present a positive real part \u03b1 (unstable torsional vibrations). If the contact model represented by Curve 2 of Figure 7 is considered, the branches of the root locus have a monotone trend (see Figure 12), since the creep force-creep curve, once passed its maximum, presents an almost constant negative slope with increasing creepage. On the other hand, Curve 3 (Figure 7), shows an inflection point for a creepage value of 0.2. Thus the equivalent damping introduced by the contact forces is negative (unstable torsional vibrations) in the range of creepages 0.04\u22120.2, and is positive for creepages higher than 0.2.3 In Table 2 the creepage values at which the real part of the vibrating modes become positive is reported for the first four vibrating modes taking into account the three contact models. The performed analysis thus confirms the possibility of unstable torsional oscillations in the locomotive powertrain only in presence of a decreasing section of the creep force-creep curve at high creepages. To verify the capability of the implemented lumped parameters torsional model in reproducing the experimentally observed behavior of the vehicle powertrain, the numerically obtained vibrating modes have been compared with the experimental data. Figure 19 depicts the shape of the vibrating mode associated with the first calculated natural frequency of the driveline. The contact model represented by Curve 2 of Figure 7 is considered and creep force-creep curve is linearised in the neighbourhood of the condition highlighted in Figure 8 (longitudinal slip \u03b5x = 0.025) for all the wheels. The amplitude of the torsional vibrations is expressed in percentage with respect to the maximum value. The damped frequency of the first vibrating mode is 15.93 Hz (while its natural frequency is 18.07 Hz). It can be observed that the obtained numerical vibrating mode is in good agreement with the experimental one shown in Figure 5, especially for the front cardan shafts. Differences for the rear cardan shafts are mainly due to the simplifications introduced into the model, i.e. the same longitudinal creepage has been imposed to all the wheels and load transfers have been neglected." ] }, { "image_filename": "designv11_101_0000752_amr.860-863.2678-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000752_amr.860-863.2678-Figure2-1.png", "caption": "Fig. 2 Mold damaged area feature processing and modeling", "texts": [ " Under such a set of parameters for single channel cladding width 2 mm, height of 0.6 mm, three-channel (lap rate 30%). 5 mm, 0.7 mm high, and the quality of the cladding is better, so choose the lap rate 30%, under the above conditions cladding. \uff083\uff09Secondary machining and clean up for repair area In order to make the cladding layer and substrate in combination with solid, cladding before the need for secondary processing, damaged area degusting degreasing and features to simplify, as shown in figure 2 (a) : characteristics of mold damage area with mechanical grinding degusting and processing, chemical processing and using acetone. \uff084\uff09Cladding area measurement and 3 d modeling After the secondary processing machinery, the use of geometric measurement tools pre cladding zone (as shown in Fig. 2 (a)) measurement, and according to the measured data by 3 d software build mold damaged part as shown in Fig. 2 (b) the three dimensional CAD model. \uff085\uff09Cladding trajectory planning and offline programming \u2460cladding trajectory planning Coaxial laser cladding found by experiment many times to send powder processing scanning track and speed had a great influence on the processing quality, especially at the end of the track initiation process and process of cladding layer due to slow scanning speed and produce nodular design trajectory will therefore uniform segment in the processing zone, the start and end period of fixture or substrate, and then USES the straight line parallel scanning trajectory, each path between two layers of perpendicular to each other, in order to improve the intensity of cladding layer" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001744_iros.2011.6095082-Figure4-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001744_iros.2011.6095082-Figure4-1.png", "caption": "Fig. 4 First concept of the crossed type crawler", "texts": [ " And they do not have complete water proof and dust proof abilities because the mechanisms of own robots are complex and the sealing is difficult to make. But SCV has these abilities by the high strength retractable skin. This chapter describes two concepts of SCV. The first concept is basic concept, and the second concept is for realize mechanisms. We explain the first concept of SCV. At first, the Flexible crawler belt is rolled in the columnar center pulley which is driving pulley and both ends of the flexible crawler belt are fixed on the edges of two end pulleys which are driving pulleys, as shown at the top of Fig. 4. Next, the flexible crawler belt is expanded from the inside by the end pulleys, and it is transformed into a diamond shape shown at the bottom of Fig. 4. When the center pulley starts to rotate, the flexible crawler belt is rotated with a diamond shape kept. The feature of this mechanism is that the inside of the flexible crawler belt is kept completely airtight. And there is no friction loss because all elements rotate without friction. With this mechanism it is possible to realize a mobile robot \u201cSVC\u201d which is a completely dust proof and water proof. The retractable rate of the flexible crawler belt is needed 300% to achieve this mechanism. When SCV try the steering motion, it needs to make the steering circular arc as the turn orbit" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001393_detc2013-12738-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001393_detc2013-12738-Figure2-1.png", "caption": "Figure 2. Outer race waviness", "texts": [ " A constant radial force is assumed to be acting on the system. A rolling element bearing with supporting a rigid horizontal rotor with an unbalanced force (Fu) is taken for theoretical simulation. To simplify the study, unbalanced rotor force is assumed to be constant in the entire wave range. The level of unbalanced rotor force has been considered as 15% of radial load (W). An important source of vibrations in ball bearings is waviness. These are imperfections with varying amplitude across the circumference of the outer race of the bearing as shown in Fig. 2. The Hertz equations for elastic deformation involving point contact between solid bodies are given by Br\u00e4ndlein et al. [18] as 3 2 2 2 3 11 25.1 Q E MKr i (1) (4.9) Here, \u03c0, K and \u03bc are Hertz coefficients which depend on the surface properties. E, 1/M and \u2211\u03c1 are the elastic modulus in N/mm 2 , Poisson ratio and the sum of curvature of the contacting bodies, respectively. The contact force (Q) is )N( 25.1 3 11 2 2 i i r K r M EQ (2) Hence the nonlinear stiffness associated with point contact is 2 2 25", " (5) is two coupled non-linear ordinary second order differential equations having parametric effect, the 1.5 non-linearity and the summation term. The \u2018+\u2019 sign as subscript in these equations signifies that if the expression inside the bracket is greater than zero, then the rolling element at angular location i is loaded giving rise to restoring force and if the expression inside bracket is negative or zero, then the rolling element is not in the load zone, and restoring force is set to zero. Consider the outer race as a circle of variable radius as shown in Fig. 2; the function describing can be expanded in a Fourier series. Taking the above into account, the profile of the raceway of the outer race can be represented by, ti N NR cage b w n n )1(2sin (6) where R is the constant component of the radii of the outer race, Nw is number of waves, Nb is number of balls, n is the order of harmonic of waviness in the outer race, and \u03a0n is the amplitude of these harmonics. Considering the waviness of outer race, an additional deformation in the contact deformation becomes ti N N cage b w n ni )1(2sin (7) The coupled non-linear second order differential equations (5) are solved by numerical integration technique to obtain the radial displacement and velocity of the rolling elements" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000160_2013-01-1757-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000160_2013-01-1757-Figure1-1.png", "caption": "Figure 1. Motor of Honda IMA system", "texts": [ "lectric automobiles are on the rise because they eliminate CO2 emissions. More electric cars will mean an increased demand for magnets used in hybrid car motors(1). Figure 1 shows a motor of Honda Integrated Motor Assist (IMA) system. The motor uses Nd2Fe14B sintered magnets with high magnetic flux density, contributing to increased torque and downsizing. When used at a high temperature, however, the magnetic flux density of these magnets lowers, which in turn reduces motor torque. This demagnetization characteristic is supplemented by Nd2Fe14B sintered magnets containing Dysprosium (Dy), to replace part of the Neodymium (Nd). High anisotropy fields of (Nd, Dy)2Fe14B make it possible to use magnets at high temperatures" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000429_compel.2013.6626446-Figure4-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000429_compel.2013.6626446-Figure4-1.png", "caption": "Fig. 4 (a) Overall block diagram (b) Prototype of proposed double three-phase full bridge SRM drive system", "texts": [ " Also, the minimum peak will be occurred at 90, 210 and 330 degree. For example, at \u03b8 = 270 deg, the maximum peak of total torque waveforms can be rewritten as (9) Similarly, at \u03b8 = 210 deg, the minimum peak of torque waveform can be rewritten as (10) Form (9) and (10), the average of torque waveform is 1.5L0IdcIpk while the amplitude of the torque ripple is 0.375L0I 2 pk Under the constant rms current value (I), Ipk can be expressed as 2 2 2*pk dc I I I = \u2212 (11) Average torque can be re-written as 2 2 0 1.5* * 2* *avg dc dcL I I I\u03c4 = \u2212 (12) Fig. 4 shows the overall block diagram of the drive system. An optical encoder is installed at the motor shaft. The rotor position and switching pattern is determined by A, B and Z signals after the Z-pulse (1 p/r) is provides by 16 bit micro-controller (dSPIC-30F4011). Two units of power module model FSBB30CH60 (Fairchild) are used. The motor current is controlled by current loop inside micro-controller. In the simulation, the current command composes of the ordinary rectangular current command and a superimposed DC offset command" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002053_fie.2011.6142725-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002053_fie.2011.6142725-Figure2-1.png", "caption": "FIGURE 2 ILLUSTRATION OF PRINT HEAD POSITIONING MECHANISM.", "texts": [ " During a recent search on IEEE Xplore \u00ae database, the author did not find any paper on the subject of using w-waste to the benefit of hands-on engineering education. While browsing recent conference proceedings, a paper was found in which the authors discuss how to design a didactic milling machine based on the reuse of a dismantled printer\u2019s part [7]. This section presents some hands-on experiments created at the Electrical Engineering Department of S\u00e3o Paulo State University, using the campus e-waste. The resulting experiments are now part of the microprocessor systems lab classes. Figure 2 shows a sketch of the utilized positioning mechanism withdrawn from a defunct ink jet printer. Beyond the parts illustrated in the sketch, the following items come ready for use (Figure 3): the print head (cartridge bearing support) and a plastic mechanical support that holds all parts in place. The print head is mounted on a stainless steel guide through a sliding bushing. The print head is tightly connected to the toothed belt which, in turn, is driven by a dc motor. 978-1-61284-469-5/11/$26.00 \u00a92011 IEEE October 12 - 15, 2011, Rapid City, SD 41 st ASEE/IEEE Frontiers in Education Conference T3G-3 As the motor runs, the toothed belt moves the print head" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001782_amr.681.219-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001782_amr.681.219-Figure1-1.png", "caption": "Fig. 1 Gear pair torsional dynamics model. Fig. 2 Calculation model of involute tooth deformation", "texts": [ " Therefore, the multi-body model was established in ADAMS, then Simulink received the angular velocities and angular displacements of gears from ADAMS to solve the meshing force which was fed back to ADAMS. The influence of speed fluctuation on time-varying stiffness, gear clearance, meshing errors were considered in the model. The establishment of non-linear multi-body gear pair model can lay the foundation for the complex gearbox modeling which contains flexible shafts and flexible box. Gear pair transmission model Analysis of Gear Meshing Dynamic Excitations. Cylinder gear joint can be simplified as torsional vibration system (Fig.1) without considering the flexibility of axis, rolling bearing and box [5] . If the contact ratio is 1~2, the torsional vibration analysis model of a gear pair can be deduced as: 1212111 )()( TeRRkReRRcRJ bamabama =\u2212\u2212+\u2212\u2212+ \u03b8\u03b8\u03b8\u03b8\u03b8 . (1) All rights reserved. No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of Trans Tech Publications, www.ttp.net. (ID: 130.216.30.116, University of Auckland, Auckland, New Zealand-14/05/15,00:20:08) 2121222 )()( TeRRkReRRcRJ abmbabmb \u2212=+\u2212++\u2212+ \u03b8\u03b8\u03b8\u03b8\u03b8 " ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001680_ccdc.2013.6561389-Figure8-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001680_ccdc.2013.6561389-Figure8-1.png", "caption": "Fig. 8: Collision interval calculation diagram 3) P(z, x) has no solution", "texts": [ "6(c)) on the line-shaped obstacle. But they don\u2019t meet the formula (1) and will be processed as non-collision between the mechanical arm and the line-shaped obstacle. (1), which can processed as collision between the mechanical arm and the line-shaped obstacle. (a) (b) (c) (d) (e) (f) Fig. 7: Collision interval calculation diagram 2) P(z, x) has the double solutions The case is that the circular arc passed by the straightened mechanical arm and the line-shaped obstacle have two cross points, as shown in Fig.8. The cross points may be the endpoints P1 and P2 (Fig.8(a)) or 2654 2013 25th Chinese Control and Decision Conference (CCDC) the points P\u2019 and P\u201d (Fig.8(d)) on the line-shaped obstacle. It is also possible that one is the endpoint P1 (Fig.8(c)) or P2 (Fig.8(b)) and another is the points P\u2019(Fig.8(b)) or P\u201d (Fig.8(c)) on the line-shaped obstacle. P1P2 (except for the cross points) in Fig.8(a), P\u2019P2 (except for the cross points) in Fig.8(b), P1P\u2019 (except for the cross points) in Fig.8(c) and P\u2019P\u201d (except for the cross points) in Fig.8(d) meet the formula (1), which are processed as collision between the mechanical arm and the line-shaped obstacle. The case is that the circular arc passed by the straightened mechanical arm and the line-shaped obstacle have not the cross point, as shown in Fig.9. They meet the formula (1) and can be processed as collision between the mechanical arm and the line-shaped obstacle. (a) 2013 25th Chinese Control and Decision Conference (CCDC) 2655 C-space By above knowable, the obstacles in C-space are expressed as the boundary function, based on which the obstacle avoidance path planning can\u2019t be done directly, thus the information of the boundary function need be described" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003522_1.4704335-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003522_1.4704335-Figure1-1.png", "caption": "FIGURE 1. Roller element bearing for physical modeling of simulation.", "texts": [ " In this research, vegetable oils which have been tested by Mia and Ohno [10], as shown in Table 1, will be used as lubricants in simulation of elastohydrodynamic lubrication, which the contact zone of this lubrication is too small so that the high pressure on the contacting elements will deform the surfaces, such as contact between mating-gears, roller element bearing, cams, etc. In this simulation, roller element bearing will be used as physical modeling which is a device used to support a rotating shaft to the bearing housing and reduce the friction between contacting surfaces, as shown in Figure 1. Because the limitation study in only two dimensional, a simple assumption for transverse length of the bearing is effectively infinite and pressure distribution is uniform in the y direction, meaning side leakage is neglected. Other assumption is simulation running in constant temperature (isothermal condition). Calculation of film thicknesses and pressures in line contact of elastohydrodynamic lubrication is based on the fast and accurate method, developed by Houpert and Hamrock [11], which used Newton-Raphson method in solving the elasticity and Reynolds equation simultaneously" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000051_sbr-lars.2012.40-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000051_sbr-lars.2012.40-Figure1-1.png", "caption": "Fig. 1. Kinematic model of a car-like robot [2].", "texts": [ " The variables x and y represent the vehicle\u2019s position in the plane of the axes XY , and \u03b8 is the orientation (heading) related to the X axis. The kinematic model of a Dubins vehicle is given by: q\u0307 = \u23a1 \u23a3 x\u0307 y\u0307 \u03b8\u0307 \u23a4 \u23a6 = \u23a1 \u23a3 \u03bd cos(\u03b8) \u03bd sin(\u03b8) \u03c9 \u23a4 \u23a6 , (1) where \u03bd (\u03bd \u2208 R +) is the linear velocity and \u03c9 the angular velocity (\u03c9 \u2208 {\u2212\u03bd/\u03c1, 0, \u03bd/\u03c1}), where \u03c1 is the minimum curvature radius of the curve that the vehicle is capable to execute. We focused the present work on Unmanned Ground Vehicles (UGV) with Ackerman steering (car-like) models (Figure 1), which represents most of the vehicles nowadays. This model introduces two new variables, L represents the distance between the front and rear axles and \u03c6 denotes the steering angle. The minimum curvature radius can be obtained by \u03c1 = L tan(\u03c6) , and consequently d\u03b8 dt = \u03bd L tan(u\u03c6), where u\u03c6 = {\u2212\u03c6max, 0, \u03c6max}. In the present approach, we consider \u03bd, L and \u03c6max are constant and a priori known. Our proposal is divided in two steps: an heuristic to approximate the cost to goal function and the algorithm to find the path to goal based on A* search" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001402_s13369-012-0232-3-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001402_s13369-012-0232-3-Figure3-1.png", "caption": "Fig. 3 The determination of the pinion with concave teeth", "texts": [], "surrounding_texts": [ "Figures 3 and 4 define the coordinate systems for obtaining the generating surface. Two fixed reference coordinate systems Sn(On, xn, yn, zn) and Sr(Or, xr, yr, zr) are rigidly connected to the frames of the concave gear and convex gear, respectively. A fixed coordinate system Sf(Of , xf , yf , zf) is rigidly connected to the frame. Coordinate system Sc(Oc, xc, yc, zc) is attached to the conical surface. The surface translates over the linear displacement S along the yf -axis. The coordinate system S2(O2, x2, y2, z2) rotates at angle \u03c62, and S1(O1, x1, y1, z1) rotates at \u03c61 and about the xn-axis and xr-axis, respectively. The angle \u03b1 is a rotary angle. Its region is 0 \u223c 2\u03c0 . To generate an envelope to the family of the ring surfaces, a coordinate transformation matrix from the Sc coordinate system to the S1 coordinate system can also be used. By applying homogeneous coordinates matrix [7] for coordinate transformation, the matrix M1c, can be obtained as follows: M1c(\u03b1, \u03c61) = \u23a1 \u23a2\u23a3 cos \u03b1 sin \u03b1 cos \u03c61 sin \u03b1 sin \u03c61 \u2212rp1\u03c61 sin \u03b1 cos \u03c61 + rp1 sin \u03b1 sin \u03c61 \u2212 sin \u03b1 cos \u03b1 cos \u03c61 cos \u03b1 sin \u03c61 \u2212rp1 cos \u03b1 cos \u03c61 + rp1 cos \u03b1 sin \u03c61 0 \u2212 sin \u03c61 cos \u03c61 rp1 sin \u03c61 + rp1 cos \u03c61 0 0 0 1 \u23a4 \u23a5\u23a6 (4) where rp1 represents the standard pitch radii of the convex gear. Similarly, matrix M2c is the coordinate transformation matrix from Sc to S2. Thus, matrix M2c can be written as follows: M2c(\u03b1, \u03c62) = \u23a1 \u23a2\u23a3 cos \u03b1 sin \u03b1 cos \u03c62 \u2212 sin \u03b1 sin \u03c62 \u2212rp2\u03c62 sin \u03b1 cos \u03c62 + rp2 sin \u03b1 sin \u03c62 \u2212 sin \u03b1 cos \u03b1 cos \u03c62 \u2212 cos \u03b1 sin \u03c62 \u2212rp2\u03c62 cos \u03b1 cos \u03c62 + rp2 cos \u03b1 sin \u03c62 0 sin \u03c62 cos \u03c62 \u2212rp2\u03c62 sin \u03c62 \u2212 rp2 cos \u03c62 0 0 0 1 \u23a4 \u23a5\u23a6 (5) where rp2 represents the standard pitch radii of the concave gear. Using Equations (1)\u2013(5) and the coordinate transformation matrix Mic(\u03b1, \u03c6i ), the family of the ring surfaces can be expressed by: Rg i (\u03b2, j , \u03b1, \u03c6i ) = Mic(\u03b1, \u03c6i )R g c(\u03b2, j ) (6) where \u03b1 and \u03c6i are independent parameters calling the parameters of motion. The variable \u03b2 and j are the design parameters as represented in Sect. 2 ( j = c, d, h). Vector Rg i (\u03b2, j , \u03b1, \u03c6i ) is the family of the ring surfaces. Superscript g is ab, bc, and cd . Subscript i is 1 and 2. The position vector of the ring surface is Rg c, as indicated in Equations (1)\u2013(3). If the family of ring surfaces is regular surface, one can find an envelope to a two-parameter family of surfaces [7]. The envelope to a two-parameter family of the ring surfaces should satisfy the equations: ( \u2202Rg i \u2202 j \u00d7 \u2202Rg i \u2202\u03b2 ) \u00b7 \u2202Rg i \u2202\u03c6i = 0 (7) ( \u2202Rg i \u2202 j \u00d7 \u2202Rg i \u2202\u03b2 ) \u00b7 \u2202Rg i \u2202\u03b1 = 0 (8) Substituting Equations (1), (2) and (6) into Equations (7) and (8) with i = 1, an envelope 1 to the family of the ring surfaces is obtained. This envelope 1 is called the convex tooth of the gear. Substituting Equations (2), (3) and (6) into Equations (7) and (8) with i = 2, an envelope 2 to the family of the ring surfaces is obtained. The envelope 2 is called concave the tooth of the pinion. Substituting Equations (1)\u2013(6) into Equations (7) and (8), the equations of meshing can be written as: \u03c6i = (yi c yi c\u03b2 + zi czi c\u03b2)xi c j \u2212 (yi c yi c j + zi czi c j )xi c\u03b2 rpi (xi c j yi c\u03b2 \u2212 yi c j x i c\u03b2) (9) \u03b2 = 0 (10) The dimensional parameters of the gear mechanism are listed in Table 1. A computer program was employed to draw the complete profile of the gear with convex tooth and the pinion with concave tooth. Using Equations (1), (2), (6), (9) and (10) with i = 1, the first tooth of the gear was created. The other teeth were copied from the first tooth and rotated at its axis by 2cy\u03c0/N1, cy = 1, . . . , N1, where N1 is the number of teeth of the gear. The complete contour of the gear is shown in Fig. 5a. Using Equations (2), (3), (6), (9) and (10) with i = 2, the first tooth of the pinion was created. The other teeth were copied from the first tooth and rotated at its axis by 2cx\u03c0/N2, cx = 1, . . . , N2, where N2 is number of teeth of the pinion. The complete contour of the pinion is shown in Fig. 5b. The assembly model of the pinion and the gear is shown in Fig. 5c. Using the proposed mathematical model and rapid prototyping technology, the pinion and the gear created by a ring surface were obtained and shown in Fig. 5d." ] }, { "image_filename": "designv11_101_0001421_iros.2012.6386014-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001421_iros.2012.6386014-Figure1-1.png", "caption": "Fig. 1. Gough-Stewart platform.", "texts": [], "surrounding_texts": [ "The Gough-Stewart platform [1], shown in Fig. (1), consists of a mobile platform (MP) and a fixed platform (FP) connected together by six actuated legs. The extremities of each leg are fitted with a 2-DOF universal passive joint at the fixed part and a 3-DOF spherical passive joint at the mobile part. A frame \u2126N+1 is attached to the MP at its center-ofmass (COM), denoted as ON+1 . Note that, with a perfect geometry, the COM is also geometric center of MP. A frame \u21260 is attached to the FP at the point O0. The configuration, i.e., orientation and position of the MP, is then defined by oriention of \u2126N+1 with respect to \u21260 and the position vector pN+1, the vector from O0 to ON+1. For each leg i, a frame \u2126i1 is attached to the leg at point Oi1, i.e., center of universal joint of leg i (see Fig. 2). Also, a frame \u2126i2 is attached to the MP at point Oi2, i.e., center of spherical joint of leg i. The following notations are used (see Figures (1) and (2)): Notations for the mobile and the fixed platforms ON+1: Center of mass (COM) of the Mobile Platform (MP) O0: Geometric center of the fixed platform \u21260: Base frame attached to the fixed platform \u2126N+1: Frame attached to the MP at its COM pN+1 \u2208 <3: Position vector of the COM of MP (position vector from O0 to ON+1) \u03c9N+1, vN+1 \u2208 <3: Angular and linear velocity of COM of MP VN+1 = [ \u03c9N+1 vN+1 ] \u2208 <6: Spatial velocity of COM of MP. \u03c9\u0307N+1, v\u0307N+1 \u2208 <3: Angular and linear acceleration of COM of MP V\u0307N+1 = [ \u03c9\u0307N+1 v\u0307N+1 ] \u2208 <6: Spatial acceleration of COM of MP. nN+1, fN+1 \u2208 <3: Moment and force exerted on the COM of MP FN+1 = [ nN+1 fN+1 ] \u2208 <6: Spatial force exerted on COM of MP mN+1: Mass of the MP JN+1 \u2208 <3\u00d73: Inertia tensor of MP about its COM IN+1 \u2208 <6\u00d76: Spatial inertia matrix of MP about its COM Notations for Leg i (i = 1, \u00b7 \u00b7 \u00b7 , 6) Oi1: Center of universal joint of leg i Oi2: Center of spherical joint of leg i \u2126i1: Frame attached to leg i at point Oi1 \u2126i2: Frame attached to MP at point Oi2 pi0 \u2208 <3: Position vector from O0 to Oi1 pi1 \u2208 <3: Position vector from Oi1 to Oi2 pi2 \u2208 <3: Position vector from Oi2 to ON+1 si1 \u2208 <3: Position vector from Oi1 to the center of mass leg i \u03b7i = pi1 \u2016pi1\u2016 \u2208 < 3: Unit vector of the leg i Qij , Q\u0307ij , Q\u0308ij \u2208 <3: Position, velocity and acceleration vectors of jth joint of leg i FiT j \u2208 <3: Control force vector at jth joint of leg i \u03c9ij , vij \u2208 <3: Angular and linear velocity of point Oij Vij = [ \u03c9ij vij ] \u2208 <6: Spatial velocity of point Oij \u03c9\u0307ij , v\u0307ij \u2208 <3: Angular and linear acceleration of point Oij V\u0307ij = [ \u03c9\u0307ij v\u0307ij ] \u2208 <6:Spatial acceleration of point Oij nij , fij \u2208 <3: Moment and force of interaction at point Oij ; ni1, fi1 are the moment and force of interaction between leg i and FP, and ni2, fi2 are the moment and force of interaction between leg i and MP. Fij = [ nij fij ] \u2208 <6: Spatial force of interaction at point Oij mi: Mass of leg i si1: Position vector from Oi1 to the center of mass of leg i Ji1 \u2208 <3\u00d73: Inertia tensor of the leg i about its COM Ii1 \u2208 <6\u00d76: Spatial inertia of the leg i with respect to point Oi1 Hij : Spatial-axis (map matrix) of leg i joint j." ] }, { "image_filename": "designv11_101_0000628_1.1735676-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000628_1.1735676-Figure1-1.png", "caption": "FIG. 1. Setup of the apparatus.", "texts": [ " 6 E. G. Richardson, Proc. Roy. Soc. (London) A61, 352-67 (1948). 6 Gilbarg and Anderson, J. App!. Phys. 19, 127 (1951). 7 A. May, J. App!. Phys. 22,1219 (1951). 8 A. May, J. AppL Phys. 23, 1363-72 (1952). Downloaded 10 Sep 2013 to 128.103.149.52. This article is copyrighted as indicated in the abstract. Reuse of AIP content is subject to the terms at: http://jap.aip.org/about/rights_and_permissions 728 A. G. DAVIS PHILIP DESCRIPTION OF THE PRESENT EXPERIMENT A diagram of the apparatus is shown in Fig. 1. A Wollensak Fastax camera, borrowed from the White Sands Missile Range, was used to photograph water drops as they fell into varying mixtures of glycerol and water. Free fall in a vacuum was used as a basis for scaling the velocities. A 20% variation in the drop velocity was noted on the film. No correction was made for air resistance as these corrections were smaller than other statistical fluctuations. The water level in the dropper tube was maintained at a predetermined level in order to keep the sizes of the drops constant" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000855_mace.2011.5988059-Figure4-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000855_mace.2011.5988059-Figure4-1.png", "caption": "Figure 4. Distribution of temperature.", "texts": [], "surrounding_texts": [ "1/2 3\nHeat Generation Rate HGEN\n(x0, y0, z0)\nvtz \u2212=0 . (1)\nv t\n(x, y, z)\n( )2 0 2 zzxDist \u2212\u2212= . 2\na g\n>= \u2264= aDistHGEN aDistgHGEN 0 . 3\nBirth and Death\nAPDL\nESEL,S,LIVE ! ESEL,INVE ! tive=v*time ! /POST1 ETABLE,centx,CENT,X x ETABLE,centz,CENT,Z y SMULT,centx,centx,centx !\nSADD,centz,centz,,,,tive SMULT,centz,centz,centz SADD,centx,centx,centz FINISH /SOLU ESEL,S,ETAB,centx,0,a2 a2 EALIVE,ALL ALLSEL,ALL", "III. 4\n720W 1mm/s 20\n5 6 7\n8\n9", "1Cr18Ni9 1400 [5]\n10 1400 700\n11 12 13 1400\nIV.\n1\n2\n3\n4\n3082007\nREFERENCES [1] L. Dubourg and J. Archambeault, \u201cTechnological and scientific\nlandscape of laser cladding process in 2007,\u201d Surface & Coatings Technology, vol. 202, pp. 5863\u20135869, 2008. [2] R. Jendrzejewski, G. liwi ski, M. Krawczuk, and W. Ostachowicz, \u201cTemperature and stress during laser cladding of double-layer coatings,\u201d Surface & Coatings Technology, vol. 201, pp. 3328\u20133334, 2006. [3] G. Zhao, C. Cho, and J. Kim, \u201cApplication of 3-D finite element method using Lagrangian formulation to dilution control in laser cladding process,\u201d International Journal of Mechanical Sciences, vol. 45, pp. 777\u2013796, 2003. [4] Y. Jia and N. Hao, \u201cThermal-mechanical Coupling Finite Element Analysis of Laser Cladding Process Part : Strain-stress Field,\u201d Proc. ICEICE2011, in press. [5] T. Zuo, Laser processing of high strength aluminum alloys. Beijing, Defence Industry publisher, 2002 (In Chinese)." ] }, { "image_filename": "designv11_101_0000363_msf.770.141-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000363_msf.770.141-Figure1-1.png", "caption": "Figure 1 Variable forming point grinding method a with cup wheel", "texts": [ " However, after the forming point on end face of cup wheel has been worn out, the two methods cannot maintain machining accuracy. [3,4] In view of this problem, this paper proposes a high efficiency and precision method that the forming point on end face of cup wheel is changed gradually as wear rule function during grinding. This grinding method improves the method for creep-feed face grinding revolving part of convex function generatrix with a cup wheel. Principle and procedures of variable forming point grinding method In this method, as shown in Figure 1, select a point A on end face of cup wheel as Forming point, make the end face of cup wheel maintain tangency contact with workpiece at Forming point through adjusting the axis of grinding wheel. The most important of this method is that with time and cutting amount varying, Forming point changes from point A to 'A gradually at un-uniform speed, which can always keep Forming point un-worn and increase suface processing quality As shown in Figure 2, the axis and displacement of grinding wheel are adjusted following the steps below: (1) install work-piece and grinding wheel on a three-axis machine tool that has three degrees of freedom in X-Z plane, 2 for pure translation and 1 for pure rotation, and then ascertain tool setting point that work-piece vertex p1 corresponds to " ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000282_icems.2011.6073966-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000282_icems.2011.6073966-Figure1-1.png", "caption": "Fig. 1 The PMAC motor with the rotor aligned asymmetrically in axial direction", "texts": [ ". INTRODUCTION For realizing some special characteristics, many PMAC motors have the rotor aligned asymmetrically in axial direction (Z-asymmetrical rotor). For example, many PMAC motors use Hall sensor to detect rotor position. To make the rotor magnetic field be detectable clearly, the rotor is aligned asymmetrically in axial direction, and this can make the sensors be close to the magnet; see Fig. 1. Another example is the spindle motor used in hard disk drive. For generating the preload to the bearing, the Z-asymmetrical rotor is also used; see Fig. 2. However, as the mechanical structure is axial asymmetrical, the magnetic field generated by the permanent magnet on the rotor acts with the stator core and generates an additional UMP in the motor operation. For the PMSM with Z-asymmetrical rotor, the magnetic field generated by the permanent magnet trays to align the rotor magnetic center with the stator magnetic center, and an UMP is thus induced" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000662_gt2011-45999-Figure5-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000662_gt2011-45999-Figure5-1.png", "caption": "Figure 5: Sensor Package", "texts": [ " A photograph of the completed sensor as well as a cross- sectional diagram is provided in Figure 4. The transceiver was also fabricated out of a PEEK ring with a channel milled out for the inductor wire, much like the passive sensor. A 1 meter coaxial cable was soldered to the ends of the inductor so as to allow an interface to the telemeter from outside of the bearing housing. This ring was then mounted on the inside of the bearing with the inductor placed 3mm away from the inductor on the cage sensor. Figure 5 shows the finished components of the telemeter and how they fit inside of the housing. The coaxial cable is not shown in this case. Sensor calibration procedure and relevant data has been published elsewhere [6]. National Instruments data acquisition hardware and custom LabVIEW software program were used to collect and analyze the displacement data. Similar to Adiletta et al. [3, 7], results from the experimental TTR investigation were corroborated with the displacements predicted by the analytical investigation" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002317_978-94-007-4201-7_5-Figure5.5-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002317_978-94-007-4201-7_5-Figure5.5-1.png", "caption": "Fig. 5.5 Virtual mechanism principle (a) a 3-RPS PM (b) virtual PM", "texts": [ " Like the KIC principle for 6-DOF PM, the VMP can also provide a uniformway to analyze the lower-mobility PM, regardless if the PM is symmetrical or not and howmanymobility it has. Thus, the process is time-saving. Themore complex the mechanism, the more effective the method. For lower-mobility PM, building square first-order KIC matrices is impossible and inverse kinematic analysis is difficult. To solve this problem, we proposed the VMP to build new reversible square KIC matrices so the velocity and accelerationanalyses can go through. In the current study, the spatial 3-DOF 3-RPS PM is taken as an example, Fig. 5.5a, to explain how to build the virtual mechanism for lower-mobility PM. To solve this problem, the virtual mechanism principle requires two conditions: 1. Transferring the lower-mobilitymechanism to a 6-DOF. To achieve this aim, some single-DOF kinematic pairs should be added to each limb until the number of the kinematic pairs is six. The added kinematic pair is called as \u2018virtual kinematic pair\u2019 and the corresponding mechanism is the \u2018virtual mechanism\u2019. The virtual kinematic pair can be in any orientation and located in any position relative to the limb", " The correctness of the virtual mechanism method is indubitable because when all the virtual inputs including velocities and accelerations all are given as zeros, none of the virtual links is moveable, the additional input links of the virtual mechanism all are fixed, and the virtual mechanism is just the real mechanism itself [1]. Many of our previous studies on lower-mobility PMs were based on VMP [12\u201318]. For the 3-RPS mechanism, one virtual revolute pair should be added to each limb, as shown in Fig. 5.5b. For the first revolute pair in the ith limb, the corresponding amplitude of velocity and acceleration are _f\u00f0i\u00de 1 and \u20acf\u00f0i\u00de 1 , respectively. Then, the original 3-DOF 3-RPS real mechanism is transferred into a virtual 6-DOF 3-RRPS chain. The sufficient and necessary conditions for the kinematic analyses of both mechanisms being equal are as follows: _f\u00f01\u00de 1 \u00bc _f\u00f02\u00de 1 \u00bc _f\u00f03\u00de 1 \u00bc 0 \u20acf\u00f01\u00de 1 \u00bc \u20acf\u00f02\u00de 1 \u00bc \u20acf\u00f03\u00de 1 \u00bc 0 (5.85) The virtual angular velocities and accelerations in a limb can be written as follows: _f1 _f2 _f6 T\u00f0i\u00de \u00bc 0 _f2 _Si _f4 _f6 T\u00f0i\u00de i \u00bc 1; 2; 3 \u20acf1 \u20acf2 \u20acf6 T\u00f0i\u00de \u00bc 0 \u20acf2 \u20acSi \u20acf4 \u20acf6 T\u00f0i\u00de i \u00bc 1; 2; 3 (5" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000654_sii.2012.6427383-Figure14-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000654_sii.2012.6427383-Figure14-1.png", "caption": "Figure 14. How to push tactile sensor", "texts": [ " The coefficient is used and the result of having presumed the value of power is shown in Fig. 12 and 13. However, since the data for study used for Fig. 12 and 13 differs, each coefficient differs. Fig. 12 shows the case where the force to which only 10 [deg] leaned the elastic body central portion to the left from perpendicular down is loaded. Fig. 13 shows the case where the force to which only 10 [deg] leaned the middle point of an elastic body central portion and a left end to the left from perpendicular down is loaded. Fig. 14 is a mimetic diagram showing the input method of force. Between 6 axis force-torque sensor and the sensor of this research, time and delay exist as a result of Fig. 12 and 13. Delay is considered that delay of picture taking in of the used camera is the cause in this time. This delay is considered that the error is producing the output of the force between 6 axis force-torque sensor and the sensor of this research. However, if this delay is canceled, the error between 6 axis force-torque sensor and the sensor of this research will decrease" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000264_2012-36-0457-Figure6-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000264_2012-36-0457-Figure6-1.png", "caption": "Figure 6 \u2013 Thermoelastic static loads.", "texts": [ "2, for fin ite element modeling; MD.NASTRAN Solver - version 2010.1.0, for the modal and static calculations; PERMAS Solver - version 13.00.216, fo r the dynamic calculations; Medina Pos-processor - version 8.2, fo r results analysis. For the modal analysis the frequency range used was 0 to 300 [Hz]. For the static analysis the pressure at the tube socket used was 2.1 [bar] which was measured in the vehicle engine. For the thermoelastic analysis, a components temperature variation list was obtained experimentally. Table 3 and Figure 6 shows the temperature gradient imposed. For the dynamic analysis, Acceleration x Time loads, in the X, Y and Z directions were applied. The loads were measured at the critical frequency range of 1541-1547 [rpm]. Figure 7 shows engine dynamic load graph. In modal analysis, the first assembly natural frequency should be equal to or greater than 182.00 [Hz], which is the 4 \u00bd order frequency of the 6 cylinder engine, for a maximum of 2200 [rpm] with a 10% safety factor, according to Campbell Diagram [5]" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000662_gt2011-45999-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000662_gt2011-45999-Figure2-1.png", "caption": "Figure 2: Introduction of Imbalance to the Rotor", "texts": [ "org/about-asme/terms-of-use Table 2: Modal response of turbocharger Mode 1(Hz) Mode 2(Hz) Mode 3(Hz) Mode 4(Hz) Mode 5(Hz) Turbocharger Assembly 1 1 1 1 1 Equivalent Assembly 0.9710 0.9344 0.9127 0.9862 0.9863 Rotor and Component Balance: The equivalent rotor assembly was dynamically balanced to ISO G2.5 standard. Each rotor part was independently balanced before assembly and later the complete rotor assembly was also balanced to the same specifications. Nevertheless, the rotor was designed to allow for adding external weights around the periphery to create controlled imbalance. Each plane has 12 equally spaced threaded holes, as shown in Figure 2, which can be plugged to add small amounts of controlled imbalance to the rotor and observe its effect on the rotor and bearing dynamics. Lubrication System: A recirculating lubrication system was used to supply pressurized oil (50 psi) to the turbocharger test rig. The lubricating oil is pumped through a heat exchanger to maintain the desired lubricant temperature before it is supplied to the turbocharger bearings, squeeze film damper and the gear box. The cooled oil is then filtered to remove contaminants particles of 10 \u00b5m and larger" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000655_imece2011-63367-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000655_imece2011-63367-Figure2-1.png", "caption": "FIGURE 2. HIGH SPEED LINEAR TEST RIG \u201dHiLiTe\u201d", "texts": [ " So the lab testing is predominant in all respects if the boundary conditions inside like temperature and ice or snow quality are similar to the test track outside, cp. [8]. In the laboratory the test set up should be kept as simple as possible. So the contact mechanics between tire and test track can be studied with a single tire tread block. In order to cover the typical parameter range of the tire tread block road interaction the High Speed Linear Test Rig \u201dHiLiTe\u201d was built at the Institute of Dynamics and Vibration Research (Figure 2). The maximum sliding velocity can be set to v = 10 m s . Normal forces can be adjusted up to N = 1000N. HiLiTe is powered with a synchronous servo motor with a standard power of P = 15,8KW and a nominal torque of MT = 53Nm. The tread block sample (Figure 1 c) is fixed under the carriage. The carriage performs a linear motion along a guide rail (cp. Figure 2). It also comprises a spring to apply the normal force and a 3D piezoelectric force sensor to measure friction and 2 Copyright c\u00a9 2011 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 01/03/2018 Terms of Use: http://www.asme.org/about-asme/terms-of-use normal force. The test rig can be equipped with arbitrary friction surfaces like glass, concrete, asphalt, snow or ice. The ambient temperature range is between \u221225\u25e6C \u2264 T \u2264 60\u25e6C. For further details see [1]. With the help of the High Speed Linear Test Rig rubber friction phenomena can be studied" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000721_s12541-013-0285-6-Figure4-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000721_s12541-013-0285-6-Figure4-1.png", "caption": "Fig. 4 (a) FEA model of mount I, (b) FEA model of mount II", "texts": [ " The last one is Marlow model with the least accuracy in this research. All constitutive constants calculated in the nine combination cases of the three stress-strain loading conditions and different maximum strain states are listed in Table 1, are listed in Table 2. Two kinds of mounts are chosen as objectives in this research. One kind of mount, mount I, bears stretching-compression deformation or pure shear deformation. The other kind of mount, mount II, bears both stretching-compression and pure shear deformations simultaneously. Rubber mount I in Fig. 4(a) can bear stretching-compression deformation in X direction and can bear pure shear deformation in Y direction. Finite element analysis (FEA) is applied to calculate the applied force and stiffness values in different directions when the deformation style is determined. Mooney-Rivlin model is used here and its constitutive constants are calculated from curves in combination loading cases. All individual loading combination cases are listed in Table 1, which include cases combined by one item from No", " 3 loading states has lower accuracy than the stiffness from the model calculated by only one curve in plane shear loading state (No. 9). From data in Table 3 in Z direction for compression and shear deformation, it can be observe that the stiffness calculated from the model fitted by two curves in uniaxial and equi-biaxial tension states has the highest accuracy, while, the stiffness from the model calculated by only one curve in equi-biaxial tension state has the least accuracy. FEA model of the main rubber spring in mount II is shown in Fig. 4(b), which bears stretching-compression and shear deformations simultaneously. In force-displacement relation calculation, force load ranges in X, Y and Z direction are: -3500N ~ 3500N; -500N ~ 500N and -1000N ~ 1000N, respectively. Like in Mount I, Mooney-Rivlin model is also used here and its constitutive constants are calculated from curves in combination loading cases. The obtained constitutive constants and the corresponding models are used in calculating force-displacement relation in the three directions" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000130_s0010952513030015-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000130_s0010952513030015-Figure1-1.png", "caption": "Fig. 1.", "texts": [ "= = III\u00b1 : x0 y0 z0, ,( ) x 0 z, ,\u00b1( ),= x2 R z+( ) 2+ R* 2 , \u03bb0 0, x2 z2+ 2 .= = = straint force is negative, if point \u0420 is located behind the Earth\u2019s center, between it and the geosynchronous orbit, which is obviously a fantastic situation. Solutions II correspond to the case when the first multiplier in the left hand side of equation (1.1) and the second multiplier in the left hand side of equation (1.2) are equal to zero. For such motions the pendu lum end should be located in the plane \u0445 = 0 on the curve presented in Fig. 1. This curve is symmetrical about axis Oz and intersects it at the point (0, 0, z*) determined from the equation R(1 \u2013 /|z + R|3) = \u2013z. For the Earth R \u2248 6378 km, R* \u2248 42164 km, and z* = 19917 km, and the tether critical length is determined by these values. The found point is located between the Earth\u2019s surface and geosynchronous orbit. The scale in Fig. 1 is R* = 1. Solutions III correspond to the case, when second multipliers in the left hand sides of equations (1.1) and (1.2) are equal to zero. For such motions point \u0420 is located on the geosynchronous orbit at a point located during the motion ahead or behind the pendu lum suspension point. In order to study the sufficient conditions of stabil ity, we again take advantage of the Routh method and investigate sign definiteness of the restriction of the second variation of the Routh function on a linear manifold determined by the constraint \u0445\u03b4\u0445 + y\u03b4y + z\u03b4z = 0", " In the original variables (\u0445, \u0443, z, \u03bb) these equations (cf. [5], p. 212, and [6]) allow one to get the characteristic polynomial in the form (4.1) The motion under study is stable in the first approximation, if both roots of the polynomial (\u03c3) from (4.1) are real and negative. The family of relative equilibria II\u00b1 forms on the plane Ozy a curve \u0393 speci fied by the equation R(l \u2013 ) = \u2013z from (2.2). Setting \u03c1* = ( R)1/4, it is convenient to represent this curve in the parametric form (4.2) The curve \u0393 specified in this way is plotted in Fig. 1. The motion under investigation is stable in the first approximation, if both roots of polynomial (4.1) are R* 3 R* \u20132 \u03c3( ) P \u03bc( ) a2\u03bc 2 a1\u03bc a0, \u03bc+ + \u03c3 2 ,= = = a2 y2 z2+( ), a1\u2013 \u20132y2 z2 3y2R* 3 R2 /\u03c1 5 ,+ += = a0 \u2013y2 1 R* 3 R2 /\u03c1 5+( ).= R* 3 /\u03c1 3 R* 3 z \u03c1( ) \u03c1* 4 /\u03c1 3 R, y \u03c1( )\u2013 \u03c1 1 \u03c1* 8 /\u03c1 8\u2013 ,= = \u03c1 \u03c1* \u221e, ).[\u2208 212 COSMIC RESEARCH Vol. 51 No. 3 2013 BUROV, KOSENKO real and negative, i.e., the following conditions are sat isfied At the branching point \u0443 = 0 discriminant d is pos itive, while quantity c1 is negative" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000105_978-3-642-20222-3_3-Figure3.57-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000105_978-3-642-20222-3_3-Figure3.57-1.png", "caption": "Fig. 3.57 A method of synthesis of the Vs vector in the first segment of the space vector plane", "texts": [ " The instantaneous position of vector Vs is determined by a phase angle \u03b8. The method of modulation using space vectors SVM will be presented on the example of the synthesis of vector Vs situated in the first sector of the voltage star of the inverter. In the other sectors the situation is similar, as a value of phase angle \u03b8 could be reduced to the range of the first sector. This vector is synthesized by adequately selected switching times of the states that determine vectors V1 and V2 as well as zero vectors V0 and V8. This is illustrated in Fig 3.57. The construction of short time averaged vector Vs involves the fact that within sufficiently short pulsation time Tp, which corresponds to a fraction of the total cycle, voltages V1, V2 and V0 or V8 are switched on for the selected duration tx, ty, tn. These intervals are obviously relative to the instant position of vector Vs determined by angle \u03b8. As a matter of simplification it is assumed that within a single pulsation time the angle \u03b8 is invariable. The determination of time intervals tx, ty, tn is performed using the relation: nyxp nnyxsp tttT tttT ++= ++= VVVV 21 (3" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001002_amm.86.850-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001002_amm.86.850-Figure2-1.png", "caption": "Fig. 2. Geometrical parameters and coordinat systems of SRSC", "texts": [ " 1, the first author introduced a novel slipping clutch which consists of an inner race, a set of cylindrical rollers, an outer race, and a cage [1] . The inner and outer races are manufactured with a special surface, in order to make spatial line contact with the skewed rollers. The centers of rollers are located on a uniparted hyperboloid. The skewed-roller slipping clutch (SRSC) possesses the following advantages: (1) Compact structure and hybrid component of both clutch and bearing; (2) Low torque ripple and noise; (3) Low temperature rise and long lifespan. Fig. 2 shows the coordinate systems and the geometrical parameters of this slipping clutch. The two coordinate systems (e.g. O-XYZ and o-xyz) are used in the analysis related to races and rollers, respectively. The geometrical parameters of the slipping clutch include: (1) the gorge circle radius, R; (2) the angle between roller and race axes, \u03a3; (3) the axial displacement of rollers, ZC, and (4) the radius, length and number of rollers, r, b and Nr. All rights reserved. No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of Trans Tech Publications, www", " Therefore, the axial equilibrium equation of either inner or outer races can be given as ( ) 0),(sincoscoscossin )0()()()( =\u00b1\u22c5\u22c5\u03a3\u2212\u22c5\u03a3\u00b1=\u00b1 \u222b \u222b\u222b ai b a oziprjotiprjra b oZir FdtdzztpNFdFN \u03b3\u00b5\u03b3\u00b5\u03b3 \u2213 (7) Based on the FEA solver developed by the first author [3-4] , the normal contact pressures pi(o)(t, z) are calculated by simultaneously solving the deformation-pressure Eq. 1 and Eq. 4 and the dynamic equilibrium Eq. 6 and Eq. 7. Furthermore, the von Mises stress can be calculated. In this paper, three roller profile modification methods, i.e. dub-off edge, one-arc and two-arc profiles, are introduced to reduce stress concentration occurring at the roller two ends, as shown in Fig. 7. The design parameters used in the calculation are RC=24mm(refer Fig. 2), \u03a3=35\u00ba, ZC/R=0.8, b=10mm, r=1.5mm, Nr=18, bp=8mm, r22=100mm (dub-off), r2=1500mm(one-arc), r22=400mm and r2=10000mm (two-arc). Distributions of Surface Contact Stress. Fig. 8 shows the surface contact stress and its contours of both the straight cylindrical roller and three types of profiled rollers. It can be seen that an obvious stress concentration occurs at the straight cylindrical roller ends (Fig. 8(a)). All the three roller profiles can significantly reduce stress concentration. Small stress concentration can still be observed for the dub-off profile (Fig" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003342_b978-0-08-096912-1.00019-8-Figure19.3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003342_b978-0-08-096912-1.00019-8-Figure19.3-1.png", "caption": "FIGURE 19.3 Dielectric losses.", "texts": [ " When an AC voltage is applied to a cable a leading charging current, Ic, flows. Also, because the cable insulation is not a perfect dielectric a small leakage current, Ir, flows in phase with the applied voltage. This leakage current causes losses in the dielectric which generate heat. The total no-load current flowing is the vector sum of leakage and quadrature currents. The equivalent circuit of a composite insulation may be represented by a parallel combination of resistance and capacitance as shown in Fig. 19.3 where M1 and M2 are different materials and the potential drop in each is given by: VM1 5 IrR1 VM2 5 Ic= j\u03c9C1\u00f0 \u00de and VM1 6\u00bc VM2 From Fig. 19.3, Ir 5 Ic tan \u03b4 The dielectric losses of a cable are given by: P5VIr 5VIc tan \u03b4 5V2\u03c9C tan \u03b4 and shows that the dielectric losses are influenced by the frequency, capacitance, the square of the voltage and the \u2018loss factor\u2019, tan \u03b4. Therefore during AC testing a considerable amount of power is absorbed in the insulation, causing heating and accelerating the ageing process. Therefore AC testing is used as a \u2018go/no go\u2019 test to determine whether or not the insulation can withstand the applied voltage" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000405_amm.121-126.433-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000405_amm.121-126.433-Figure2-1.png", "caption": "Fig. 2 CAD model of impeller and its casting mould", "texts": [ " Target part slip off entirely, and complex parts can be obtained by SLPP. Most of resin coated powder can be taken as material in SLPP. The resin films of coated sand will be overheated and carbonized while being heated up to 600\u2103, and the joint strength is decreased to nearly zero, then, the thermosetting property is completely destroyed[11]. Case study for rapid casting Complex parts can be rapid casted by using SLPP to manufacture casting mould. Impeller casting are presented as case studies. 3D CAD model of impeller\u2019s casting mould which is shown in Fig. 2 was designed on computer when reasonable casting process being determined. Fig. 3 shows the impeller\u2019s coated sand mould manufactured by self-developed laser RP system, while laser power is 45W and scanning speed is 6mm/s. After surface coating treatment for moulds, the metal impeller was obtained by centrifugal casting, as shown in Fig. 3. A self-developed five-axis Coordinate Measurement Machine (CMM) processed the surface measuring of complex parts. Worktable of CCM moves in X-Y plane and rotates around Z axis" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001600_sisy.2011.6034297-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001600_sisy.2011.6034297-Figure1-1.png", "caption": "Figure 1. Idealized motion of elastic body according to D. Bernoulli.", "texts": [ " The kinematic model of the system is created (direct and inverse kinematics) in Section 3. Section 4 analyzes a simulation example. Section 5 gives conclusion. II. DYNAMICS The source form of the Euler-Bernoulli equation of the elastic line of beam bending has the following form: 0)\u02c6( \u02c6 \u02c6 \u02c6 \u02c6 1,11,12 1,1 2 1,12 1,1 1,1 2 1,1 =\u2212\u22c5\u22c5+ \u2202 \u2202 \u22c5 xx dt yd m x y el\u03b2 . (1) Where ])[\u02c6( \u02c6 \u02c6\u02c6 1,11,12 1,1 2 1,11,1 Nmxx dt yd mM el \u2212\u22c5\u22c5= is the load moment, in these source equations encompassing only inertia, ][ \u02c6 \u02c6 \u02c6 2 1,1 1,1 2 1,11,1 Nm x y \u2202 \u2202 \u22c5= \u03b2\u03b5 is bending moment. See Fig. 1. The Euler-Bernoulli equation (1) is expanded in [3], [4] from several aspects in order to be applicable in a broader analysis of elasticity of robot mechanisms. In that case, the model of elastic line of the elastic link\u2019s first mode has the new form of the Euler-Bernoulli equation: 0 \u02c6 )\u02c6\u02c6( \u02c6 \u02c6 \u02c6 2 1,1 1,11,11,1 2 1,11,11,12 ,1 2 ,1 1,1 = \u2202 \u22c5+\u2202 + +\u22c5\u25ca++ \u22c5 + \u22c5 x yy Fjh dt yd H juk Tj j \u03b7 \u03b2 \u03b5 . (2) The model of the elastic line of this complex elastic robotic system is given in the matrix form: 0\u02c6\u02c6\u02c6\u02c6 2 2 =+\u22c5\u0398\u22c5\u25ca+\u22c5++\u22c5 \u03b5\u03b5uk T e Fjh dt yd H ", " Generally, coordinates 1tox , 1toy are the total of elastic deformations, but precisely, in geometrical terms, it is the total of projected elastic deformations on axes 1x , 1y respectively. Equation (29) has a significance as elastic deformation for each mode for Meirovitch [8] and his followers and in this way defined is entered in the total dynamic model. In this paper, as explained above, (29) has completely new meaning. Equation (29) is a solution of dynamic models, i.e. form of elastic lines in space of Cartesian coordinates. The motion of the considered robotic system mode is far more complex than motion of the body presented in Fig. 1. This means that the equations that describe the robotic system (its elements) must be also more complex than (28) and (29), formulated by Euler and Bernoulli. This fact is overlooked, and the original equations are widely used in the literature to describe the robotic system motion. This is very inadequate because valuable pieces of information about the complexity of the elastic robotic system\u2019s motion are thus lost. Hence, we should emphasize the necessity of expanding the source equations for the purpose of modeling robotic systems and this should be done as in [3], [4]" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001652_ecj.10392-Figure5-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001652_ecj.10392-Figure5-1.png", "caption": "Fig. 5. Path avoiding self-collisions.", "texts": [ " Namely, the method first calculates the locus Ln of the movable end of the n-th link, and extracts the inhibited points on Ln while examining whether the individual attitudes of the n-th link constituted by point pairs on Ln\u22121 and Ln collide with other links or not. Then, for each inhibited point S on Ln, a path that connects S1 and S2 on Ln is recalculated to generate a self-collision-free locus from Hn to Dn while assigning infinity to S as its distance from Hn. Here S1 and S2 are self-collision-free points on Ln, and inhibited point S is a point in the sublocus {S1, . . . , S2} included in Ln. As a consequence, the shortest path algorithm generates a sublocus from S1 to S2 that avoids inhibited point S. Figure 5 shows an example. In the figure, the attitude (P, S) of the n-th link collides with other links, and therefore the distance at point S is set to infinity, and the original sublocus {S1, . . . , S, . . . , S2} is replaced by the self-collision-free sublocus {S1, . . . , S\u2032, . . . , S2}. For collision-free points S1 and S2, the current implementation of the BFA selects them as points that are two distance units apart from inhibited point S. According to several simulation results, this method successfully removes self-collisions even in cases where manipulators must pass through narrow holes, as discussed in Section 6" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000256_amm.275-277.751-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000256_amm.275-277.751-Figure1-1.png", "caption": "Fig. 1. Overview of finite element modeling of vehicle lifting jack", "texts": [ " Establishment of vehicle lifting jack finite element model 3-dimensional modeling for vehicle lifting jack is conducted by PRO/E 3-dimensional mapping software, and finite element modeling is conducted for equipment by adopting HyperMesh v10.0; static strength analysis for vehicle lifting jack is conducted by using large general finite element analysis software ANSYS. According to the Figure data of vehicle lifting jack and the relationship Figures between parts, 3-dimensional modeling for vehicle lifting jack is conducted. Pease refer to Fig. 1 for the 3-dimensional modeling of vehicle lifting jack. All rights reserved. No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of Trans Tech Publications, www.ttp.net. (ID: 128.210.126.199, Purdue University Libraries, West Lafayette, USA-04/07/15,15:32:34) establish finite element modeling; all structures that contribute to overall rigidity and local strength of vehicle lifting jack will be taken into account. To ensure the accuracy of calculation, the finite element modeling of vehicle lifting jack mainly consists of hexahedron solid elements" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002030_ijhvs.2013.053008-Figure19-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002030_ijhvs.2013.053008-Figure19-1.png", "caption": "Figure 19 Shape of the natural vibrating mode at 18.07 Hz", "texts": [ "3 In Table 2 the creepage values at which the real part of the vibrating modes become positive is reported for the first four vibrating modes taking into account the three contact models. The performed analysis thus confirms the possibility of unstable torsional oscillations in the locomotive powertrain only in presence of a decreasing section of the creep force-creep curve at high creepages. To verify the capability of the implemented lumped parameters torsional model in reproducing the experimentally observed behavior of the vehicle powertrain, the numerically obtained vibrating modes have been compared with the experimental data. Figure 19 depicts the shape of the vibrating mode associated with the first calculated natural frequency of the driveline. The contact model represented by Curve 2 of Figure 7 is considered and creep force-creep curve is linearised in the neighbourhood of the condition highlighted in Figure 8 (longitudinal slip \u03b5x = 0.025) for all the wheels. The amplitude of the torsional vibrations is expressed in percentage with respect to the maximum value. The damped frequency of the first vibrating mode is 15.93 Hz (while its natural frequency is 18", " Torsional oscillations characterised by a frequency of about 15.9 Hz can be observed in the cardan shafts response predicted by the MB vehicle model (Figure 28-right). The maximum oscillations amplitude (20 kNm) is reached at the secondary front cardan shaft (Figure 27), while oscillations amplitude is similar for the other shafts. These results are in good agreement with the experimental data shown in Section 3 and reported in Figure 28-left and improve predictions offered by the torsional model presented in Section 4 (first vibrating mode, Figure 19). The implemented MB model in fact allows reproducing not only the torsional deformability of the driveline, but also the load transfers occurring during the manoeuvre. During the start-up, the wheelsets of the front bogie, especially the first one, are dynamically unloaded and, as a consequence, creepage of these wheelsets becomes higher. This has produced larger torsional vibrations in the secondary front cardan shaft. The results referred to a start-up manoeuvre on a low adherence rail (e.g., a rail sprayed with oil) are instead reported in Figure 29" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001733_iemdc.2011.5994905-Figure6-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001733_iemdc.2011.5994905-Figure6-1.png", "caption": "Figure 6. Rotor current and Stator flux in both the stator and rotor reference frame", "texts": [], "surrounding_texts": [ "position sensors Slip and rotor angle estimation are based on having a same parameter in two different reference frames, which in this case is rotor current. The rotor current in the rotor reference frame can be obtained by directly measuring the current at the rotor terminals, knowing that the rotor current is oscillating with the slip frequency, the angle of the measured current in its own reference frame is a good estimation of the slip angle. At the other hand, the rotor current in the synchronous reference frame is calculated using the machine voltage and flux equations; which is oscillating with the synchronous frequency. The difference between the rotor current angles in the rotor and synchronous reference frame will result in the rotor angle which can be used to transform any parameter from the synchronous frame to rotor frame and vice versa. Therefor by knowing the difference between the rotor current and the stator flux in the synchronous frame we can calculate the stator flux angle which is the d axis in the rotor frame. Figure (6) shows the stator flux vector and rotor current vector in the rotor, stator and stator flux synchronous reference frames. The equations regarding the calculation of the slip and rotor angles are as follows: )( r Ir s Ir s s r sslip \u03b8\u03b8\u03b8\u03b8\u03b8 \u2212\u2212== \u03a8\u03a8 (18) In (18) the angle of the stator flux in the stator frame \u03b8\u03c8s s can be calculated by the estimation of the stator flux in (8). Therefore we have: )(tan 1 \u222b \u222b \u2212 \u2212 = \u2212 \u03a8 s ss s s s ss s ss s IRV IRV \u03b1\u03b1 \u03b2\u03b2 \u03b8 (19) Also, the angle of the rotor current in the stator reference frame is calculated by using the estimated rotor current in the following equation: ) )( )( (tan 1 s ss s ss s s s ss s ss s ss Ir ILIRV ILIRV \u03b1\u03b1\u03b1 \u03b2\u03b2\u03b2 \u03b8 \u2212\u2212 \u2212\u2212 = \u222b \u222b\u2212 (20) Finally, the angle of the rotor current in the rotor frame can be directly calculated from the measured rotor currents. r r r rr Ir I I \u03b1 \u03b2 \u03b8 1tan \u2212= (21)" ] }, { "image_filename": "designv11_101_0000198_indin.2013.6622905-Figure9-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000198_indin.2013.6622905-Figure9-1.png", "caption": "Fig. 9. TowerPro Brushless Speed Controller [10].", "texts": [ " In this project, the PCB solution will be used only in the future for controlling the ControlledBioBall due to its low cost, since electronic devices are cheap and highly flexible because is easy to add additional features. The selected solution used for controlling the motor A28L was the commercial R/C brushless controller because is the cheapest and easiest solution that can be applied in this phase of conceptual design of the ControlledBioBall. The controller used was the commercial TowerPro Brushless Speed Controller [10] (fig. 9). It can send PWM pulses up to 8 kHz with a power supply of 10 VDC. However, for future implementation, PCB electronic circuit was already developed to control the motor without Hall sensors. Figure 10 shows the developed electronic circuit that will be used in the future to control the sensor less brushless DC Motor A28L. The signals from Hall sensors are replaced by the back EMF voltages from coils of the motor that are read by ADC ports. Resistances are used to ensure the back EMF in the range from 0 V to 5 V for ADC reading" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001712_0954406212438142-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001712_0954406212438142-Figure3-1.png", "caption": "Figure 3. Schematic diagrams of a symmetric model of a turbocharger with two identical journal bearings supported in rigid housings. (a) Cross-section through each bearing; CB is the bearing centre; CJS is the static journal centre; CJD is the dynamic journal centre; O is the centre of the stationary casing and ! is the speed of rotation of the rotor. (b) Schematic of the rotor-bearing system. (c) Co-ordinate system: and are the angular co-ordinates about s and r, respectively; J! is the angular momentum of the rotor with polar moment of inertia J about its spin axis z and l is the distance between the bearings.", "texts": [ " Although it is a gross simplicity, it avoids complicated mathematical modelling which enables better physical insight. Before conducting a numerical analysis of this system, a study is first conducted into a simple system containing a symmetric rotor mounted in rigidly-supported bearings rather than floating ring bearings. This allows analytical expressions to be derived which give physical insight into the interplay of the internal moments. The influence of the gyroscopic moment on the rest of the moments is revealed and the stability of the system is studied. Figure 3 shows schematic diagrams of a symmetric model of a turbocharger with two identical journal bearings supported in rigid housings, which is representative of the simple system described above. Figure 3(a) shows a cross-section through each bearing with the oil-film greatly exaggerated. Figure 3(b) shows the schematic of the rotor-bearing system with two identical bearings with rigid backing. Figure 3(c) shows the co-ordinate system (for clarity, the discs representing the turbine and the compressor at each end of the shaft are not shown). Only tilting motion and the moments due to the oil-film forces and the moments from the gyroscopic action are of interest here. The rigid rotor has a mass 2m and transverse moment of inertia I. It is mounted in two identical bearings of 360 oil-film (i.e. no cavitation or oil-film rupture11). The angular co-ordinates about the s and r axes are and , respectively, and l is the distance between the bearing centres", " Given the assumptions mentioned above, and for a balanced rotor, the equation of angular motion of the rotor is given by I\u20ach\u00fe J!!_h\u00fe t \u00bc 0 \u00f01\u00de where h \u00bc T in which T denotes the transpose, J is the polar moment of inertia of the rotor assembly and ! \u00bc 0 1 1 0 couples the two equations giving !_h \u00bc _ , _ T ; t \u00bc Ms Mr T ; Mr \u00bc Fsl and Ms \u00bc Frl are the moments about r and s axes, respectively, due to the oil-film forces Fr, Fs in the bearings; J!!_h is the gyroscopic moment due to the angular momentum J! about the spin axis z shown in Figure 3(c). When the spinning rotor tilts about the s axis by angle , it experiences a moment about r axis as shown in Figure 3(c). A similar moment causes angular acceleration of the system about the s axis due to a tilt about r. The gyroscopic effect cross-couples the angular motion about the axes s and r. The linearised oil-film forces are used here. These are determined by at UNIV OF MICHIGAN on June 20, 2015pic.sagepub.comDownloaded from at UNIV OF MICHIGAN on June 20, 2015pic.sagepub.comDownloaded from integrating the pressure profile given by the Reynolds equation based on the short-bearing approximation, and using the full-film condition" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000476_amm.419.795-Figure7-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000476_amm.419.795-Figure7-1.png", "caption": "Fig. 7 The pattern IV which meets the situation Er\u03b7 >", "texts": [ " Pattern IV: The stair vertex with the larger label shown as the right subfigure's i+1 in Fig. 6 keeps in contact with the track, and exerts a force on the driving wheel. While the stair vertex with the smaller label shown as the right subfigure's i in Fig. 6 keeps in no contact with the track. The criterion of pattern IV can be obtained as: 2 2 2 2 ( ) 2 E A E r r h b i h b l h r \u03c0\u03b7 \u03b6 \u2212 ++ + \u2264 \u2264 \u2265 or . (7) If the expression 2 \u03c0\u03b6 < is not satisfied simultaneously for the criterion of pattern III, as is shown in Fig. 7, pattern IV might be mistaken for pattern III. STATEFLOW [4] module in MATLAB is used to simulate the whole process of WT wheelchair robot' stair-climbing, as is shown in Fig. 8, which can realize the function similar to a finite state machine and make a determination for wheelchair's pattern in the stair-climbing process. All the abscissa of the curves in Fig. 8 is time t. The ordinate of subfigure a in Fig. 8 is\u03b3cr, and the left of the black line at about t=16(s) means stair-climbing process while the right of the black line means it finished" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001260_s11465-011-0225-z-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001260_s11465-011-0225-z-Figure1-1.png", "caption": "Fig. 1 A shield machine containing a cutterhead, the hydraulic cylinders of the thrust subsystem, a segment erector, etc.", "texts": [ "eywords redundant actuation, parallel mechanisms, five-bar mechanisms, segment erectors, shield machines The shield machine is a heavy construction machine used for tunnel excavation; it contains the cutter subsystem, thrust subsystem, segment erector subsystem, etc. (Fig. 1). When the shield machine is cutting and thrusting, a segment erector is adopted to place the pre-cast concrete segments to the required locations of the tunnel inwall in a safe, efficient, and precise manner. The classical segment erector has three degrees of freedom with the ability to implement the necessary erecting actions, although it cannot adjust the pose of the segment [1]. To improve the quality and efficiency of the segment erection, the modern 6-DOF (degree of freedom) erector is invented [2]" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001277_amm.391.72-Figure6-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001277_amm.391.72-Figure6-1.png", "caption": "Fig. 6 Model in recurdyn and follower angle \u03c6 and its velocity v", "texts": [ " 11, the pressure angle \u03b1 is expressed as 2 sin cos cos 1 (tan sin ) (17) The parameters of the spherical engine are given in Table 1 as follows: By the using the software MATLAB, we plot the spherical cam profile and the pressure angle as shown in fig.5. (a)Spherical cam profile (b) pressure angle Fig. 5 Spherical cam profile and pressure angle By loading datas generated by the MATLAB software into modeling software SolidWorks, using simulation software Recurdyn, the kinematic model of the spherical engine is built as shown in Fig.6. From the meshing Eq. 14, we know the contact lines are througn the origin point o and each point on it have the same relative velocity, so there is no glide between the cam and the follower when they are in meshing movement and the motion is pure roolling contact.The quality of the engagement is fine. From Eq. 15, we see the cam profile surface is developable ruled surface, so the manufacturing process of the surface is easy. From the expression of the pressure angle seen in Eq. 17, we conclude that the value of the pressure angle is just related to the rotary angle \u03b8, disrelated to the radius r of the follower, and the presure angle of each point on contact line as well as the driven force are the same" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001847_1350650111403037-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001847_1350650111403037-Figure2-1.png", "caption": "Fig. 2 Calculation of moment contribution from element", "texts": [ " Several forces act on the tilting pad and these forces must counterbalance each other, giving moment equilibrium. The forces originate from contact force F C, oil pressure forces F P, and oil shear forces on the pad surface F S. The balance of moment M on the pad around the pivot centre is calculated by summing the contribution from all elements n. S is the distance vector from element centre to pad pivot centre M \u00bc Xn e\u00bc1 F P e \u00fe F S e\u00f0 \u00de S e\u00bd F C SC \u00f08\u00de The elements of equation (8) are shown in Fig. 2. The pivot of the bearing is assumed frictionless and hence it has no influence on the pad angular equilibrium position. In practice, the pivot will always cause a resistance towards motion and this will require Mj j4M Resistance in order for the pad to adjust position according to the external loads. A typical relationship between M and pivot tilting angle is shown in Fig. 3. For all operating conditions, M \u00bc 0. M 6\u00bc 0 exists only mathematically and is encountered in the seeking of the parameters giving M \u00bc 0" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003368_b978-0-08-098332-5.00006-1-Figure6.16-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003368_b978-0-08-098332-5.00006-1-Figure6.16-1.png", "caption": "Figure 6.16 Self-excited induction generator. The load is connected only after the stator voltage has built up.", "texts": [ " machine, however, the induction motor has only one winding that provides both excitation and energy converting functions, so given that we want to get the terminal voltage to its rated level before we connect whatever electrical load we plan to supply, it is clearly necessary to provide a closed path for the would-be excitation current. This path should encourage the build-up of magnetizing current \u2013 and hence terminal voltage. \u2018Encouraging\u2019 the current means providing a very low impedance path, so that a small voltage drives a large current, and, since we are dealing with a.c. quantities, we naturally seek to exploit the phenomenon of resonance, by placing a set of capacitors in parallel with the (inductive) windings of the machine, as shown in Figure 6.16. The reactance of a parallel circuit consisting of pure inductance (L) and capacitance (C ) at angular frequency u is given by X\u00bcuL 1/uC, so at low and high frequencies the reactance is very large, but at the so-called resonant frequency (u0\u00bc 1/OLC ), the reactance becomes zero. Here the inductance is the magnetizing inductance of each phase of the induction machine, and C is the added capacitance, the value being chosen to give resonance at the desired frequency of generation. Of course the circuit is not ideal because there is resistance in the windings, but nevertheless the inductive reactance can be \u2018tuned out\u2019 by choice of capacitance, leaving a circulating path of very low resistance. Hence by turning the rotor at the speed at which the desired frequency is produced by the residual magnetism (e.g. 1800 rev/min for a 4-pole motor to generate 60 Hz), the initial modest e.m.f. produces a disproportionately high current and the flux builds up until limited by the non-linear saturation characteristic of the iron magnetic circuit. We then get balanced 3-phase voltages at the terminals, and the load can be applied by closing switch S (Figure 6.16). The description above gives only a basic outline of the self-excitation mechanism. Such a scheme would only be satisfactory for a very limited range of driven speeds and loads, and in practice further control features are required to vary the effective capacitance (typically using triac control) in order to keep the voltage constant when the load and/or speed vary widely. The term doubly-fed refers to an induction machine in which both stator and rotor windings are connected to an a.c. power source: we are therefore talking about wound rotor (or slipring) motors, where the rotor windings are accessed through insulated rotating sliprings" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003407_icepe-st.2013.6804316-FigureI-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003407_icepe-st.2013.6804316-FigureI-1.png", "caption": "Fig. I.", "texts": [ " Considering that different shooting distance result in inconsistent size of the spacer image, this paper uniform scale the image size to 256*256 pixels. After experiments and statistical, open Operation structural elements in a radius of 10 pixels is best. B. Perspective dis tortion correction (1) Correction model During image acquisition, if the optical axis of the imaging lens is not perpendicular to the plane where the spacer locates, it will lead perspective geometric distortion on the image plane, which affects the shape determination result, as shown in Fig. I. Therefore, there is a need for a perspective distortion correction. Assume that the polar coordinate of the undistorted ideal pixel on the image plane is Cp, g), The polar coordinate of distortion pixel is Cp', ()I), the distance between the vertical object plane and the lens is d, the distance from the imaging point on the object plane to the intersection of the optical axis and the object plane is c, while the object plane is tilted, the projections of which on the optical axis and perpendicular object plane are respectively a and b, focal length isf, so we can get Camera Geometrical Relationship p/c=f/d p'/b = f/(d -a) a2 + b2 = c2 From the formula (5)(6)(7) I fpb p= --jc -pa Notes b/c = sina, a/c = cosa, so I f psina p= f -pcosa (5) (6) (7) (8) (9) Formula (9) establishes the relationship between the geometric coordinates of the ideal image and that of the distorted image" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003879_0471238961.micrwoo.a01-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003879_0471238961.micrwoo.a01-Figure2-1.png", "caption": "Fig. 2. Schematic diagrams of a two-chambered MFC in cylindrical shape (a) and rectangular shape (b).", "texts": [ " A number of configurations for MFCs have been developed. These are run in batch, fed-bath, or continuous mode, and they have a variety of shapes. A two-chambered configuration is a typical MFC, which consists of an anode and a cathode chamber separated by a PEM such as Nafion (DuPont, Wilmington, Del.) (25,26) or Ultrex (27,28) so as to create a potential difference between them. The compartments can take various practical shapes (29,30). The schematic diagrams of two-chambered MFCs are shown in Figure 2. Organics are injected into the anode chamber under anaerobic conditions, while oxygen is supplied to the cathode chamber. Two-chambered MFCs are typically run in batch mode often with a chemically defined medium such as glucose or acetate solution to generate energy. Single-Chambered Microbial Fuel Cells. MFCs are difficult to scale up, even though they can be operated in either batch or continuous mode. Singlechambered MFCs are quite attractive for increasing the power output because they can be run without artificial aeration in an open-air cathode system and can reduce the internal ohmic resistance by avoiding the use of a catholyte as a result of combining two chambers" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002239_978-94-007-5006-7_3-Figure3.1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002239_978-94-007-5006-7_3-Figure3.1-1.png", "caption": "Fig. 3.1 ZYZ Euler angles", "texts": [ " The EAJs have the specific advantage that they make a unified representation of multiple-DOF joints and also provide computational efficiency in formulating dynamics algorithms, as highlighted in Sects. 4.2.1 and 6.3, respectively. Such advantage is not available by following the Euler Angle representation. The definition of Euler angles is first revisited in this chapter before the Euler-AngleJoints are introduced. According to Euler\u2019s rotation theorem (Shabana 2001), any three-dimensional spatial rotation can be described using three sequential angles of rotations about three independent axes. These angles of rotation are called Euler angles. Figure 3.1a\u2013c show the sequence of rotations, (a) by an angle 1 about ZM axis, (b) by an angle 2 about rotated YM axis, and (c) by an angle 3 about the current ZM axis. The OR-XRYRZR and OM-XMYMZM denote the reference frame and moving frame, respectively. The three angles 1, 2, and 3 are called the ZYZ Euler angles. In a similar way, one can define 12 such sets of Euler/Bryant angles depending on the sequence of axes about which the rotations are carried out. They are ZYZ, ZXZ, ZXY, ZYX, YXY, YZY, YXZ, YZX, XYX, XZX, XZY, and XYZ" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001094_robio.2011.6181622-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001094_robio.2011.6181622-Figure1-1.png", "caption": "Fig. 1. Walking on uneven terrain with the different inclination and height. The solid-line rectangle and small circle represent the boundary and the position of the foot, respectively.", "texts": [ " In this paper, the 3-D CS is defined as a novel navigational command set as follows: Definition 1: 3-D command state (3-D CS) is defined as 3-D CS \u2261 [T ss l/r T ds l/r Sl/r Ll/r Hl/r \u03c6l/r \u03c8l/r] where T ss l/r: single support time during left/right support phase; T ds l/r: double support time from left/right support phase to right/left support phase; Sl/r: sagittal step length of left/right leg; Ll/r: lateral step length of left/right leg; Hl/r: foot height of left/right leg; \u03c6l/r: foot pitch angle of left/right leg; \u03c8l/r: foot roll angle of left/right leg. Fig. 1 shows the walking on the uneven terrain, in which there exists a board with the different inclination and height compared to the flat terrain. The solid-line rectangle and small circle represent the boundary and the position of the foot, respectively. fl, fr, ff , and fb denote the distances from the foot position to each boundary, respectively. ol and ow denote the length and width from the foot position to the nearest point of the board, respectively and oh denotes the height of the board at the point" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000655_imece2011-63367-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000655_imece2011-63367-Figure1-1.png", "caption": "FIGURE 1. a) TYPICAL WINTER TIRE TREAD PATTERN, b) NOMENCLATURE OF TIRE TREAD BLOCK1, c) TIRE TREAD BLOCK SAMPLE", "texts": [ " of different actions m Result, based on measurement s Result, based on simulation x\u0307 Derivation to time t x\u2032 Derivation to space coordinate 1 Copyright c\u00a9 2011 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 01/03/2018 Terms of Use: http://www.asme.org/about-asme/terms-of-use Winter tires gain more and more attention due to safety reasons. They increase the traffic safety on wintry roads by increasing the friction potential on snow and ice. This is very important because a decreasing road friction level normally causes an increasing accident probability [1]. The winter tire pattern (Figure 1 a) is characterized by cuts within the single tread block, called sipes (Figure 1 b). The development of winter tires is strongly based on vehicle tests. They can only be done in specific areas and certain seasons around the world where cold climate prevails. To speed up the development, tires or their components can be investigated in the lab for the evaluation of the performance. On the other hand simulations can be done to study the friction potential of different tire constructions or the contact mechanics in general. An advantage of computing single parts of the tire like e", " So the contact mechanics between tire and test track can be studied with a single tire tread block. In order to cover the typical parameter range of the tire tread block road interaction the High Speed Linear Test Rig \u201dHiLiTe\u201d was built at the Institute of Dynamics and Vibration Research (Figure 2). The maximum sliding velocity can be set to v = 10 m s . Normal forces can be adjusted up to N = 1000N. HiLiTe is powered with a synchronous servo motor with a standard power of P = 15,8KW and a nominal torque of MT = 53Nm. The tread block sample (Figure 1 c) is fixed under the carriage. The carriage performs a linear motion along a guide rail (cp. Figure 2). It also comprises a spring to apply the normal force and a 3D piezoelectric force sensor to measure friction and 2 Copyright c\u00a9 2011 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 01/03/2018 Terms of Use: http://www.asme.org/about-asme/terms-of-use normal force. The test rig can be equipped with arbitrary friction surfaces like glass, concrete, asphalt, snow or ice", "org/about-asme/terms-of-use The deflection curve r = r(x, t), the bending angle \u03d5 = \u03d5(x, t) and the load q = q(x, t) depend on the space coordinate x and the time t. The density \u03c1 , the Young modulus E and the shear modulus G = E/3 are material constants. The dottet variables display the derivation of time, the dashed variables were derivations of the space coordinate x. The geometry of the beam is described by the second moment of area I I = w`3 12 (5) The corrected cross section area AK of the beam is (cp. [10]) AK = 6 7 A = 6 7 wl, (6) using the length of the beam ` and its width w (cp. Figure 1 b). The value of the correction factor is 6/7 as recommended by Cowper [11]. Equation 3 and Equation 4 are partial differential equations (PDE). Analytical solutions for PDE exist only for simple problems. Therefore PDE are generally solved numerically. In the next chapter one numeric solving technique is presented with the Finite Difference Method (FDM). The Finite Difference Method (FDM) is based on the approximation of differential quotients with difference quotients. The difference quotients can be derived from the Taylor series, cp", " Additionally the finite difference quotients are easy transferrable into code. Beam models are simplified descriptions of physical structures. In this chapter such a description is developed for a tire tread block. Simulation results of the model will be shown and compared with experimental results. In a first step only a single block element is regarded. The requirements of the introduced model are . displaying the geometry of the single tread block with its width w, block length ` and sipe depth d (Figure 1 a) . considering the rubber material parameter like Young\u2019s modulus E, shear modulus G and density \u03c1 . giving the possibility to implement an arbitrary number of single tread block elements . considering the interaction between the single block elements and surface but also between neighbored blocks . low computational load During friction experiments with HiLiTe on transparent ice the contact area was investigated with a camera. Lifting of the back of the element was often observed. Additionally Lindner [14] describes a pressure peak in the front of the block element, which occurs mainly on new samples" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002397_978-3-642-33926-4_57-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002397_978-3-642-33926-4_57-Figure1-1.png", "caption": "Fig. 1. A human with a robot walking helper (left) and the configuration of a human with a robot (right)", "texts": [ " Since a human walks synergistically with the robot, interaction forces and geometric constraint are included in the model during support phase. Furthermore, impact forces are considered in the model. To achieve stable walking, zero moment point (ZMP) is utilized and friction constraint is considered on reaction force from the ground. Finally, simulations of the human with robot walker, using the model, are performed to obtain optimal gait trajectories, human applied joint torques and helper robot support forces. A conceptual diagram of a human walking with a robot helper is shown in Fig.1 (left). The human model is composed of a torso, an arm and two identical legs. The arm is composed of an upper arm and a forearm. Each leg is composed of two links and a foot [14]. The robot walking helper [9,10] consists of a support frame and two wheels with motors. It provides good support and stability for human walking. For the design of the robot walking helper, the dynamics of human with helper is described below. Fig. 1 (right) shows the configuration of human with a robot walking helper in joint coordinates given by [ ]Tqqqq 821 = . From Lagrange\u2019s equations, the dynamic equations of motion of a human in single support phase can be expressed as yyxx FEFEuqGqqqCqqD \u2212\u2212=++ )(),()( (1) where [ ]Tuuuu 821 = is the human joint torques, D is the inertia matrix, C is the centrifugal and Coriolis matrix, G the gravity matrix, [Fx Fy] is the human applied force on helper, and [Ex Ey] the corresponding Jacobian matrices" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003315_b978-0-08-098332-5.00010-3-Figure10.6-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003315_b978-0-08-098332-5.00010-3-Figure10.6-1.png", "caption": "Figure 10.6 Hybrid (200 step/rev) stepping motor. The detail shows the rotor and stator tooth alignments, and indicates the step angle of 1.8 .", "texts": [ " By repetitively switching on the stator phases in the sequence ABCA, etc. the rotor will rotate clockwise in 30 steps, while if the sequence is ACBA, etc. it will rotate anticlockwise. This mode of operation is known as \u20181- phase-on\u2019, and is the simplest way of making the motor step. Note that the polarity of the energizing current is not significant: the motor will be aligned equally well regardless of the direction of current. A cross-sectional view of a typical 1.8 hybrid motor is shown in Figure 10.6. The stator has eight main poles, each with five teeth, and each main pole carries a simple coil. The rotor has two steel end-caps, each with 50 teeth, and separated by a permanent magnet. The rotor teeth have the same pitch as the teeth on the stator poles, and are offset so that the centerline of a tooth at one end-cap coincides with a slot at the other end-cap. The permanent magnet is axially magnetized, so that one set of rotor teeth is given a north polarity, and the other a south polarity", " The step angle is a property of the tooth geometry and the arrangement of the stator windings, and accurate punching and assembly of the stator and rotor laminations is therefore necessary to ensure that adjacent step positions are exactly equally spaced. Any errors due to inaccurate punching will be non-cumulative, however. The step angle is obtained from the expression Step angle \u00bc 360 \u00f0rotor teeth\u00de \u00f0stator phases\u00de The VR motor in Figure 10.5 has four rotor teeth, three stator phase-windings, and the step angle is therefore 30 , as already shown. It should also be clear from the equation why small angle motors always have to have a large number of rotor teeth. The most widely used motor is the 200 step/rev hybrid type (see Figure 10.6). This has a 50-tooth rotor, 4-phase stator, and hence a step angle of 1.8 (\u00bc360 / (50 4). The magnitude of the aligning torque clearly depends on the magnitude of the current in the phase-winding. However, the equilibrium position itself does not depend on the magnitude of the current, because it is simply the position where the rotor and stator teeth are in line. This property underlines the digital nature of the stepping motor. From the previous discussion, it should be clear that the shape of the torque\u2013 displacement curve, and the peak static torque, will depend on the internal electromagnetic design of the rotor" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001129_amr.230-232.554-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001129_amr.230-232.554-Figure3-1.png", "caption": "Figure 3 -2(a) sketch map of the force to worm.", "texts": [ " such as chamfer, fillet and keyway and so on . 3.2 Setting material properties of the model. the three parts are of steel, setting the Young's modulus as 199GPa and Poisson's ratio 0.27\u00b5 = 3.3 Setting the solid model, taking the constraints into account .only the X, Y, Z three directions of moving freedom, in order to apply Freedom of Motion to the model, establishing the cylindrical coordinate system, taking the theta direction instead of the DOF rotational degrees of freedom. Worm constraints as shown in Figure 3-2 (b), the worm Z-axis, limiting the freedom of R and Z , after the model of restraint shown in Figure 3-2 (a) below; planetary bound plane diagram in the Figure 3-2 (d), binding surface for the planetary gear center hole surface,the constraint model to planetary gear. Figure 3-2 (c) 3.4 Setting the load of the model .the worm is active body, so it has an input torque T1.In the Pro / MECHANICA, in the real model, it can only impose TLAP (Total Load At Point) torque, before the torque applied, first of all you should create surface point by the force then addingTLAP at this point, 3.5 Creating a contact region which is peculiar step of contact analysis. between the planetary gear and the worm or stator tooth contact part, establishing contact region, running the overall intervention at the same time, testing whether there is intervention between the parts", "7 definiting the analysis tasks, operating analysis, checking the \"include the contact region\"(include contact regions), selecting the single-channel to adaptation (Single-Pass Adaptive), selecting the \"Localized Mesh Refinement\", improving the contact region of the grid density. Operating analysis, after analysis the operation status from the maximum 7, the stress error is 0.4%. 3.8 The picture shows results. In the \"Display Settings\", select \"part / layer\", respectively, each part shows the stress distribution of the contact region. Figure 3-5 for the worm where the stress contours, Figure 3-6 for the planetary gear stress contours. Fig.3-5 Stress distribution in the worm Fig.3-6 Stress distribution in the planet 4.Result We can find theresult from Figure 2-5: 4.1 The worm stress on each contact region is uneven. Close to the torque input, the contact stress is maximum (1309MPa), as shown in Figure in the left. At the other side of the worm, the minimum contact stress (390MPa), as shown in Figure in the right. the ratio was 3.3. 4.2 Each contact region shape is long strips on the worm. At each side of the contact region, the stress is great, but the middle is less, showing a \"dumbbell-shaped\" distribution" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001230_gtindia2012-9586-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001230_gtindia2012-9586-Figure2-1.png", "caption": "Figure 2. GEAR BOX AND ITS ASSEMBLY", "texts": [ " Predictions are excellent for the data at the same rotational speed of testing as that of training. Experiments were performed on a Machinery Fault Simulator\u2122 (MFS) and a schematic diagram of it is shown in Figure 1. Present experiments were conducted to study of the fault detection and diagnosis in gears. In MFS experimental setup, 3\u2212phase induction motor is mounted to rotate the rotor, which is connected to gear box through a pulley and belt mechanism. The motor speed can be manually control by a controller. The gear box and its assembly are illustrated in Figure 2. In the study of faults in gears, three different types of faulty pinion gears namely the chipped tooth, the missing tooth and worn gear along with a healthy gear were used (illustrated in Figure 3 (a\u2212c)). The real-time data in time domain were recorded using a tri-axial accelerometer (illustrated in Figure 3 (d) with sensitivity: x\u2212axis 100.3 mV/g, y\u2212axis 100.7 mV/g, z\u2212axis 101.4 mV/g) and data-acquisition hardware. The measurements were taken for running speed of 10 Hz to 30 Hz in the intervals of 5 Hz for each of four gears with faults and no-fault conditions" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002278_978-1-4471-4628-5_3-Figure3.8-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002278_978-1-4471-4628-5_3-Figure3.8-1.png", "caption": "Fig. 3.8 Cam with knife-edge follower", "texts": [ " Moreover, the follower axis can be offset from the cam center of rotation. Usually, the cam designer represents the follower motion or travel in terms of the angle of rotation of the cam. This is achieved through the inversion method which generates the different cam profile curves such as the pitch and working curves and the pressure angles. The flat-faced follower arrangement is slightly complicated because the contact with the follower does not take place at a fixed offset. We will rather consider a simple knife-edge follower as shown in Fig. 3.8. The velocity of the cam at contact point in the vertical direction will be given by y(\u03b8) which is assumed to be known function determined from cam profile. For example, in a cam giving simple harmonic motion, y(\u03b8) = \u23a7\u23a8 \u23a9 h 2 ( 1 \u2212 cos \u03c0\u03b8 \u03b2 ) for 0 \u2264 \u03b8 \u2264 \u03b2 0, otherwise (3.8) where h is the maximum follower travel, \u03b2 is the angle associated with motion event, 0 < \u03b8 < \u03b2/2 is the rise duration, \u03b2/2 < \u03b8 < \u03b2 is the fall duration, and the rest is the dwell duration. The velocity and acceleration profiles are derived from the displacement profile through successive derivatives and by assuming unit angular speed", " Here, the cam is driven through a torque input and its angular Fig. 3.10 Schematic diagram of cam drive with torque input ,C CK R y inT \u03b8 ( )r \u03b8 RiseFall Dwell Follower m 1 1,P A 2 2,P A Mg 'R Friction xF yF speed is variable. In this case, the cam rotation speed depends on the load dynamics. The normal and friction forces from the follower oppose the cam rotation. Moreover, the self weight of the cam and its rotary inertia are also significant factors that decide cam speed. For knife-edge follower, we can refer to Fig. 3.8 where the normal and frictional forces are shown as FN and \u03bcFN with \u03bc as the coefficient of kinematic friction. The direction of the surface normal at contact point is defined by an angle \u03c6 which is related to the pressure angle and can be written in terms of the cam angular position. Thus, \u03c6 is a dependent variable. We can use other kinds of friction forces. The component of the forces along the follower motion direction is Fy = FN sin \u03c6 + \u03bcFN cos \u03c6 (3.10) If the follower reaction force is known, then FN = Fy/ (sin \u03c6 + \u03bc cos \u03c6)" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000054_20110828-6-it-1002.02539-Figure7-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000054_20110828-6-it-1002.02539-Figure7-1.png", "caption": "Fig. 7. Simulated Lego Mindstorms NXT in the arena.", "texts": [ " The Lego Mindstorms can play an important role in the first years of competition because competitors can easily prototype a robot that fulfills all the requisites in order to compete and also to promote the participation of those who are taking their first steps in the world of mobile robotics. Added to all the advantages of the Kit Lego Mindstorms, simulation can also be a very important aid, allowing competitors to test their approaches without accessing to real hardware. An example of a simulated robot prototyped with Lego Mindstorms Technology is shown in Figure 7. The modeling and simulation of the Lego Mindstorms NXT was previously presented Gonc\u0327alves et al. [2009]. The simulated robot is provided, in the simulator, with its position in the arena. If the competitor is working with real hardware its position can be given by an external external localization system for robots Moreira et al. [2001], provided by the organization. This system will identify the robots using a pattern that must be placed on top of each robot and can provide the position and orientation of the robot" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000476_amm.419.795-Figure4-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000476_amm.419.795-Figure4-1.png", "caption": "Fig. 4 Primary classification of the patterns", "texts": [ " Meanwhile we introduce a parameter l EH= to calculate the numerical of calibration, where EH is the length of the line EH where H is the point crossed by the line EH which is parallel to the stair slope line and the line AH which is parallel to the level ground. The criterion expressed by l for calibrating a stair can be deducted as: 2 2 2 2 2 2 2 2( 1)A A E E r h b r h bb b i h b r l i h b r h h h h + ++ + \u2212 \u2264 < + + + \u2212 . (1) The stair's calibration number i can be deducted as: 2 2 2 2 A Ehl r h b r b i floor h h b + + \u2212= + . (2) where ( )floor \u2022 is rounding function. Pattern Classification As stair's horizontal surface is tangential to the bottom surface of the driving wheel as shown in Fig. 4, can be used as a primary classification of the patterns. When the driving wheel moving in the horizontal plane as shown in Fig. 4, the parameter l keeps unchanged, and the relationship can be obtained as: 2 2 2 2 ( )E Ar r h b i h b l h \u2212 ++ + = . (3) Analysis and determination between pattern I and Pattern II Pattern I: The stair vertex with the smaller label shown as the left subfigure's i in Fig. 5 exerts a force on the track and then on the driving wheel. The criterion of pattern I can be obtained as: 2 2 2 2 cr ( )E Ar r h b i h b l h \u03b3 \u03b3 \u2212 + + + > \u2264 . (4) where \u03b3 is the angle formed by the line crossing from stair vertex to driving wheel's center and the line perpendicular to stair slope line, and \u03b3cr is the angle formed by the line crossing from the tangent point of track and driving wheel to driving wheel's center and the line perpendicular to stair slope line" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000764_amr.188.566-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000764_amr.188.566-Figure1-1.png", "caption": "Fig. 1. The structure of aerostatic thrust bearing", "texts": [ "General Universitaria, Sevilla, Spain-25/03/15,17:28:58) traditional film area to orifice up stream, which would be more consistent with the actual condition. For the finite element method used to solve fluid motion equation, it is a effective method to solve the non-linear problem of aerostatic thrust bearing, which could avoid over-simplified analysis of the problems in traditional method. This paper uses circular-disk surface thrust bearing, with groove in order to raise load capacity and stiffness of the thrust bearing. Its structure is shown in Fig.1. Table 1 shows the principal bearing dimensions of aerostatic thrust bearings treated in this paper. In the structure of aerostatic thrust bearing, orifice is inherently compensated. Aerostatic thrust bearing studied in this paper is installed in pairs on the thrust plate in the back of shaft. This paper considers flow region of one aerostatic thrust bearing as solution domain, which would reduce the scale of computing operation as shown in Fig.2. In relation to the diameter of orifices and film thickness, the depth and diameter of feed channel are much larger, and the pressure loss is quite small in the conveying process from gas source to orifices" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000762_1.3555500-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000762_1.3555500-Figure2-1.png", "caption": "Fig. 2. Frame with suspended weight and coordinate board.", "texts": [ "1063/1.1863792 This article is copyrighted as indicated in the article. Reuse of AAPT content is subject to the terms at: http://scitation.aip.org/termsconditions. Downloaded to IP: 141.161.91.14 On: Thu, 27 Aug 2015 01:46:37 156 The Physics Teacher \u25c6 Vol. 49, March 2011 DOI: 10.1119/1.3555500 to the frame with large binder clips or spring clamps; this coordinate board is just a large sheet of laminated graph paper. Typical experiment setup A typical arrangement with a hanging mass is shown in Fig. 2. The arrangement is sized to fit in the working area of 5 ft x 5 ft and a typical mass is 10 kg. With a dry erase marker the students mark the cord angles on the coordinate board clamped to the frame. In our labs whenever angles need to be measured, we use the Cartesian coordinate boards to in- Christy Heid and Donald Rampolla, Chatham University, Pittsburgh, PA Many illustrations and problems on the vector nature of forces have weights and forces in a vertical plane. One of the common devices for studying the vector nature of forces is a horizontal \u201cforce table,\u201d1 in which forces are produced by weights hanging vertically and transmitted to cords in a horizontal plane" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002317_978-94-007-4201-7_5-Figure5.4-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002317_978-94-007-4201-7_5-Figure5.4-1.png", "caption": "Fig. 5.4 A 6-6R Multi-Loop Mechanism (a) sketch (b) one branch", "texts": [ " The multi-loop robot is composed of N branches. When N \u00bc 1, the mechanism becomes an open-loop manipulator. When N \u00bc 2, the mechanism is a single-loop chain. When N > 2 the mechanism is a multiloop device. When N > 6, at least N-6 branches are passive, which means no actuator is found in the N-6 branches. The platform has six DOF, so at least six inputs are necessary. Theoretically, any six single-DOF kinematic pair from this mechanism may be freely selected as inputs to make its motion well-controlled. The 6-6R mechanism, Fig. 5.4a, is a typical 6-DOF parallel manipulator[8] consisting of six limbs. Each limb has six single-DOF revolute pairs. Its kinematic analysis is an important theoretical foundation suitable for all parallel mechanisms, including 6-DOF and lower-DOF parallel mechanisms. In the following two sections, we focus on the kinematic analysis of the 6-6R parallel mechanism. In 1985, Huang made an important contribution by first setting both the velocity and acceleration analyses of the complicated 6-6R parallel mechanism using KIC method, setting its dynamic modal, and analyzing a numerical example [5, 6]. In this section, we introduce the KIC principle. Let us first discuss the velocity analysis and derive the first-order-KIC. Initially, let us consider one of the limbs which is a serial chain, Fig. 5.4b [6]. The absolute angular velocity of the platform can be expressed as vh \u00bc \u00bdGh \u2019 _w (5.34) where vh \u00bc ohx; ohy; ohz T ; _w \u00bc _\u20191 _\u20192 _\u20196f gT ; and _\u2019i , the ith component of the generalized velocity vector, is either yi or _Si depending on whether joint i is a revolute or prismatic joint. Gh \u2019 h i is a 3 6 first-order KIC matrix defined as Gh \u2019 h i \u00bc @oh @ _\u20191 @oh @ _\u20192 @oh @ _\u20196 (5.35) For the serial open chain, the first-order KICs can be evaluated as Gh \u2019 h i \u00bc Si; for revolute pair 0; for prismatic pair (5.36) If the joints are all revolute pairs as shown in Fig. 5.4b, Gh \u2019 h i \u00bc \u00f0S1 S2 S6\u00de (5.37) In Eqs. (5.36) and (5.37), Si is the unit vector of the i th revolute pair. The linear velocity of an interested point on the central platform, say point P, is given by VP \u00bc GP \u2019 h i _w (5.38) where VP \u00bc \u00f0vPx vPy vPz\u00deT (5.39) GP \u2019 h i is a 3 6 matrix corresponding to the first-order KIC of point P GP \u2019 h i \u00bc @VP=@ _\u20191 @VP=@ _\u20192 @VP=@ _\u20196\u00f0 \u00de (5.40) For the serial open chain, the column i of this matrix can be expressed as GP \u2019 h i :i \u00bc Si \u00f0P Ri\u00de i n i revolute pair Si i n i translation pair 0 i>n 8< : (5" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001455_2012-36-0254-Figure14-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001455_2012-36-0254-Figure14-1.png", "caption": "Figure 14. Journal bearing", "texts": [], "surrounding_texts": [ "The replacement of journal bearings by rolling bearings has proved to be a good solution for efficiency increase, leading to CO2 emission reductions and allowing engine downsizing. Other benefits are the increase of load carrying capacity, simpler and smaller oil supply system. However, it includes a step on the valvetrain assembly process, as the needle bearings must be press-fitted to the housing. The functionality of the system is not impaired. Needle roller bearings rating life, safety factor and hertz pressure have acceptable levels. This is also true for the static and fatigue safety factors of the camshaft and the stresses in the camshaft housing. It is important to highlight that analysis hereby conducted considered only the maximum camshaft rotation speed, as it is the most critical bearing life time and system structural analysis, due to the higher cam contact forces and higher number of revolutions of the system. In order to obtain more reliable results for the rolling bearing (rating life and safety factors) a complete duty cycle should be considered." ] }, { "image_filename": "designv11_101_0000282_icems.2011.6073966-Figure8-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000282_icems.2011.6073966-Figure8-1.png", "caption": "Fig. 8 can be used to express the way to calculate [X(\u03b1), Y\u03b1)], the location coordinate of the UMP center. The following equations can thus be obtained,", "texts": [], "surrounding_texts": [ "For the PMSM with Z-asymmetrical rotor, the magnetic field generated by the permanent magnet trays to align the rotor magnetic center with the stator magnetic center, and an UMP is thus induced. Such an UMP may contain the components in different directions, but only its component in axial direction is considered in this paper, as it is clear, this is the dominant component of this kind of UMP. There are many kinds of PMAC motors [3][4]. In this paper, the analysis concentrates on the motors with surface PM ring. This kind of motor is widely used in low power applications, e.g, the spindle motor used in hard disk drive, and the cooling fan motor used in PCs. Fig. 3 A simplified model for describing the PMAC motor with Zasymmetrical rotor (inner rotor) Fig. 4 A simplified model for describing the PMAC motor with Zasymmetrical rotor (Outer rotor) III. INFLUENCE OF STATOR SLOTS TO THE UMP INDUCED BY Z-ASYMMETRICAL ROTOR The magnetic field on the axial edge of the motor is related certainly with the permanent magnets on the rotor, and also the stator core structure, especial the slots on the core. If there is not slot on the stator, the geometrical relationship between the rotor and stator is fixed, i.e., from the viewpoint of rotor, the magnetic relationship between the rotor and stator is not affected by the rotor position. In this case, as the field generated by the armature winding currents in the axial edge is normally much weaker then the field generated by the magnet, the magnetic pull produced by the permanent magnet is constant. It also means, for the slotless PMAC motor, if the UMP in one rotor position can be known, the UMP on the other positions can also be known. However, the existence of the stator slots makes the geometrical relationship between the rotor and stator be complicated. The influence of the edgy magnetic field varies with the rotor rotation. This makes the UMP varies with the rotor rotation. One example is the motor shown in Fig. 5 and Fig. 6. It is a spindle motor with outer rotor. From the viewpoint of the magnetic circuit, the effective airgap of the motor axial edgy can be expressed with the simplified model shown in Fig. 7. From he viewpoint of the rotor, the length of the local airgap changes when the rotor is at different position. IV. UMP CHARACTERISTIC IN THE SPACE DOMAIN From Fig. 7, the effective permeance of the axial edgy airgap can be expressed as 0 0 ( ) [1 ( )]\u03b8 \u03bb \u03b8 = \u039b = \u039b + \u22c5\u2211A A An n Cos nZ , (1) where, \u039bA0 is the effective average permeance of the airgap, which is related with \u03b5, the difference between the lengths of rotor and stator; see Fig. 3 and Fig. 4. \u039bA0 is also linked with the thickness and permeability of the magnet, and the stator slot structure. Z is the number of stator slot. \u03bbAn, is the coefficient of the nth order unit permeance of the airgap. In the motor operation, the magnetic-motive-force (mmf) generated by the PM ring can be described by, ,2 1 0 ( , ) [(2 1) ( )]\u03b8 \u03b1 \u03b8 \u03b1\u2212 = = \u2212 +\u2211A A n n F K Sin n p , (2) where, \u03b8 is the position of the field, p is the pole-pair of the PM ring, and \u03b1 is rotor phase difference to the stator reference point. It is clear, \u03b1 is a function of time and rotor speed. Using magnetic-circuit method, the effective magnetic flux density in the airgap can be expressed as ( , ) ( ) ( , )\u03b8 \u03b1 \u03b8 \u03b8 \u03b1= \u039b \u22c5A A AB F . (3) The local radial force area density at position \u03b8 can thus be calculated by 20( , ) ( , ) 2 \u03bd\u03b8 \u03b1 \u03b8 \u03b1=A Ap B . (4) where, \u03bd0 is the reluctivity of vacuum. Therefore, the UMP, PA(\u03b1), can be calculated by using the following equation, 2 2 20 0 0 2 2 20 0 ( ) ( , ) ( , ) 2 ( ) ( , ) 2 \u03c0 \u03c0 \u03c0 \u03bd \u03b5\u03b1 \u03b5 \u03b8 \u03b1 \u03b8 \u03b8 \u03b1 \u03b8 \u03bd \u03b5 \u03b8 \u03b8 \u03b1 \u03b8 = = = \u039b \u222b \u222b \u222b A A A A A A A A A A A RP p R d B d R F d , (5) where, \u03b5A is the effective length difference between the stator core and permanent magnet; see; see Fig. 3 and Fig. 4. RA is the average radius of the airgap. From (1) and (2) it can be known that, 2 0 ( ) ( )]\u03b8 \u03b8 = \u039b = \u22c5\u2211A Am m l Cos mZ , (6) and 2 2 0 ( , ) [(2 ( )]\u03b8 \u03b1 \u03b8 \u03b1 = = \u22c5 \u2212\u2211A A n n F f Cos np . (7) The definitions of lAm and fAn can be found in [5]. Using the results shown in (6) and (7), the UMP can be expressed as, 0 2 20 0 0 ( ) 2 [ ( )] [2 ( )] \u03c0 \u03bd \u03b5\u03b1 \u03b8 \u03b8 \u03b1 \u03b8 = = = \u00d7 \u22c5 \u22c5 \u2212\u2211 \u2211\u222b A A A Am A n m n RP l Cos mZ f Cos np d . (8) From the orthogonality of the triangular function, it can be known that the integrations of all the items in (8) are zero except the items which can meet the following condition, 2=mZ np . (9) Therefore, the following result can be obtained, 0 2 0 0 0 0 (2 )] ( )] ( ) [ 22 [ 2 \u03b1 \u03bd \u03b5 \u03c0 \u03bd \u03b5 \u03b1 \u03c0 \u03b1 = = = = = + \u2211 \u2211 \u2211 A A A Am A n m n A A A Aq Aq q P R l f mZ Cos np Cos n q p RP l f , (10) where, PA0 is the zero order of the UMP, and q is the common multiple of Z and 2p. Equation (10) shows clearly that, in the motor operation, the UMP induced by z-asymmetrical rotor is not a constant, and its frequency of fundamental harmonic in space domain is the minimum common multiple of Z and 2p. This result is very helpful in analyzing the vibration measurement results of PMAC motor. PA0 is normally the maximum component in the UMP, and its value can be predetermined by the lengths of rotor magnet and stator core, and the specifications of magnet material. The equation shows also that, selecting the reasonable matching between the numbers of stator slot and magnetic pole-pair of the motor can increase the order of the fundamental harmonic of UMP, and this can normally reduce the amplitude of UMP variation. V. VARIATION OF UMP CENTER IN MOTOR OPERATION If the location of the axial UMP center varies in the motor operation, even if its UMP is constant, the center variation can still induce acoustic noise and vibration, especially to the motors whose ratio of motor diameter to length is big, like the spindle motor used in hard disk drive. As it was analyzed in Section-IV, in normally, the UMP is not constant in the motor operation. If the UMP center varies in the motor operation, the UMP influence becomes very complicated. It is necessary to investigate which factors affect the variation of the UMP center. When the distribution of the local z-asymmetrical axial force in tangential direction is known, and 2 0 0 0 ( , ) ( ) ( ) ( ) ( , ) ( ) ( , ) \u03c0 \u03c0 \u03c0 \u03b5 \u03b8 \u03b1 \u03b8 \u03b8 \u03b1 \u03b1 \u03b8 \u03b1 \u03b8 \u03b8 \u03b8 \u03b1 \u03b8 = = \u222b \u222b \u222b A A A A A A A A R p R Sin d Y P R p Sin d p d . (12) From (6) and (7), the integrations on the numerator of the right side of (11) and (12) can be written as 2 20 0 0 [ ( )] [(2 ( )] ( ) \u03c0 \u03b8 \u03b8 \u03b1 \u03b8 \u03b8 = = \u2212\u2211 \u2211\u222bA Am A n m n R l Cos mZ f Cos np Cos d , (13) and 2 20 0 0 [ ( )] [(2 ( )] Sin( ) \u03c0 \u03b8 \u03b8 \u03b1 \u03b8 \u03b8 = = \u2212\u2211 \u2211\u222bA Am A n m n R l Cos mZ f Cos np d . (14) If the stator slot is even, mZ is also even, and the functions to be integrated in above two equations can only form the odd triangular functions, and it is clear, the integrations for these triangular functions are zero. That is, if the stator slot is even, the following results can be obtained, ( ) 0 ( ) 0 \u03b1 \u03b1 =\u23a7 \u23a8 =\u23a9 X Y . (15) Equations (15) means also, for the PMAC motor possessing even stator slot, its UMP center is always located at the center of the motor. If the stator slot is odd, the integration of (13) and (14) can generate the following formats, 2 1 2 1 1 1 (2 ) (2 )\u03b1 \u03b1= + = \u2212 = = \u22c5 + \u22c5\u2211 \u2211n np mZ n np mZ n n U Cos np V Cos np (16) and 2 1 2 1 1 1 (2 ) (2 )\u03b1 \u03b1= + = \u2212 = = \u22c5 + \u22c5\u2211 \u2211n np mZ n np mZ n n U Sin np V Sin np , (17) where Un and Vn are obtained by integrating (13) and (14), separately, and m is any of integers. Equations (16) and (17) show that, when Z is odd, the fundamental orders of the numerators of (11) and (12) are the multiple number of the magnetic poles of the motor. In this case, as the denominators of (11) and (12) are formed by the even harmonics whose order is also the multiple of the magnetic poles, it can be concluded that the variation of the UMP center is formed by the even harmonics in both X and Y directions, and the order of the harmonic is the multiple of the motor magnetic poles. VI. NUMERICAL ANALYSIS ON THE UMP For verifying the analytical results obtained in the Section IV and V, several PMAC motors are calculated with 3D finite element method (FEM), and the FEM results are compared with the analytical ones. Here, the calculation results for two spindle motors are introduced, and they are named as Motor-A and Motor-B, separately. Motor-A has 12 stator slots and 4 magnetic pole-pairs, whose structure has been shown in Fig. 5 and Fig. 6. Motor-B is shown in Fig. 9. It has 3 stator slots and 2 magnetic pole-pairs. Both these two types of motor EM structure can be found in many applications. Fig. 10 shows the flux lines of Motor-A obtained with 3D FEM. From the analytical analysis in the Section-IV, as Motor-A has 12 stator slots and 4 magnetic pole-pairs, its minimum common multiple of Z and 2p is 24. Therefore, the cycle width of fundamental harmonic of the UMP is 15\u00b0. This analytical deduction is confirmed by the UMP curve shown in Fig. 11, which is obtained from the 3D FEM results. Fig. 10 The flux-lines of Motor-A For checking the variation of the UMP center, FEM is also used for the center position calculation, and the results of Motor-A are shown in Fig. 12. As the motor has 12 slot and 4 magnetic pole-pairs, the rotor position range from 0\u00b0 to 50\u00b0 (mechanical degree) is enough for checking the variation cycle of its UMP. For this motor, the number of its stator slots is even, therefore, from the analytical analysis in the section-V, the center of UMP shouldn\u2019t change in the motor operation, and it should always locate at the center of the motor. It seems that the FEM results shown in Fig. 12 express the variation of the UMP center. As the average radius of Motor-A\u2019s airgap is 14.35 mm and the variation of the axial UMP center is less than 0.07 mm, the center variation shown in Fig. 12 can be considered as zero. It should be pointed out, the curves in the figure varies randomly; they are actually the numerical error which is ineluctable in using FEM. Using optimized FE mesh density and element shape can reduce such an error. The FEM result of Motor-B\u2019s flux lines are shown Fig. 13, and motor\u2019s UMP is shown in Fig. 14. As the motor has 3 stator slots and 2 pole-pairs, the minimum common multiple of Z and 2p is 12. Therefore, the cycle width of UMP\u2019s fundamental harmonic is 30\u00b0, and this is confirmed by the UMP curve shown in Fig. 14, which is obtained with FEM. The average radius of the Motor-B airgap is same as MotorA, i.e., 14.35 mm. As its Z=3 and p=2, from the analysis in the Section-V, the location of the UMP center varies in the motor operation, and it is confirmed by the FEM results shown in Fig. 15 and Fig. 16. It is clear, as the variation is in a range whose maximum distance to the motor axis is 4.11mm, comparing with the radius of the motor airgap, the influence of such an UMP center variation is not neglectable in many applications. From Fig. 15, it can be known that, the UMP center varies 4 times in one rotor revolution, and this fundamental order is just 2p, the magnetic pole-pair of Motor-B. This confirms also the analytical deduction to the UMP center variation in the Section-V. From the calculation results of Motor-A and Motor-B, it can be known that, the axial UMP of Motor-A contains weak ripple than Motor-B. The reason of such a phenomenon is that, Motor-A has 12 slots and Motor-B has only 3. As increasing the slot number and magnetic pole-pair can make the minimum common multiple of slot and magnetic pole-pair be big, the order of UMP fundamental harmonic can thus be increased. This measure, in generally, can reduce the variation of the axial UMP." ] }, { "image_filename": "designv11_101_0001113_amm.433-435.17-Figure5-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001113_amm.433-435.17-Figure5-1.png", "caption": "Fig. 5. Stress at the moment of PGS failure.", "texts": [ "2 r/min, while obvious fluctuation exists and relative fluctuation reaches 175.2 % under the failure condition. The obvious speed fluctuation of planet gear can lead to large impact force to gear teeth. Needle roller fracture, which causes PGS to be \u2018locked\u2019, is predicted to be the reason of PGS failure. Therefore, in the dynamic simulation, a rotational joint between planet gear and PGS is replaced by fixed joint. The rotational speeds and torques are from Table 2 and the simulation time is set for 0.001s. The stress at the moment of PGS failure is shown in Fig. 5. The maximum stress of PGS reaches 1.31 GPa and it can cause the failure of PGS. As shown in Fig. 5, the red area is easy to emerge failure and it is the same as the actual fracture location. At the moment of PGS failure, the maximum impact force on PGS in the actual test is larger than it in the simulation, therefore PGS failure is more likely to occur under actual complex condition. To avoid PGS failure, it is recommended that the bearing cage should be installed and good lubrication should be ensured. This paper provides an efficient rigid-flexible coupled dynamic simulation method. A new threshold function algorithm for wavelet denoising is used to denoise test torque and rotational speed data" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003607_978-3-642-23026-4_10-Figure10.56-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003607_978-3-642-23026-4_10-Figure10.56-1.png", "caption": "Fig. 10.56 See Exercise (34)", "texts": [ "5 N and area A = 0.2 m2 is placed over a thin film mg d = constant fvis Lubricant A (33) How fast will an aluminum sphere of radius 1 mm fall through water at 20\u25e6C once its terminal speed has been reached, see Fig. 10.55? Assume that the msg Fvis t FB (34) The viscous force on a liquid flowing steadily through a cylindrical pipe of length L is given by: Fvis = 4\u03c0\u03b7Lvm where \u03b7 is the viscosity of the liquid and vm is the maximum speed of the liquid which occurs along the central axis of the pipe, see Fig. 10.56. If the pressures in the rear and front horizontal segments of the pipe are respectively P1 and P2, where P1 > P2, then show that vm will be given by: vm = (P1 \u2212 P2) r2/4\u03b7L where r is the radius of the cylinder. (35) Blood of viscosity \u03b7 = 4 \u00d7 10\u22123 PI is passing through a capillary of length L = 1 mm and radius r = 2 \u03bcm. If the speed of this blood as it travels through the center of this capillary is found to be vm = 0.66 mm/s, calculate the blood pressure (in pascal and mm Hg) using the result from the previous exercise" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000900_amr.562-564.654-Figure8-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000900_amr.562-564.654-Figure8-1.png", "caption": "Fig. 8 Ellipsograph mechanism", "texts": [ "4 ,0 0Im z\u03b1 \u03b2 \u03b3= , the common constraint of link group EF is ( )0,0 0gz IIm z\u03b1 \u03b2= , so the virtual constraint is ( ) (2,4 0,0 0 0 I II X gz IIm m z m\u03b1 \u03b2 \u03b1 \u03b2= + ) ( ),0 0 0,0 0 3z z\u03b1 \u03b2= = , the mobility of loop II is 2 6 3 1IIF = \u2212 + = \u2212 . Link group EF is a generalized group with one constraint, and the mobility of the mechanism is Example 5. Fig.7 shows a Watt\u2019s six bar mechanism, In loop I, 3X I Im m= = . Links 3 and 4 which compose the virtual pair are adjacent, so 3X II IIm m= = and 7 3 2 [(3 3) (3 3)] 1F = \u2212 \u00d7 + \u2212 + \u2212 = . Fig. 7 Watt\u2019s six bar mechanism Example 6. Fig.8 shows a Ellipsograph mechanism. Links OA=AB=AD, OC\u22a5OD. In loop I, ( ),0 0 3X I Im m z\u03b1 \u03b2 \u03b3= = . In loop II, point D on link 2 can translate along y-axis and rotate about z-axis, therefore, 4 constraints are added on point D. The constraint of the generalized pair 2,5 IG in point D is 2,5 2,5 ( 0, 0 )D D I Im m x z\u03b1 \u03b2= . Similarly, the constraint of the generalized pair 2,5 IG in point B is 2,5 2,5 ( 0,0 )B B I Im m y z\u03b1 \u03b2= . It needs to be explained emphatically that in most cases, the quantities and characters should be unanimous for motion parameters of different points on two components which composed the generalized pair with relative motion" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001260_s11465-011-0225-z-Figure4-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001260_s11465-011-0225-z-Figure4-1.png", "caption": "Fig. 4 The PRRRP mechanism in an existing 6-DOF erector. The sections marked as joints A1 and B1 denote the two driving hydraulic cylinders; joints A2, B2, and C are three revolute joints; and load M denotes the segment", "texts": [ " Here, the letters P and R denote the prismatic and revolute joints, respectively. Obviously, these two cylinders should move synchronously. Notably, the synchronization of two cylinders is not only implemented by the hydraulic control system, but is also ensured by the mechanical structure, such as link C in Fig. 3, showing an especially rotating angle for the circumferential mechanism not drawn in this figure. 2.2 A PRRRP mechanism in the existing 6-DOF erectors The PRRRP mechanism is adopted in the existing 6-DOF erector to adjust the segment pose (Fig. 4) where the circumferential mechanism and the pose adjustment mechanism are not drawn. There are two degrees of freedom in this radial mechanism. The two hydraulic cylinders can implement synchronous lifting and pose adjustment when the cylinders move synchronously and differentially, respectively. The synchronization of two cylinders depends entirely on the hydraulic control system. In the practical control system, no matter what hydraulic elements are adopted (e.g., the synchronal valves or advanced controlling strategies), the precise synchronization of the two cylinders is difficult to achieve" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000718_iccme.2012.6275659-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000718_iccme.2012.6275659-Figure1-1.png", "caption": "Fig. 1 Schematic of a micropipette in contact with Oocyte membrane.", "texts": [ " In addition, the terms (kZik - PTV1Sik) in lateral equations of motion provide a clear indication of motion-induced stiffness variations that might be affecting the beam stability. The lateral bending stiffness decreases in accelerating axially oscillating rigid base (VI >0) and vice versa. The stability of a beam experiencing motion-induced stiffness variations was investigated in [14]. In next section, the equations of motion are solved using MATLAB\u00ae ode14x extrapolation fixed-step solver based on linearly implicit Euler method [16]. III. MICROPIPETTE TIP FREE MOTTON In this section, the dynamic model of the micropipette, shown in Fig. 1, previously developed with and without embedded mercury is simulated with various boundary conditions. Two distinct simulation cases have been tested: first, with a micropipette tip free in vacuum and second, with the tip in contact with oocyte membrane. The simulation parameters listed in Table I are the same as that presented in [8]. The amplitude and duration of the piezoelectric actuator axial input pulse used to excite the micropipette shoulder (see TABLE I SIMULA nON PARAMETERS [81 Parameter Value Parameter Value Length, L 1.1 mm Medium density 1000 kglm' Inner diameter, d; 8J.1m Mercury density 13543 kglm3 Outer diameter, do 10J.lm Glass density 2290 kglm Young's modulus, E 63.4 GPa Pulse duration 60 J.lS Fig. 1) are 80 nm and 60 IlS, respectively. Note that for the present analysis, it is assumed that the input pulse amplitude is completely transferred up to the micropipette shoulder without attenuation or growth. Furthermore, the micropipette is considered to be perfectly aligned with the piezoelectric actuator. In order to model the micropipette dynamics accurately, four mode shapes are utilized in our simulations. Fig. 3 demonstrates the resultant dynamic responses of the tip in vacuum and without embedded mercury due to the input pulse" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000593_j.proeng.2011.08.169-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000593_j.proeng.2011.08.169-Figure1-1.png", "caption": "Figure 1. An underwater vehicle developed by Shanghai University", "texts": [ " However, the underwater vehicles will be impacted by the inevitable surge, which will result in errors in the attitude control, affecting their normal operation. Many control strategies have been developed for the underwater vehicles [3-7]. Some of these works are summarized in [8]. The work has been shown that the control overshoot must be suppressed in the ROV control system design [9]. In this paper, calculation and simulation of the second order wave are given. Then, anti-disturbance control strategy is put forward which is used in an underwater vehicle developed by Shanghai University, shown in Figure 1. The wave surface which can be described by simple function is defined as regular wave. The regular wave can present the flow and it is also the basic of reach on the irregular wave. The long wave can be described as a composition of so many single regular waves which are independent, have different wavelengths, different amplitudes and stochastic phase, so it can be described as 0 1 ) cos( )ai i i i i t k x t\u03b6 \u03b6 \u03c9 \u03b5 \u221e = = \u2212 +\u2211\uff08 (1) Where\uff0c \u03b6 is the distance that wave surface deviate from the water horizontal plane; 0x is the coordinate of a point on the water horizontal plane; ai\u03b6 \uff0c ik , i\u03c9 , i\u03b5 are amplitude, wave number, frequency, initial phase of the first order harmonics respectively" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002709_smasis2011-5076-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002709_smasis2011-5076-Figure3-1.png", "caption": "FIGURE 3. Experimental configuration used in measuring transverse load\u2013displacement behavior of the VHB while under a biaxial prestretch of 300%", "texts": [ " Both sets of experiments were conducted using a MTS Insight 1 kN load frame and a 5 N load cell. The results of the uniaxial experiment are shown in Figure 2. A stretch rate of 5.64\u00d710\u22125 s\u22121 was used which illustrates minor viscoelasticity at this load rate. The uniaxial experiment is compared to a set of transverse loading experiments where the VHB was biaxially prestretched to 300%. This experimental configuration more closely corresponds to the aerodynamic loading experienced during wind tunnel experiments. A schematic of the loading configuration is illustrated in Figure 3 where the VHB membrane is prestretched and adhered to a rigid ring with an inner diameter of 50.8 mm. The load is applied normal to the biaxial pre-stretched direction using a circular push rod with a diameter of 34.3 mm which was attached to the MTS load cell. The nominal strain induced in the membrane was larger than strains observed during aerodynamic loading, thus providing estimates on bounds for the elastic and hyperelastic behavior. Also note that the displacement rate during this experiment is comparable to the uniaxial experiment; however, the strain rate is inhomogeneous due to the variable contact area of the push-rod" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001153_s1068798x12040120-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001153_s1068798x12040120-Figure2-1.png", "caption": "Fig. 2. Dependence of the clutch torque on the relative rotary angle of the half clutches: hysteresis loop of the torque oscil lations around the rated value (Tr) in a single cycle.", "texts": [ " 4 2012 TEMPERATURE VARIATION OF ELASTIC ELEMENTS IN CLUTCHES ON DAMPING 323 driving and driven half clutches, respectively; d is the damping, which is proportional to the angular velocity, N m s/rad; \u03a91 is the azimuthal frequency of the har monic load at the clutch, rad/s; Ccl is the dynamic rigidity of the clutch in torsion, N m/rad; is the amplitude of the rotary torque; J1 and J2 are the moments of inertia of the driving and driven half clutches, respectively, kg m2. The relative damping coefficient of the elastic clutches is (Fig. 2) [1, 5] where Ad is the energy absorbed in one damping cycle; Ael is the energy consumed in elastic deformation of the elastomer. We know that [5] where \u03c90 = is the eigenfrequency of the system (the resonant frequency with no damping). Then, the power losses due to damping take the form (1) According to the fundamental laws of thermody namics, the heat liberated by the source is where \u03b1 is the thermal conductivity, W/(m2 K); \u0394\u03b8 = \u03b8 \u2013 \u03b80 is the temperature variation, K; \u03b8 is the working temperature of the clutch\u2019s elastic element, K; \u03b80 is the ambient temperature, K; S is the cooled surface area of the elastic element; t is the time" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003001_detc2013-12231-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003001_detc2013-12231-Figure2-1.png", "caption": "Fig. 2 The schematic diagram of the intersected beveloid gear set", "texts": [ " In the mesh model, the coupling effects of the comprehensive assembly errors and the elastic deformation of the tooth surface were considered. Then, loaded tooth contact analysis was performed to investigate the coupling and tooth surface profile modification effects. Finally, a test rig was setup and the contact characteristics are investigated experimentally to verify the computed results. The research hasa significant value for the beveloid gear tooth surface design and optimization for intersected beveloid gears used in marine transmission. The schematic diagram of the intersected beveloid gear transmission is shown in Fig. 2. The position vector of the pinion tooth surface is represented by 1r and the one of gear is represented by 2r . According to the spatial meshing theory, the surface equation of 2r can be represented by 2 21 1 12 0 r M r n v (1) Where 21M is the coordinate transformation matrix, n denotes the mutual normal vector of the engaged tooth surfaces. 12v is the relative velocity. The relative velocity 12( )v is defined in the mutual tangent plane to guarantee the continuous contact condition, which can be represented by 12 0( ) n v (2) The transformation relationship of the coordinate systems of the pinion and gear is shown in Fig" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001281_amm.448-453.3115-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001281_amm.448-453.3115-Figure1-1.png", "caption": "Fig. 1 Finite element model of gear pair in PSD", "texts": [ " All the heat generated by friction is absorbed by the tooth surfaces. Gear parameters. The material of gears is 20CrMnTi, and Table 1 shows its physical parameters. Table 1. Material parameters Elasticity modulus [MPa] Poisson\u2019s ratio Density [kg/m 3 ] Specific heat [J/kgK] Heat conductivity [W/mK] Expansion [10 -5 /K] 210,000 0.3 7800 590 32 1.17 Simulation model of thermal-mechanical coupling. The heat conduction during the gear meshing is a three-dimension problem [4]. The mesh of the gear pair in PSD is obtained using Hypermesh software, as shown in Fig. 1. A thermal-mechanical analysis is conducted during the gear meshing process in ABAQUS. Fig. 2 shows the flowchart for the thermal-mechanical coupling analysis. Boundary conditions. Being convenient to apply the loads and boundaries, two reference points are built, as shown in Fig. 1. The degree of freedoms (DOFs) of reference points about the axis of the planet gear and the half axle gear are released, and all other DOFs are fixed. According to the simulation results under the typical working condition, the rotation speed V=200 rad/s is applied at the reference point RP-1 of the half axle gear, and the resistance torque T=150000 Nmm is applied at the reference point RP-2 of the planet gear. Fig. 3 shows the temperature field distributions on the planet gear and the half axle gear" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000563_detc2011-48891-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000563_detc2011-48891-Figure1-1.png", "caption": "FIGURE 1. TRANSLATIONAL-ROTATIONAL, LUMPEDPARAMETER MODEL OF A PLANETARY GEAR.", "texts": [ " The formulated harmonic balance is verified by finite element analysis and numerical integration. In summary, the major objectives of the present work are to: establish lumped-parameter and finite element models of planetary gears including bearing clearance, tooth separation, and mesh stiffness variation; apply the harmonic balance method with arc-length continuation and stability analysis to obtain the dynamic response; investigate nonlinear behavior, bifurcations, and chaos. The lumped-parameter model of a planetary gear in Figure 1 considers the carrier, ring, sun, and planets as rigid bodies with each having two translational and one rotational degree of freedom. Translational displacements xl , yl (l = c,r,s) are assigned to the carrier, ring, and sun. The radial and tangential displacements of the planets are denoted by \u03bei, \u03b7i (i = 1,2, ...,N) with respect to the carrier fixed reference frame. N is the number of planets. The rotational displacements are uv = rv\u03b8v (v = c,r,s,1, . . . ,N) where \u03b8v is the rotation in radians, and rv is the base circle radius for the sun, ring, and planets and the radius to the planet center for the carrier", " The solution loses stability, or gains it if the multiplier enters the unit circle. 3. Secondary Hopf bifurcation: A pair of complex Floquet multipliers leave the unit circle from complex numbers. A quasi-chaotic regime is generated afterwards. The examined planetary gear detailed in Tables 1 and 2 is a variation of that in a helicopter drive train. This planetary gear includes two pairs of diametrically opposed planets. The planets are mounted at 0, 32\u03c0 63 , \u03c0 , and 95\u03c0 63 measured counter-clockwise with respect to the xc axis in Figure 1. The lowest seven natural frequencies without considering any bearing clearance are translational modes at 1761, 1771, 3282, and 3302Hz; rotational modes at 2181 and 5281Hz; and a planet mode at 3582Hz. The mode types are defined in [34, 46]. Damping in the finite element model includes material damping and bearing damping. Material damping is represented as Cm = \u03b1M+\u03b2K, where \u03b1 = 10sec and \u03b2 = 10\u22128sec\u22121. The damping matrices of the sun and carrier bearings are Cb,s = Cb,c = diag(cr,s,cr,s,c\u03b8 ,s) The damping matrices of the planet 5 Copyright c\u20dd 2011 by ASME Downloaded From: http://proceedings" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003859_9780857094537.13.789-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003859_9780857094537.13.789-Figure2-1.png", "caption": "Figure 2. The free body diagrams for disks 2 and 3 of the four-to include a model of a gearbox inertia idealisation of a drive system", "texts": [ " The four equations of motion of this simple system, assuming viscous damping in the shaft, are 2 1 1 1 1 2 1 1 2 1 2 2 1 2 1 1 2 1 2 2 23 3 3 2 3 4 2 3 4 3 3 23 4 4 2 4 3 2 4 3 4 I c k T I c k T R F I c k T R F I c k T (1) where 23F represents the instantaneous force acting between gears 2 and 3 along the common tangent and where 1 2 3 4, , ,T T T T are the instantaneous torques from external sources being exerted on inertias 1, 2, 3 and 4 respectively. It appears anomalous at first that 23F contributes positively to both gear wheels but inspection of the free body diagrams for disks 2 and 3, shown in Figure 2, reveals that this is correct. There are four rotation co-ordinates in Equation (1) but only three of these are independent because of the constraint between 2 and 3 . We choose 1 2 4, , as independent coordinates, and 3 is eliminated using 3 2 . The forcing term 23F in Equation (1) is unknown, and this is removed by subtracting times the third equation from the second. The resulting three equations in 1 2 4, , are 1 1 1 1 2 1 1 2 1 2 2 2 2 3 2 1 2 1 2 2 4 1 2 1 2 2 4 2 3 4 4 2 4 2 2 4 2 4 I c k T I I c c k k T T I c k T (2) These equations are equivalent to those obtained from a three-inertia model of a simple rotor but comprising only three inertias" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003035_6.2012-1736-Figure7-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003035_6.2012-1736-Figure7-1.png", "caption": "Figure 7. Two Helical mode shapes as example: Left: Axial translation of ISS Middle: No modal deformation (Reference), Right: Rotation of ISS around one of its radial bearings", "texts": [ "2 01 2- 17 36 In helical modes, the mode shapes are predominantly manifested in the helical gear stages of the gearbox and consist of rotations or axial translations of respectively the high speed shaft, intermediate speed shaft and/or low speed shaft. These modes are similar to the modes typically found in helical gear systems. For the investigated gearbox, eigenfrequencies can be categorized into low speed helical stage modes and high speed helical stage modes, according to the helical stage in which they are manifested. Mode shapes involving the ISS are assigned to both categories. Two of these modes are visualized in Figure 7. The global modes, on the other hand, are manifested in both the planetary gear stages and the helical gear stages of the gearbox. Especially with regard to torsional modal displacements these modes are important. For further details on the characteristic behaviour of multimegawatt gearboxes the reader is referred to [22]. In addition to the global modes, which can be determined by means of rigid multibody models also the modes related to the structural components are important. Main structural components are gearbox housing, planet carrier (PC), low speed shaft (LSS), intermediate speed shaft (ISS) and high speed shaft (HSS)" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000478_kem.568.169-Figure5-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000478_kem.568.169-Figure5-1.png", "caption": "Figure 5 presents the series-wound mechanism of planetary transmission with small teeth difference and 2K-H planetary transmission; Figure 6 shows the planetary train that is formed in series by two K-H-V planetary trains.", "texts": [ " Figure 3 illustrates the topological graphs of basic structure of PGTs. It can be seen that sun gears may have fixed joint which is denoted by small triangle or rotary pair which is default with the frame. As the connection between the fundamental units is different, they construct various PGTs. Composition of PGTs with small teeth difference Because there are the explicit requests on input and output about PGTs with small teeth difference, the combination of fundamental units of such PGTs and those of other PGTs are limited, usually in series. Figure 5 shows several kinds of planetary trains with small teeth difference: in Fig.1, W mechanism in 2K-H planetary trains has zero teeth difference, it is composed by two pairs of inner gear pairs, which have the same teeth difference and the same modulus number, but the different teeth number. Here, the planetary gear is two-gear cluster. In Fig .2, K-H-V gear train is the usual planetary train with small teeth difference. There are several kinds of K-H gear train such as the coaxial translation, tricyclic planetary train with small teeth difference, the cycloid pin gear, etc" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002357_978-3-642-23244-2_49-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002357_978-3-642-23244-2_49-Figure1-1.png", "caption": "Fig. 1. Design of preliminary conception", "texts": [ " The role of the wheels can be used not only for increasing efficiency of locomotion, but also for overcoming obstacles [4]. Independent leg articulation [2] supplemented by wheel can be effective solution for experimental robot suited for research and education in mechatronics. The developed novel construction of the hybrid undercarriage described in this paper was designed to fulfill following requirements: speed 8 km/h, 2.2 m/s; climbing ability 30\u00b0; step obstacle - ability to go upstairs (maximum height of a step is 10 cm). After many preliminary conceptual variants (Fig. 1) the resulting robot equipped with seven servo drives and four independent wheel motors was built. During the development of several variants of the topology and geometry of the robot, the multi body system simulation was extensively used. The complete model of robot, servodrives and environment was built in Matlab/SimMechanics [7] environment. Fig. 2 shows the example of the kinematical analysis of one free leg movement possibilities. Three rigid legs are fit to the ground by spherical kinematic links (thus cannot slip on the ground)" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002278_978-1-4471-4628-5_3-Figure3.65-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002278_978-1-4471-4628-5_3-Figure3.65-1.png", "caption": "Fig. 3.65 Bond graph model of solenoid", "texts": [ " Solenoids A solenoid transforms an electrical signal into mechanical movement. A typical solenoid consists of a coil of wire wrapped around a fixed iron core. There is a movable iron core named as armature and a spring and damper as shown in Fig. 3.64. An increase in voltage (Vin), causes the current to rise in the coil, thus increasing the core flux (\u03c6). The increasing flux generates a magnetic force in the air gap which pulls the armature closer to the static housing core. The bond graph model of solenoid is shown in Fig. 3.65. The bond graph model can be divided into electrical, electromagnetic, and magneto-mechanical parts. The electrical model accounts for the change of current in the coil. The effective voltage across the coil (effort in bond 3 in bond graph shown in Fig. 3.65) is given by e = Vin \u2212 R1i (3.50) Here, Vin is the voltage input from the driver, R1 is the resistance in the coil, and i is the current. The electromagnetic model describes the change of magnetic flux in the moving and stationary iron core, and air gaps. We can apply Faraday\u2019s law to the coil, in order to model this part. We know from Faraday\u2019s law that e = \u03bb\u0307 = N \u03c6\u0307 (3.51) M = Ni (3.52) where \u03bb is the flux linkage, N is the number of turns in coil, \u03c6 is the flux, and M is magnetomotive force. Equations 3.51 and 3.52 are modeled using a gyrator element with gyrator modulus N as shown in Fig. 3.65. The effort which establishes the flux in the core is the magnetomotive force which can be further written as M = Mg + Mc (3.53) where Mg and Mc are the magnetomotive forces in the gap and core, respectively. The 1 junction in the model indicates that the flow (flux) in gap and core are same. In the iron core the magnetomotive force (Mc) and flux are related by a nonlinear relation which is a function of material properties and geometry. To characterize a core material, usually the B-H curve is used where B is the magnetic flux density (weber/m2), H is the field strength (Ampere-Turn/m), and Mc = f (\u03c6) (3" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001051_j.piutam.2011.04.017-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001051_j.piutam.2011.04.017-Figure2-1.png", "caption": "Fig. 2. The model used in the simulation. The legs were massless and telescoping to permit toe clearance.", "texts": [ "; (2) constraints on the desired motion, for example, prescribed initial and final conditions, step length, or maximum speed; and (3) constraints that set limits on physical structure, energetics, and Newton's laws that form constraints that specify a physically valid solution [3]. Taking these constraints into account, the generalized positions, velocities q(t), and controls u(t) are determined that minimize an objective function min f(q,u); with the constraints h(q,u) as a nonlinear programming problem (NLP). Such an approach was used in the present study to simulate a non-periodic gait with the objective of traversing a specified distance in minimum time. The model used in the simulation (Figure 2) was composed of the following elements: a point mass at the hip; damped, compliant, and massless legs with telescoping axial actuators; and point feet (each with a mass that was small compared to point mass at hip). Telescoping axial actuators apply compressive forces (P1 and P2) necessary for foot clearance and the torque (T) around the hip is essential for swing of legs. The ground was modeled as a viscoelastic medium. Using spacetime constraints, the optimal control problem of traversing a specified distance in minimum time from rest was transformed to a NLP problem", " Similar differences in toe length were found in the present study, with sprinters having toes that were 5%-10% longer than height-matched non-sprinters. Our successful simulation of non-periodic steps at the start of a sprint will form the basis for future simulations using more complex models that will permit analysis of the advantages conveyed to sprinters by the measured differences in foot and ankle anatomy. We are presently at work adding feet, ankles, and ankle muscles to the model shown in Figure 2, and we plan to use the resulting simulation to perform sensitivity studies to establish the dependence of sprint performance upon variation in pfMA and toe length. Establishing links between foot and ankle structure and function is important not only for understanding the determinants of sprint performance but for identifying naturally-occurring structural variations that may contribute to movement abnormalities or age-related loss of mobility. The mechanisms by which limb and joint structure affect performance in such cases is likely to be different from that which determines sprint performance, and we anticipate that simulations will be similarly valuable in characterizing structure-function relationships in these other populations" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000023_icecc.2011.6067768-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000023_icecc.2011.6067768-Figure1-1.png", "caption": "Figure 1. Figure1. Definition of coordinate system", "texts": [ "Another coordinate system,with the same origin as the earth-fixed system and the axes parallel to the main axes of the vessel,is used for computing the forces acting on the vessel. Let [ , , ]Tx y\u03b7 \u03c8= be the vector denoting the position and heading of the vessel relative to an earth-fixed frame, and let [ , , ]Tu v r\u03bd = represent the velocities of the vessel decomposed in a body-fixed frame.For both systems, the z-axis is defined along the vertical,with the x-y plane in the calm water surface. The coordinate system definitions are shown in Fig. 1.In this figure,OEXEYE is the earth-fixed frame,OXY is the body-fixed frame,and O is the center of gravity of the vessel. the equations of motion in surge,sway,and yaw can be written[7]: ( ) ( )x y HX WaX WiX CX TXm m u m m vr F F F F F+ \u2212 + = + + + + (23) ( ) ( )y x HY WaY WiY CY TYm m v m m ur F F F F F+ + + = + + + + (24) ( )Z Z H Wa Wi C TI J r N N N N N+ = + + + + (25) Where the following definitions are used: u , v is velocity in X and Y direction; \u03c8 , r is heading and heading rate; m is the vessel mass; TXF , TYF , TN is the control input vector of force and moment provided by the actuator system; xm , ym , ZJ are inertial coefficients which are assumed to be constants; ZI is the moment of inertia about the body-fixed Z-axis; This paper studies a new type of dynamic positioning ships" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003409_cp.2013.2376-Figure5-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003409_cp.2013.2376-Figure5-1.png", "caption": "Figure 5. Antenna pattern rotating when no intentional interference", "texts": [ " Therefore, the SNR and hence the reliability of receiving VLF messages can also be improved significantly. The method does not need to change the mobile platform heading to adjust the directivity of magnetic antenna. This is beneficial to the mobility of the mobile platform. Interference When there is no intentional interference, the jamming source mainly is environment interference, natural noise, etc. These jamming signals may be considered as uniform distribution in all directions. When receiving messages, the rotation of the antenna pattern is shown in Fig. 5, namely the peak of the antenna pattern points to the direction of messages signal. The rotation scanning controller in Fig. 3 scans angle . When \u03b2 \u03d5= (or 180\u03b2 \u03d5= \u00b1 ), the peak of the antenna pattern points at the direction of messages signal whose azimuth is \u03d5 . The antenna pattern rotation system outputs a maximum value of messages signal, and the receiver outputs the received message with maximum SNR. Angle may also be input manually when the message signal direction is known. By means of adjusting the antenna pattern, the peak of the antenna pattern can be pointed to the direction of messages signal for receiving VLF messages, so that the antenna pattern rotation system outputs the maximum SNR, environment interference and the natural noise are suppressed significantly" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000347_icdma.2011.75-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000347_icdma.2011.75-Figure2-1.png", "caption": "Figure 2. Geometric scheme of TIVT[10]", "texts": [ " structure of TIVT[10] Force on the roller by hydraulic oil Fk1 is set constant to be 11309.7N, inner radius of roller rk is set to be 20mm. Then hydraulic oil pressure Pk is 9N/mm2 by calculation on the following equation: 2 1 k k k r FP \u00d7 = \u03c0 (N/mm2) (1) To calculate the maximum contact force generated by hydraulic oil pressure, the maximum angle Sk between direction of Fk1 and normal direction of roller at contact point between roller and input disc is to be calculated. So geometric scheme of [10] is needed, which is shown in figure 2. 978-0-7695-4455-7/11 $26.00 \u00a9 2011 IEEE DOI 10.1109/ICDMA.2011.75 280 As shown in figure 2, set input disc radius as R1, output disc radius as R2, tilting angle of roller as , included angle of input disc axis and line connecting left roller sphere head center and input disc sphere surface center as input angle 1, included angle of output disc axis and line connecting right roller sphere head center and output disc sphere surface center as output angle 2, left roller sphere radius as r1, and right roller sphere radius as r2. When roller axis is parallel to input disc axis, 1 becomes 1, 2 becomes 2, H is distance of left roller sphere head center from axis of input disc, H is distance of right roller sphere head center from axis of input disc, L is distance of left roller sphere head center from symmetry plane of controller, L is distance of right roller sphere head center from symmetry plane of controller. Then from figure 2, three equations can be obtained: ( ) ( ) ( ) ( ) ( ) ( ) ( ) ( )2222 2222 1111 1111 coscos sinsin coscos sinsin \u03d5\u03b1 \u03b1\u03d5 \u03d5\u03b1 \u03d5\u03b1\u03b8 \u2212\u2022\u2212+ \u2212\u2022\u2212= \u2212\u2022\u2212+\u2032 \u2212\u2022\u2212= rRL rR rRL rRtg (2) 11 1sin rR H \u2212 \u2032 =\u03d5 (3) 22 2sin rR H \u2212 =\u03d5 (4) Let H'=H, L'=L, R1=R2=R and r1=r2=r, where R means disc sphere surface radius and r means roller sphere head radius. Then: ( ) ( ) ( ) ( ) ( ) ( ) ( ) ( )22 22 11 11 coscos sinsin coscos sinsin \u03d5\u03b1 \u03b1\u03d5 \u03d5\u03b1 \u03d5\u03b1\u03b8 \u2212\u2022\u2212+ \u2212\u2022\u2212= \u2212\u2022\u2212+ \u2212\u2022\u2212= rRL rR rRL rRtg (5) 21 sinsin \u03d5\u03d5 = \u2212 = rR H (6) The corresponding equation for (5) in [10] is wrong" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001430_s12206-012-1266-x-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001430_s12206-012-1266-x-Figure1-1.png", "caption": "Fig. 1. Damage on the lock nut.", "texts": [ " Kasei [2], Friede and Lange [3] carried out a bolt loosening test of bolted lap joint by applying a repeated shear load to induce a slippage between the fastened plates. Bhattacharya, Sen and Das [4] tested the anti-loosening ability of threaded fasteners by applying vibration perpendicular to the bolt axis. In this paper, the cause of the lock nut loosening is investigated using three-dimensional finite element analysis of the drive axle assembly, macroscopic observation of the failed lock nut surface and experimental simulation on the test bed. The top surface of the lock nut was originally one flat surface, as shown in Fig. 1(a). It was in contact with the planet carrier face axially holding the wheel shaft with the tightening force. The failed lock nut was as shown in Fig. 1(b). The annular region was lowered. The lowered region\u2019s radial thickness is 3.5 mm. This is just the radial thickness at the region with overlapping the planet carrier face. The lock nut appeared as if it were being collapsed by the excessive compressive contact stress. The contact stress was to be checked. CATIA structural analysis module was used as a finite element analysis tool. Some parts irrelevant to the problem were deleted from the assembly. Only six parts in Fig. 2 were modeled in the structural analysis" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000881_apec.2013.6520300-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000881_apec.2013.6520300-Figure1-1.png", "caption": "Figure 1. Structure of the SMIIR.", "texts": [ " The voltage can not only transfer power to the rotorside inverter but also cause current response which contains rotor position information. And an initial +d-axis detection algorithm without using magnetic saturation is also devised. The feasibility of the proposed method is verified by experimental test on the prototype SMIIR with associated control system. II. OPERATION PRINCIPLE OF SMIIR The SMIIR is based on an wound rotor synchronous motor, but an inverter is integrated inside the rotor, as shown in Fig. 1. The rotor-side inverter consists of DC link capacitor and switching devices and its own controller. And the rotor-side inverter rotates synchronously as the rotor rotates and is electrically isolated from the outside of the rotor. Therefore, there is no brush and slip-ring. This is the main difference between the SMIIR and the conventional wound rotor synchronous machine. The equivalent circuit of the SMIIR in the d-q rotor reference frame is shown in Fig. 2. The voltage source dqsv and dqrv denote the output voltages of the stator-side and the rotor-side inverters respectively" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002278_978-1-4471-4628-5_3-Figure3.94-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002278_978-1-4471-4628-5_3-Figure3.94-1.png", "caption": "Fig. 3.94 Schematic diagram of use of tachogenerator", "texts": [ " This type of sensors are used as bumper sensor in walking robots to detect the presence of an obstacle, so that the robot can change its path. Incremental encoders can be used to measure the angular velocity of the shaft. Here we have to find the number of pulses produced per second in order to determine the angular velocity of the shaft. Tachogenerator is a sensor to measure the angular or linear velocity (through a pinion). In principle it is similar to AC generator. It consists of a rotor which is mounted on a rotating shaft as shown in Fig. 3.94. The rotor rotates in a magnetic field produced by either a permanent magnet or an electromagnet. When rotor rotates in the magnetic field an alternating emf is produced in the rotor. The amplitude or the frequency of the alternating emf produced can be a measure of angular velocity of the shaft. The alternating emf may be rectified to give DC output voltage. Figure 3.95 shows the bond graph model of the tachogenerator where the R element between the two transformers models the friction between the two wheels" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000609_amm.86.653-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000609_amm.86.653-Figure1-1.png", "caption": "Fig. 1 The schematic diagram of planetary gear Fig.2 Dynamics model of planetary gear system", "texts": [ " The time-series of wind speed is obtained from the LS-SVM model, then the time-series of external load of the planetary gear system of wind turbine is obtained combine with the aerodynamics theory and used as the external excitation for simulating dynamic performance of the planetary gear system of wind turbine. The planetary gear used in multibrid technology wind turbine (MTWT) with 1.5MW power is taken as an object for study, where the planet carrier serves as input, the sun serves as output, the ring is rigidly fixed to the gearbox, the schematic diagram of the planetary gear system is shown in Fig.1. Bearing Force and Time-Varying Stiffness. Fig.3 shows the parameters and schematic diagram of the rolling element bearing. The bearing is modeled as a two-DOF system, while ignoring mass and inertia of the rolling elements. The two orthogonal DOFs are related to the inner race (rotor). Contact forces are summed over each of the rolling elements to get the overall forces on the shaft in the load direction and at right angles to it. The total nonlinear bearing force in the x and y directions for a ball bearing with b n balls can be obtained based on the Hertzian contact relationship [9,10]" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001788_pacc.2011.5979014-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001788_pacc.2011.5979014-Figure2-1.png", "caption": "Figure 2. Generalized Coordinates For The AltAz Telescope Mechanics", "texts": [ " The basic pointing mechanism consists of a rotating base with a rotating sight mounted on the base, which gives the telescope mechanics two degrees of freedom and allows it to aim at object in three dimensional space. Two actuators are needed to rotate the base and sight. Angle sensors are most commonly used as feedback elements. Servo control is applied on the base and sight actuators in order to direct the telescope to the object by giving it reference commands. The reference commands in turn are normally given either as static coordinates, or outputs from target tracking mechanism. Figure 2 and 3 shows the generalized coordinates used to calculate the equations of motion for the dynamic load with Lagrangian mechanics. To set up the Lagrangian , the kinetic and potential energies for the system must be expressed as a function of generalized coordinates. The potential energy of the system is expressed as :- Ep = \u222b g.r.sin\u03b1.dm = \u03c1.g.sin\u03b1. \u222b l 0 r.dr = \u03c1.g.sin\u03b1.[r2/2]l0 = \u03c1.g.sin\u03b1.[l2/2] = m.g.[l/2].sin\u03b1 The kinetic energy of the system is expressed as:EK = EK,Base +EK,Sight 978-1-61284-764-1/11/$26" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001433_ispcc.2013.6663464-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001433_ispcc.2013.6663464-Figure1-1.png", "caption": "Fig. 1. Sarika 1 Micro Air Vehicle", "texts": [ " Section 2 gives the model description and modeling of Sarika-1 for its longitudinal dynamics. Section 3 gives the controller design specifications and principles of fixed order robust H2 controller. Section 4 briefly explains the H2 performance robustness analysis using LMI techniques. The results and analysis of offline simulation, real time HILS experiment using the designed controller and its robust performance analysis is given in section 5. Finally conclusions are drawn in section 6 II. MAV DESCRIPTION AND ITS MODELLING MAV named Sarika-1 is shown in Fig. 1, which is a remotely piloted small flying vehicle of about 1.28 m span and 0.8 m length and weights around 1.75 kg at takeoff. It has a rectangular wing of planform area of 0.2688 m2 and a constant area square section fuselage of width 0.06 m. The control surfaces are elevators, ailerons and rudder. The power plant is a 4 cc propeller engine (OSMAX -25LA), which uses methanol plus castor oil as fuel, with 10 to 15 % nitro-methane to boost the engine power. Sarika-1 has a provision to carry video camera and sensor payloads" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003483_allerton.2013.6736717-Figure4-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003483_allerton.2013.6736717-Figure4-1.png", "caption": "Fig. 4: The control line uniquely determines a section of an optimal trajectory for a Dubins car.", "texts": [ " In this fashion, the trajectory \u2018bounces\u2019 along the control line until the goal is reached. There are three classes of control line trajectories: generic, singular, and tacking. 1) Generic trajectories: Generic trajectories are those for which the initial placement of the rigid body with respect to the control line completely determines the structure of the optimal trajectories. Along a generic trajectory, the set of times for which multiple controls maximize the Hamiltonian is of measure zero; these switching times correspond to well-defined switches between controls. Figure 4 shows an example of a generic trajectory for the Dubins car. 2) Singular trajectories: Translation of the rigid body parallel to the control line does not change the set of currently maximizing controls, since the maximizing controls depend on the distance and angle of the body relative to the control line. Therefore, such translation may occur indefinitely, and be followed by any of several other maximizing controls. Therefore, in this singular case, no single trajectory structure is determined by the configuration of the body with respect to the control line; there may be many trajectories consistent with the configuration, all of which satisfy the Maximum Principle" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001506_iros.2013.6696627-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001506_iros.2013.6696627-Figure2-1.png", "caption": "Figure 2. Inner structure of tissue-engineering-based mechanoreceptors", "texts": [ " Both types consist of an outer layer and an inner structure including mechanoreceptors which is responsible for sensing and processing. The outer layer is an elastic artificial skin-like thin film which is responsible for absorbing interaction. Its 978-1-4673-6358-7/13/$31.00 \u00a92013 IEEE 2030 outer and inner surfaces have the ridged texture like human finger-print. The inner structure consists of an artificially tissue-engineered axons, mechanoreceptor capsules, guide tubes, and growth solution, as shown in Fig.2. The guide tubes work as tunnels leading the axon terminals from parent fibers to prepared mechanoreceptor capsules. The growth solution provides inner pressure and keeps the mechanoreceptors living since organic components may decay easily. In order to supply growth solution frequently, we are thinking about using a two-way exchange system like human circulatory system, which is not illustrated here. This type of system may decrease the mobility but increase the stability of the tactile sensor" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003029_dscc2013-3998-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003029_dscc2013-3998-Figure2-1.png", "caption": "Figure 2. MODEL OF HELICOPTER TESTBED", "texts": [ " The extrudedaluminum guide rails keep the helicopter 84 inches (3-times the main rotor diameter of 28 inches) above the floor to avoid aerodynamic phenomena such as ground effect. There is also adequate space above the helicopter to ensure that no unusual aerodynamic conditions are introduced. To minimize vertical force on the guide rails, the main rotor thrust force must be large enough to support the helicopter, pivoting base, and suspended load. The procedure used to determine this thrust force as a function of command to the helicopter is given in [1]. Figure 2 shows a planar sketch of the helicopter. This model is similar to previous planar models of externally-loaded helicopters near hover [3, 4]. The helicopter is pinned through its center of gravity to a horizontally-sliding cart. The helicopter\u2019s horizontal location is x(t), and its pitch angle is \u03b8(t). The suspended load has swing angle \u03c6(t) relative to vertical. A thrust force F(t) is produced by the main rotor, and the angle of the thrust vector relative to the helicopter body is \u03b1(t). To capture the effect of main rotor stiffness, a torsional spring with stiffness k is attached from the helicopter to a massless rod that coincides with the thrust vector [4]. This torque is produced by the main rotor disk (not pictured in Figure 2). It is assumed that the suspension cable is inextensible, and that aerodynamic effects on the load are negligible. The helicopter has mass M and rotational inertia I, and the load has mass m. There are dampers with coefficients c and b on the helicopter position x(t) and helicopter pitch angle \u03b8(t), respectively. In helicopterbody-fixed coordinates, the load suspension point is at distance dL below the helicopter\u2019s center of gravity, and the main rotor thrust force is applied at distance dF above the helicopter\u2019s center of gravity. Two important dynamic elements are not pictured in Figure 2 \u2013 horizontal force on the helicopter caused by 2 Copyright \u00a9 2013 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 01/29/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use its pitch rate is given by f \u03b8\u0307 , and torque on the helicopter caused by its horizontal velocity is given by \u03c4x\u0307 [4]. A commercial multi-body dynamics software package [5] was used to obtain the following equations of motion: M+m dLmcos(\u03b8) Lmcos(\u03c6) mcos(\u03c6) dLmcos(\u03c6\u2212\u03b8) Lm dLmcos(\u03b8) I +md2 L dLLmcos(\u03c6\u2212\u03b8) x\u0308 \u03b8\u0308 \u03c6\u0308 = \u2212x\u0307c+ \u03b8\u03072m[dLsin(\u03b8)+Lsin(\u03c6)] +\u03b8\u0307 f \u03b8\u0307 \u2212T sin(\u03b1+\u03b8) \u2212\u03b8\u03072dLmsin(\u03c6\u2212\u03b8)\u2212gmsin(\u03c6) x\u0307\u03c4x\u0307 + \u03b8\u03072dLLmsin(\u03c6\u2212\u03b8)\u2212 \u03b8\u0307b \u2212dLgmsin(\u03b8)+ k\u03b1+dFT sin(\u03b1) (1) Several parameters of the helicopter-load system were directly measured" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001629_amr.383-390.2963-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001629_amr.383-390.2963-Figure1-1.png", "caption": "Fig. 1. Field-circuit coupled model of the motor", "texts": [ " Definite solution of transient electromagnetic field can be expressed as in 1 2 1 2 1 1 : : 0 1 1 : Z S A A dA J x x y y dt A A A J n n \u03c3 \u00b5 \u00b5 \u00b5 \u00b5 \u2202 \u2202 \u2202 \u2202 \u2126 + = \u2212 \u2212 \u2202 \u2202 \u2202 \u2202 \u0393 = \u2202 \u2202 \u0393 \u2212 = \u2202 \u2202 (3) Where \u2126 is solving area, 1\u0393 is the boundary between stator excircle and rotor inner circle, 2\u0393 is the PM boundary, SJ is the equivalent surface current density of the PM boundary, A is magnetic vetor potential, \u00b5 is permeability, ZJ is the additional axial current density, dA dt\u03c3\u2212 is the eddy current density. The finite element method is adopted to solve in this paper. To ensure the computational accuracy, solving area was divided into lots of triangular elements. Fig. 1 shows the field-circuit coupled model of the motor. Results and analysis. The end effect of the motor is expressed as constant leakage inductance in the winding circuit equation. The calculation results which are obtained by the formula method are that the end resistance is 0.403295\u2126 and the end inductance is 0.049135H. The main parameters of the motor are listed as following. The rated power is 5kW, the rated voltage is 220V, the number of the stator slots is 48, the number of the poles is 10, and the length of the stator core is 225mm" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002278_978-1-4471-4628-5_3-Figure3.9-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002278_978-1-4471-4628-5_3-Figure3.9-1.png", "caption": "Fig. 3.9 Bond graph model of cam drive with flow input and no follower offset", "texts": [ "8) where h is the maximum follower travel, \u03b2 is the angle associated with motion event, 0 < \u03b8 < \u03b2/2 is the rise duration, \u03b2/2 < \u03b8 < \u03b2 is the fall duration, and the rest is the dwell duration. The velocity and acceleration profiles are derived from the displacement profile through successive derivatives and by assuming unit angular speed. If the follower is constrained to move vertically and the follower offset is zero as shown in the figure, then the follower velocity is given by dy dt = dy d\u03b8 . d\u03b8 dt = \u03a8 (\u03b8) \u03b8\u0307 (3.9) where \u03a8 (\u03b8) is purely a function of the cam\u2019s angular rotation from a reference position. Figure 3.9 shows the bond graph model of the cam follower (knife-edge follower) system. Here the valve jump is modeled by contact stiffness Kc and contact damping Rc. If there is no contact then the stiffness and damping forces are made zero. \u2022 Model with effort input: Figure 3.10 shows the general schematic diagram of a cam follower system with a flat-faced follower. Here, the cam is driven through a torque input and its angular Fig. 3.10 Schematic diagram of cam drive with torque input ,C CK R y inT \u03b8 ( )r \u03b8 RiseFall Dwell Follower m 1 1,P A 2 2,P A Mg 'R Friction xF yF speed is variable" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001454_opl.2012.468-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001454_opl.2012.468-Figure1-1.png", "caption": "Figure 1. Cut away profile of the GRIME.", "texts": [ " Thus, the intent of this investigation was three fold: to exploit the geometric and electrical properties of graphene to fabricate a robust micro electrode whose dimensions could be characterised and controlled; to more efficiently exploit the greater electrochemical activity of graphene edge sites (as opposed to basal plane) in the electrode design; and to use that electrode to study the intrinsic electrochemical response of graphene by use of a well known redox probe. To these ends we fabricated the first Graphene RIng Micro Electrodes (GRIMEs), Figure 1, by dip coating fibre optics with subsequently reduced graphite oxide (GO). The graphene used in this study was fabricated by chemical delamination of graphite by its oxidation to GO. The GO generation was achieved by use of a solution phase, oxidative method [12], modified from the original Hummers method [13]; a synthetic route that incorporates oxygen containing functionalities onto both the basal plane and edges of the component carbon layer structure of the graphite starting material. Briefly this method proceeds as follows: 5 g of carbon powder (<20 \u03bcm dia" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002278_978-1-4471-4628-5_3-Figure3.12-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002278_978-1-4471-4628-5_3-Figure3.12-1.png", "caption": "Fig. 3.12 Schematic diagram of gear drive with resistive load", "texts": [ " This type of gearing is used to transform either linear motion to rotary motion or rotary motion to linear motion. When two gears are in mesh, the larger wheel is called gear, while smaller wheel is called pinion. Any combination of gear wheels by means of which motion is transmitted from one shaft to another shaft is called gear train. A single gear train may include any or all forms of gear wheels, such as spur, bevel, helical, spiral, etc. Let us consider a gear drive driving a resistive load as shown in Fig. 3.12. Figure 3.13 shows the bond graph of the system where Ip, Ig , and RL represent the pinion rotary inertia, gear rotary inertia, and load resistance, respectively. Here it is to be noted that backlash and teeth flexibility has been modeled by C f and R f elements. If meshing stiffness, friction, and backlash are neglected then the bond graph can be reduced as shown in Fig. 3.14. In this Fig. 3.14 we observe that the I element corresponding to pinion is differentially causalled, showing its speed dependence on gear speed" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003418_ica.2011.6130140-FigureI-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003418_ica.2011.6130140-FigureI-1.png", "caption": "Fig. I. Model of the free-joint manipulator", "texts": [ " (a) 6,,(j vs. initial angle of \u00a2 (b) final angle vs. initial angle of \u00a2 Finally, the performance of the control method was ex amined based on the database of information vectors. The performance was good when the information vectors were distributed throughout the range of the target angles. The worst performance was obtained when the distribution was either too thin or too wide. The performance is expected to be improved by adjusting the input range according to the dynamics of the manipulator. Fig. I O. Distribution of data for (j and \u00a2 with parameter set A 500' \ufffd\ufffd--\ufffd --------\" 200r----------------, -2$00=-\ufffd-1\ufffd0 \ufffd\ufffdO\ufffd\ufffd100=-\ufffd20 initial angle of Ij)[dcgl [I] M. Kishi, K. Kimura, J. Ota and S. Yamamoto, \" Shrinkage prediction of a steel production via model-on-demand \" , Proceedings of 11th IFAC Symposium on automation in Mining, Mineral and Metal processing , 2004 [2] Ken-ichiro Fukuda, Shun Ushida and Koichiro Deguchi. Just-In-Time Control of Image-Based Inverted Pendulum Systems with a Time-Delay, SICE-ICASE International Joint Conference 2006, pp" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000666_amm.229-231.723-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000666_amm.229-231.723-Figure1-1.png", "caption": "Fig. 1 - Features of the deformation of two cylindrical bodies: a) - with the same elastic properties of bodies, b) - with an elastic-plastic deformation of solid parts deforming element", "texts": [ " However, there are no evidence of this in any literature. There is a fundamental difference between features contact while squeezing the same on the mechanical properties of cylindrical bodies and the case where one body is elastic and more pliable. Difference between a static element in the implementation of the deforming surface of another body in the elastic deformation of bodies with similar mechanical properties and elasto-plastic deformation at its rolling on the surface of the elastic-plastic lies in the fact that in the first case (Fig.1a) contacting a flat figure, as a body are deformed equally and have almost identical mechanical properties. In the second case (Fig.1b) deforming member having a greater hardness, introduced in part to a certain depth, a value which is determined by the applied force, the size of the deforming elements and mechanical properties of the deformable surface. [1] All rights reserved. No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of Trans Tech Publications, www.ttp.net. (ID: 142.103.160.110, University of British Columbia, Kelowna, Canada-07/07/15,18:37:45) Deformation of the deforming calculate the element when it is squeezed between two opposing forces" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002117_icef.2012.6310331-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002117_icef.2012.6310331-Figure3-1.png", "caption": "Figure 3. The result of magnetic flux density in the end winding", "texts": [ " In the mechanical analysis, the equations of end winding components can be simplified as: 0, 1 ( ), ,2 1 fij j i u uij i j j i ij ij ijkkE E \u03c3 \u03b5 \u03bd \u03bd \u03b5 \u03c3 \u03c3 \u03b4 + = = + + = \u2212 \u23a7 \u23aa \u23aa\u23aa \u23a8 \u23aa \u23aa \u23aa\u23a9 (4) where \u03c3ij,j is the stress; fi is the body force; \u03b5ij is the strain; u is the displacement; i, j, k = 1,2,3; E is the Young\u2019s modulus; and \u03bd is the Poisson\u2019s ratio. IV. ANSLYSIS RESULTS OF ELECTRMAGNETIC MODEL Prior to analysis of electromagnetic force on end windings, the magnetic flux density distribution should ensure proper. In Fig. 3 shows the analysis result of magnetic flux density in the end winding. It shows a periodic distribution around the circumference, and high value appears in the involute part of the end winding. One phase belt\u2019s local electromagnetic force density was analyzed for the other phase would have similar situations. Three groups of values were have been shown in figure 4- 6.They are /,F Vi r i, - and \u02c6 /, ~F Vi r i, , /, ,F Vi i\u03d5 - and \u02c6 /, ~F Vi r i, /,F Vi z i, - and \u02c6 /, ~F Vi z i, , respectively. In Fig" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000433_12.2003376-Figure5-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000433_12.2003376-Figure5-1.png", "caption": "Figure 5. Small EMPL mounted on the wooden vest insert; projectile is visible inside the EMPL", "texts": [ " In an effort to condense the electromagnetic field during firing, the coil with the 16 AWG wire was also shielded with steel pipe around the perimeter of the coil as can be seen in Figure 4. The projectiles tested were punched out of 19 Gauge sheet steel to a diameter of 15/16\u201d to freely pass through the 1\u201d opening in the plastic coil housing. The projectile used throughout the EMPL test plan weighed 6.3 grams. The projectiles were roughly the size of a U.S. quarter. The coil assembly was mounted onto the plywood inserts in the vest as seen in Figure 5. For the preliminary user study, after the EMPL test plans were carried out, a simple projectile return system was created to make the tactor more repeatable. The return system was created by drilling a small hole through the middle of the projectile and by using a cotter pin to attach a one-inch long spring. The other end of the spring was attached to a thin sheet of plastic, which was fixed to the outside of the plywood insert. Unenergized and unsprung, the projectile sat flush with the closest edge of the coil" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001956_opl.2011.1003-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001956_opl.2011.1003-Figure3-1.png", "caption": "Figure 3: Magnetic assembly control on a planar substrate : i) random magnetic orientation ii) platinum evaporation iii) magnetic alignment (iv-vi) self assembled structures: iv) linear translator resulting from magnetisation as shown in iii v) sideways translator from magnetisation perpendicular to magnetic cap vi) rotor from mixing swimmers magnetised in both orientations.", "texts": [ " This finding also illustrates a key mechanistic difference between these non-conducting particles, propelled away from the catalyst site, and conductive bi-metallic nano-rod translators, which travel towards the catalytically active end.[3] While it appears that the assembly process occurs at random to produce a wide range of relative configurations and trajectories, in some cases it may be desirable to encourage specific configurations, and trajectories to form. An approach to achieve this is shown in Figure 3. If the individual swimming components contain a ferromagnetic core, it is possible to magnetize them to align the poles at a specific orientation relative to the thrust generating platinum cap. If assembly is then allowed to proceed, both the ability to constrain the orientation of the individual components by an external field, and the interaction between the poles on attaching swimmers will direct the preferred configuration. We have physically realized this scheme to produce doublet swimmers with the favored configuration shown in Figure 3(v). Figure 4 shows the trajectory that this swimmer exhibits, with the propulsion direction being perpendicular to the long axis as expected. In addition, the dipole alignment makes it possible to steer the swimmer using an external field. In this case a Ushape trajectory is achieved by reversing the external field direction during swimming. It should be noted that the field is homogenous and only controls orientation, and does not produce the translation, as confirmed by a control experiment in which swimming proceeds with the same velocity in the absence of the applied field" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000873_j.proeng.2013.07.025-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000873_j.proeng.2013.07.025-Figure1-1.png", "caption": "Fig. 1. A device for simulating under-g conditions, consisting of a custom made harness, supported by ropes which are attached to four pneumatic cylinders. Two precision regulating valves (one valve for two cylinders) ensure a constant pressure of the pneumatic cylinders. (a) Schematic figure of the unloading device, where a ring replaces the costume made harness and the instrumented force treadmill is replaced by a box. (b) Picture of the constructed unloading device with a subject integrated.", "texts": [], "surrounding_texts": [ "Eleven healthy male subjects (age 26 \u00b1 3 yrs; height 180 \u00b1 7 cm; weight 76 \u00b1 8 kg; mean \u00b1 SD) participated in this study. All subjects were free of pain and injury. To calculate the individual unloading situation a gradual load reduction was carried out, whereas the pressure was increased in 20 kPa steps, from 0 Pa up to 500 kPa, both in a static and in a dynamic (running at 3.0 m/s) condition. The maximum pressure level was individually adapted to the unloading situation. Maximum pressures ranged from 300 kPa up to 500 kPa. For every pressure level in the dynamic condition (3 m/s), ten stance phases of the right leg were analyzed. To control the load reduction, vertical GRF were measured by means of the instrumented force treadmill. In the dynamic situation the mean of the maximum GRF of 10 stance phases of the right leg was analyzed for each trial. Every stance phase began with the heel contact and ended with toe off. For calculations a threshold of 50 N was used. For each trial within the static condition the mean vertical GRF (= average body weight) was included. The entire analysis was conducted using custom written Matlab (R2010a, The Mathworks, Natick, USA) code. From those measurements a relative force reduction over pressure function was derived. To detect correlations between the individual unloading situation and the adjusted pressure level a linear regression analysis was performed, whereas the pressure level represents the independent variable and the resulting vertical GRF the dependent variable. In the fitting procedure, regression lines were forced to always match 100% body weight in the 0 Pa condition (Fig. 2.). Therefore, only the slope of the regression line was varied in order to find the regression line that minimizes the sum of squared residuals between measured and predicted data points. Due to this approach, linear relative unloading pressure formulae between subjects differed only with respect to the slope of the regression line. Based on this, two methods were developed to calculate the individual unloading situation with respect to the body weight. The first method assumed that a linear relationship between relative unloading and respective pressure levels was existent. Therefore, four calibration measurements at different pressure levels (e.g. 100, 200, 300 and 400 kPa) had to be performed. Based on those calibration data, an individual regression line was generated, allowing necessary pressure levels to be calculated from the requested weight reduction. The second method was developed in order to avoid calibration measurements and the necessity to use an instrumented treadmill. It was assumed that the individual slope of a load reduction over pressure regression line calculated from the static condition was mass. Therefore, nal regression line was derived, taking all 12 subjects into account. Coming back to that final regression, it was possible to determine an individual pressure-unloading relationship and no further calibration measurements." ] }, { "image_filename": "designv11_101_0001740_9781782421702.12.749-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001740_9781782421702.12.749-Figure3-1.png", "caption": "Figure 3 \u2013 Geometrical model #1 (vane is in contact with vane slot at V1)", "texts": [ " Point V2 is a point at the left hand side of the vane that is of a distance of the cylinder radius, rc, from the cylinder center, C. This is the contact point when the vane contact is at the left hand side of the vane. Points Vb1 and Vb2 are points at the right hand and left hand sides of the base of the vane. They are of distances of the rotor radius, rr, from the rotor center, R. The angle of the cylinder, \u03d5c, is defined as the angle between the vertical line, which passes through C and R, and a line that connects point C and the midpoint of the vane slot opening, S0 (see Figure 3). Point S2 is the tip of the left hand side of the vane slot opening. Point R1 is the projection of point S1 at the midline of the vane, while point R2 is the projection of point S2 at the midline of the vane. When the vane is in contact with the vane slot at its right hand side as illustrated in Figure 2, the distance between points R1 and S1 is equal to half of the vane width, wv. Point C1 is a point at the line that connects point C with the midpoint of the vane slot opening, So. The distance between point C1 and point S1 is equal to half of the width of the vane slot opening, wvs. There is also point C2 (not shown in the figures) that is of the same distance from point S2. However, because this point always coincides with point C1, point C2 will be merged with point C1 in the subsequent discussions. Figure 3 also shows that points V1, V0, V2, S1, S0 and S2 are all of the distances of the cylinder radius, rc, from the cylinder center, C. To determine the angle of the cylinder, the triangles CS0S1, CV1R and RV0V1 shown in Figure 3 are used. By observing the triangles, the relationship between the various geometrical parameters according to Equation (1) can be derived. If the contact point is at the left hand side of the vane, the equation should be adjusted accordingly. Equation (1) can then be used to find the angle of the cylinder, \u03d5c, when the vane is in contact with the vane slot. ( ) ( ) r v vCSScccrCSScc w werrer \u03d5\u03b8\u03d5\u03d5\u03b8\u03d5 cos 2 25.0cos2sinsin 222 1010 +\u2212+\u2212+=+ (1) During the operation, the vane extends and retracts in and out of the vane slot" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001873_phm.2011.5939468-Figure6-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001873_phm.2011.5939468-Figure6-1.png", "caption": "Figure 6. A prototype of ED/EP for the gearbox", "texts": [ " Similarly, it needs to determine the location of the hole from the center of the gear and its diameter and depth, so that drilling such a hole just causes an acceptable impact on the structure integrity. Here natural frequencies and modal shapes of the gear are two important factors. Next we consider six bearings and four gears as ten candidate locations of embedded sensor modules. Then an optimal model is built based on quantitative signed directed graph (QSDG) of fault vibration propagation, which is used to determine how many and where sensors to be embedded optimally. The final result corresponds to {@, @, @}. Finally, a modified gearbox shown in Fig.6 is designed and manufactured according to the results and a miniaturized vibration sensor module is designed [23], where the location of sensor 1 corresponds to @. An artificial fault is seeded in the inner race of the bearing CD and its fault frequency is calculated as 60.2 Hz. Two meshing frequencies at high and low speed stage are 480Hz and 125.6Hz respectively. Then vibration signals from three embedded sensors and an outside sensor on the housing are collected. These signals are analyzed and their power spectrums are shown as Fig" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000120_gt2013-95575-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000120_gt2013-95575-Figure1-1.png", "caption": "FIGURE 1. SEAL MODEL AND COMPUTATIONAL DOMAIN", "texts": [ " Then the first order bulk-flow equations are solved using the 2 Copyright c\u00a9 2013 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/conferences/asmep/76978/ on 03/04/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use CFD derived base state primitive variables. The reaction forces are determined by integrating the perturbed pressure field over the surface of the rotor and the rotordynamic coefficients are determined by using the equations of motion. Figure 1 shows a cut away view of a hole-pattern seal around a rotor along with the sector fluid domain geometry used in the hybrid approach. Following Ha and Childs [4] and Shin and Childs [6], the two-control volume governing equations for non-isothermal, adiabatic flow of an ideal gas with standard Hirs bulk-flow assumptions [1] (after manipulation) are \u2202 \u2202 t [\u03c1 (h+ \u03b3chd)]+ \u2202 \u2202 z [\u03c1wh]+ 1 R \u2202 \u2202\u03b8 [\u03c1uh] = 0 (1) \u03c1h [ \u2202w \u2202 t + u R \u2202w \u2202\u03b8 +w \u2202w \u2202 z ] +h \u2202 p \u2202 z + \u03c4z = 0 (2) \u03c1h [ \u2202u \u2202 t + u R \u2202u \u2202\u03b8 +w \u2202u \u2202 z ] + h R \u2202 p \u2202\u03b8 + \u03c4\u03b8 = 0 (3) \u03c1h [ \u2202e \u2202 t + u R \u2202e \u2202\u03b8 +w \u2202e \u2202 z ] + \u03b3c\u03c1Hd \u2202e \u2202 t + \u2202 \u2202 z [pwh] + 1 R \u2202 \u2202\u03b8 [puh]+R\u03c9\u03c4r\u03b8 = 0 (4) p = zc\u03c1RgT (5) where e = 1 zc(\u03b3\u22121) p \u03c1 + 1 2 [ u2 +w2] (6) Equations (1) - (4) represent the conservation equations for mass, axial momentum, circumferential momentum, and energy in the jet flow region" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003294_978-3-642-39128-6-Figure2.7-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003294_978-3-642-39128-6-Figure2.7-1.png", "caption": "Fig. 2.7 Replacing the patch clamp pipette with a microstructured chip. a Whole-cell configuration of the conventional patch clamp technique. The tip of a glass pipette is positioned onto a cell using a micromanipulator and an inverted microscope. b Whole-cell recording using a planar chip device having an aperture of micrometre dimensions. Cells in suspension are positioned and sealed onto the aperture by brief suction [12]", "texts": [ " In both lateral and planar patch clamping, the cell is positioned on the pore and suction is applied to facilitate gigaseal formation. In comparison with conventional techniques, planar and lateral patch clamping configurations do need costly equipment, such as a precise manipulator, a high-magnification microscope and an anti-vibration table. However if the material of the chip is transparent then lateral patch clamping can provide optical access to the patching site. Fertig et al. [12] conducted one of the earliest attempts in planar patch clamping. Their design is shown in Fig. 2.7. The figure also shows the difference between conventional and planar patch clamp configurations. The idea of bringing cells to the patching site rather than bringing the pipette to the cells has drawn the attention of many researchers and hundreds of different designs have been developed, each with some advantages over the others. In planar patch clamping the same hole is used for both cell positioning and gigaseal 2.3 Attempts to Improve Patch Clamping 10 2 Development of Patch Clamping formation" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001498_1071181311551342-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001498_1071181311551342-Figure1-1.png", "caption": "Figure 1: Footwear design variables", "texts": [ " Two marks, on the lateral side corresponding to the point at which the foot touches the ground, and the end of the fifth metatarsal were marked with a pen in order to consistently locate the foot on the footbed simulator. The footbed simulator allowed the quick change of footbed shapes while the F-Scan system (Tekscan, 2010) was used to obtain the plantar foot pressure profiles. The independent variables were heel height, wedge angle and type of midfoot support. Two heel heights (50 and 75 mm), three wedge angles (Figure 1) nested under heel height (4\u00b0, 10\u00b0, 14\u00b0 at 50 mm; 14\u00b0, 18\u00b0, 22\u00b0 at 75 mm) and three different types of mid-foot supports (45- PU, HL-PU, HL-SS) corresponding to different seat lengths (45 mm and anatomical heel length, HL) and two types of materials, PU and SS were the experimental conditions. The toe spring was controlled C op yr ig ht 2 01 1 by H um an F ac to rs a nd E rg on om ic s S oc ie ty , I nc . A ll rig ht s re se rv ed D O I 1 0. 11 77 /1 07 11 81 31 15 51 34 2 at UNIV OF LETHBRIDGE on October 18, 2015pro" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000181_amm.128-129.142-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000181_amm.128-129.142-Figure1-1.png", "caption": "Fig 1 Ancillary air bursa control mode", "texts": [], "surrounding_texts": [ "The airship control system consists of attitude controller and altitude controller. The attitude control system of airship can not only realize to control attitude of airship, but also control altitude rapidly. Especially when airship flies at high velocity, the aerodynamic lift force is very considerable. The attitude control system of airship is given in Fig (2). In this figure, the attitude of airship is controlled by elevator and ancillary air bursa, which both works at moment mode. When the airship ancillary air bursas charge or deflate simultaneously, the weight and buoyancy are changed; the airship can rise or decline. So the ancillary air bursas can not only control the attitude of airship to track the altitude command, but also control the altitude directly. In this control system, the attitude command can also be tracked when efficiency of air helms is low. The parameters of the attitude controller and altitude control system are design by using fuzzy self-tuning PID. This includes classical PID controller and fuzzy controller. The structure of the fuzzy self-tuning PID controller is given in Fig (3) The parameters of controller , ,p i dK K K : { } { } { }0 0 0, ; , ; ,p p p k i i i k d d d kK K K e eck K K K e eck K K K e eck= + \u2206 = + \u2206 = + \u2206 , ,p i dK K K\u2206 \u2206 \u2206 represents modified efficiency of PID, which is computed by fuzzy controller. ,ke eck represents error and error rate, 0 0 0, ,p i dK K K are the parameters of classical PID. The fuzzy controller adjusted the classical PID parameter according to the fuzzy logical." ] }, { "image_filename": "designv11_101_0000933_ecce.2013.6647385-Figure5-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000933_ecce.2013.6647385-Figure5-1.png", "caption": "Fig. 5. Rotor cross section of 0.6 kW 4-pole PM motors.", "texts": [ " - The compressor motors have relatively small current and large (Lq \u2013 Ld), which results in the small \u03b1 of about 0.2. Even if the (Lq \u2013 Ld) or the current is doubled, the \u03b1 is only increased to about 0.4. Therefore, the less saliency will be preferable (see section 4). - The HEV motors have large current and little (Lq \u2013 Ld), which results in the large \u03b1 of about 0.6. An increase in the (Lq \u2013 Ld) or in the current almost directly increases total output torque, and hence the large saliency will be preferable. IV. APPLICATION TO 4-POLE 0.6KW PM MOTORS Fig. 5 shows the two kinds of rotor cross section for 0.6 kW 4-pole PM motors driven by a PWM inverter. Table 2 shows specifications. The model A is the original structure, composed of V-shaped magnets in order to increase the Lq and hence the reluctance torque, while the model B is a proposed measure, consisting of flat-plate magnets and smaller outer core region so that the Lq is reduced. In the following, all results are described in per-unit system. Reference values for the per-unit system are the peak values of the rated phase voltage \u22c52 UN = \u22c52 57" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000678_mmar.2013.6669966-Figure5-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000678_mmar.2013.6669966-Figure5-1.png", "caption": "Fig. 5. Scheme of a DC motor.", "texts": [ " 3, for the additional state variables in order to get the open loop tracking system. 4. The closed loop tracking systems is obtained by calculating the feedback gains and connecting these gains according to the Fig. 4. The proposed methodology to obtain the closed loop tracking system in a bond graph approach is applied in the next section V. ILLUSTRATIVE EXAMPLE The theory presented in previous sections can be conveniently illustrated by designing a controller which will cause the output of the controllable second-order linear plant whose figure 5 shows a scheme of a DC motor. The bond graph of the DC motor is shown in Fig. 6. The key vectors of the bond graph are x = [ p 3 p 7 ] ; \u2022 x = [ e 3 e 7 ] ; z = [ f 3 f 7 ] ;u = [ e 1 e 8 ] D in = [ f 2 f 6 ] ;D out = [ e 2 e 6 ] ; y = [ f 3 f 7 ] the constitutive relations of the elements are L = diag {R a , b} (26) F = diag { 1 L a , 1 J } (27) and the junction structure is e 3 e 7 f 2 f 6 f 3 f 7 = 0 \u2212n \u22121 0 1 0 n 0 0 \u22121 0 1 1 0 0 0 0 0 0 1 0 0 0 0 1 0 0 0 0 0 0 1 0 0 0 0 f 3 f 7 e 2 e 6 e 1 e 8 (28) The goal of this system is ,to track the command input vector [ v 1 (t) v 2 (t) ] = [ 2t t ] , (0 \u2264 t <\u221e) In this case it can be seen that the command input is of the form (9) with r = 2 and applying the proposed procedure, Fig" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001742_amr.442.246-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001742_amr.442.246-Figure2-1.png", "caption": "Fig. 2.Contact relationships among the stress recovery model", "texts": [ " And directions of 4, 5 and 6 degrees of freedom correspond to the directions rotating around x, y and z axis. The established modal shrinkage-reducing model of the block is shown in Fig.1. The major difference between the modal shrinkage-reducing model and the stress recovery model, of the block, is the contact relationship of every part in the block component. Contact relationships among modal shrinkage-reducing model are tie-style binding relationships while among stress recovery model are common contact relationships, as shown in Fig. 2. Moreover, the node set that extracts the main DOF is defined in the modal shrinkage-reducing model. All rights reserved. No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of Trans Tech Publications, www.ttp.net. (ID: 128.210.126.199, Purdue University Libraries, West Lafayette, USA-01/06/15,21:34:39) With modal reduction calculation of every module finished via FE software, dynamic calculation begins at the dynamics simulation platform", " The structure of the block is complex, so, to be convenient to analyze, it\u2019s divided into several parts. As shown in figure 3, the block is divided into 7 parts named A~G in the horizontal direction and is divided into 3 parts named 1~3 in the longitudinal direction. There are 3 dangerous areas that has low safety ratio----area at the top of the block, area near the inner end of the cam bearing hole in the mid of the block and area at the knuckles of the main bearing end which is at the bottom of the block. It\u2019s shown in the diagram area of figure 2. Under that 3 explosion pressure conditions, distribution nephograms of safety coefficient are similar. Figure 4a~c show the safety coefficient distribution of the 3 dangerous areas. Figure 5 shows the stress development chart of the points that has the least safety coefficient in that 3 dangerous areas, and their data are shown in Table 2. According to the data from the table, the least safety coefficient of the block is 2.142, located on the inner end of the 6th cam bearing hole, as shown in figure 3b" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001113_amm.433-435.17-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001113_amm.433-435.17-Figure2-1.png", "caption": "Fig. 2. Rigid-flexible coupled model of PSD. Fig. 3. Flexible model of PGS.", "texts": [ " Data containing noise will lead to abnormal operation of model in the simulation, thus the noise in test data must be reduced. The wavelet denoising method has a good effect on non-stationary signal [5] and it has been applied to many engineering fields [6,7]. In this paper, an improved threshold function [8] in wavelet denoising method is used to denoise test data. Through establishing the rigid-flexible coupled virtual prototype of PSD, the dynamic simulation at the moment of PGS failure is carried out. The maximum stress of PGS is obtained. The rigid-flexible coupled model is shown in Fig. 2. The flexible PGS is added rigid regions using ANSYS software, as shown in Fig. 3. Then the modal neutral file with 20th mode is imported to ADAMS software to carry out the dynamic simulation. All rights reserved. No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of Trans Tech Publications, www.ttp.net. (ID: 132.174.254.159, Pennsylvania State University, University Park, USA-24/05/15,22:39:40) The load torque of left half axle, driving rotational speeds of PSD housing and right half axle are denoised by new threshold function (Eq" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001402_s13369-012-0232-3-Figure8-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001402_s13369-012-0232-3-Figure8-1.png", "caption": "Fig. 8 Coordinate systems for determining a rack cutter", "texts": [ " The maximum stress of von-Mises was 19.7 MPa. Although the mathematical model of the gear and pinion was obtained in Sect. 3, the rack cutter for manufacturing the pinion or the gear was unknown. Based on inverse envelope concept [5,6], the corresponding rack-cutter contour for manufacturing the pinion or the gear can be determined. The developed mathematical model of the pinion in Sect. 3 became the generating surface, and the contour of the rack cutter is considered as the generated surface. Figure 8 shows three coordinate systems that mathematically describe the geometric model of the rack cutter. Coordinate system S\u2032 2(O \u2032 2, x \u2032 2, y\u2032 2, Z \u2032 2) was rigidly attached to the pinion, where z\u2032 2 axis is the rotation axis of the pinion. The pinion rotated through angle \u03c6\u2032 2 at the z\u2032 2 axis. Coordinate system S\u2032 c(O \u2032 c, x \u2032 c, y\u2032 c, Z \u2032 c) was rigidly attached to the workpiece, which translated over the linear displacement S along the y\u2032 c axis. Coordinate system S\u2032 f(O \u2032 f , x \u2032 f , y\u2032 f , Z \u2032 f) was rigidly attached to the fixed frame" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002083_ifost.2011.6021074-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002083_ifost.2011.6021074-Figure3-1.png", "caption": "Figure 3. The control vector diagram by the side of grid.", "texts": [ " Meanwhile, the inner ring control can regulate the electromagnetic torque directly. We set the given rotor speed is W *, the given q-axis component of stator current is iq *. III. THE CONTROL STRATEGY BY THE SIDE OF GRID The controlling goal by the side of grid is making sure that the voltage of the DC capacitance is constant. It also needs to ensure that the output reactive power is O. Therefore, the direction of grid voltage integrated vector is taken as the d axis, and the q-axis is ahead of d-axis 90 degrees. The control vector diagram by the side of grid is shown in Fig.3. ed is d-axis component of the grid voltage, igd is d-axis component of the grid current, and Bg is the angle between the d-axis and the phase A, which is given by e () = arccos-'!... g e, (12) where ea is the phase A voltage of the grid, and es is the value of the grid voltage integrated vector. The grid voltage can be known from the system, so that the value of the grid voltage integrated vector is given by (13) In this case, we can know that eres, eq=O, and the power equation by the side of grid can be simplified as {p = \ufffd (edigd +eqigq) = \ufffdeJgd (14) Q 3( " ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001431_amr.694-697.512-Figure4-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001431_amr.694-697.512-Figure4-1.png", "caption": "Fig. 4 Face gear tooth surface curve Fig. 5 Simulation of offset face gear", "texts": [ " Concrete steps: Insert\u2192 curved surface\u2192 Surface lofting\u2192 Surface seaming\u2192 Array\u2192 Finish. Because face gear surface is asymmetric, the other surface can not be obtained directly through the mirror and needs double counting. Obtained by face gear processing principle, the top and bottom surfaces of gear are flat along the axial direction, and the front and rear sides, which are close to the inside and outside diameter, are cylindrical surface [5]. Then get the simulation of offset orthogonal face gear (Fig. 4, Fig. 5). By deriving the surface equation and writing programs, the tooth surface simulation model of the offset orthogonal curved tooth face gear has been obtained finally. Importing tooth surface data points, which is created by programming, into SOLIDWORKS, and modeling the whole offset orthogonal curved tooth face gear, that makes preparations for later strength analysis, dynamic analysis and so on. [1] R.P. Zhu, S.C. Pan and D.P. Cao: Journal of Nanjing University of Aeronautics and Astronautics Vol" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000741_sustech.2013.6617332-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000741_sustech.2013.6617332-Figure3-1.png", "caption": "Fig. 3. Harmonic flux trajectory in half a switching cycle.", "texts": [ " This method is valid for an inductance load or a SPM synchronous machine load as well. Computer simulations based on MATLAB are conducted to verify the accuracy of the proposed method. II. CURRENT THD CALCUALTION FOR INDUCTANCE LOAD This session reviews the traditional method to calculate current THD for an inductance load. Based on Kolar and Abraham\u2019s research work, Hava concluded space vector approach as the most straightforward method for analytical investigation of the switching frequency harmonic characteristics of a PWM-VSI [5]. As illustrated in the vector diagram in Fig. 3, the sequence of the applied voltage vectors follows V0 V1 V2 V7 in the first half switching cycle. For these four stages, a deviation between the applied voltage vector and the command voltage vector always exists. The harmonic flux linkage may be obtained by integrating the voltage deviation as shown in (1) [5], where . In Fig. 3, the corresponding harmonic flux linkage trajectory follows \u201cO A C B O\u201d in this half switching cycle. In the second half of this switching cycle, the harmonic flux can be calculated symmetrical to the first half. In the space-vector theory, the kth harmonic of the threephase flux in the time domain can be translated to a complex flux vector which rotates in the complex coordinates with the angular velocity k e in the following, where , The RMS value of the total harmonic flux linkage in one fundamental cycle, , may be calculated using (3), where T is the fundamental cycle" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001514_s12257-011-0363-5-Figure5-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001514_s12257-011-0363-5-Figure5-1.png", "caption": "Fig. 5. New mechanism for electro-enzymatic nitrate reduction.", "texts": [ " However, this proposed nitrate reducing mechanism cannot be applied to the electro-enzymatic denitrification system since in the biological nitrate removal system, the electrons are transferred to nitrate reductase via cofactors. But in the electro-enzymatic system, the electrons are obtained directly from the bio-electrode and the electrons are instantly supplied to nitrate reductase. Furthermore, the proposed mechanism does not contain a detailed electron transfer pathway. For these reasons, a new mechanism for bio-electrochemical nitrate reduction is needed (Fig. 5). The process of denitrification through the proposed electro enzymatic denitrification system is composed of five steps. As shown in Fig. 4, in the first step, the hydroxyl group is detached and nitrate is attached to the metal center of nitrate reductase. Figs. 4A, 4B, and 4C shows this process. The nitrate reduction process requires 2 electrons, and the electrons are supplied during (i) step. For these reasons, the following assumption can be made. Considering the structure of the Mo orbital, the hydroxyl group is detached first and the nitrate molecule is attached later", " Step (c), because the enzyme substrate complex is unstable, the complex gets another electron and is changed into the activated form. After forming the activated enzyme substrate complex, the complex is separated releasing nitrite and just the enzyme ((ii) step in Fig. 4). This step is included in the proposed electro-enzymatic mechanism as step (d). Finally, in step (e), the enzyme gets another proton and is changed back into its initial form. 3.3. Proposed model equation of nitrate reduction rate From the proposed mechanism (Fig. 5), the following equation for nitrate reduction can be proposed. E + H+ + e\u2212 \u2192 E* + H2O (a) E* + NO3 \u2212 E* NO3 \u2212 (b) E* NO3 \u2212 + e\u2212 E** NO3 \u2212 (c) (1) E**NO3 \u2212 E\u2212 + NO2 \u2212 (d) E\u2212 + H+ \u2192 E (e) In this electro-enzymatic nitrate removal system, the electrons and protons are readily supplied from the bioelectrode. For this reason, the first step and final step occur immediately. Therefore, the proposed mechanism can be reduced to 3 steps (2). E* + S E** S (a) E** S + e\u2212 E** S (b) (2) E** S E\u2212 + P (c) (E: nitrate reductase, S: nitrate, P: nitrite) Among these 3 steps, it was assumed that the 2-c step was the rate determining step since in the 2-a step, the enzymes exist in the radical form and in the 2-b step, the electron is readily supplied from the bio-electrode" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000046_amm.186.194-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000046_amm.186.194-Figure2-1.png", "caption": "Fig. 2 The angular position of the tooth, \u03b8(t)", "texts": [ " Torsor and tool position determination The aim being to evaluate the torsor at the tooth tip using results of the experimental protocol were performed several steps. With the data provided by the position transducers, the instant real positions on the three linear axes and the angular position in the center of the tool are known. Starting at this position, tooth tip position is calculated. Furthermore, the transport of moments can be done, from the dynamometer\u2019s center to the milling tool center and to the tooth tip of the tool. Evaluation of the instant position at point OD, representing the tooth tip is represented in Fig. 2. Applied Mechanics and Materials Vol. 186 195 Determining the instant position at the center of the tool in point OS(t). Based on the on-line acquisition system of tool position the axis coordinates of the tool can be known at any time (t). So, the instantenous position of the point of the tool's rotation axis, representing the center of the mill tool in point OS(t) is made known with Eq. 1 [7]: )( )( )( :)( tZ tY tX tO S S S O O O S (1) Determining the instant position at the tooth tip of the tool OD(t)" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003551_20130828-3-uk-2039.00008-Figure10-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003551_20130828-3-uk-2039.00008-Figure10-1.png", "caption": "Fig. 10. Task 2: Lane Assist", "texts": [], "surrounding_texts": [ "Since the first cycle of both courses is still running, only preliminary results from a short survey can be given. Due to this, a comparison of learning processes to a preliminary year of students was not possible until now. All in all, the feedback of the participating students was very positive (e.g.: \u2019More such sessions wanted!\u2019). Many students declared that the participation increased their understanding of the course topics. The possibility to get a hands-on approach to control engineering was often acknowledged. One very positive effect, which could be seen in the student contest was, that the contest participants come from different fields (such as informatics as well as mechanical and electrical engineering). Hence, they were able to support each other very well because of their completely different backgrounds." ] }, { "image_filename": "designv11_101_0001650_i2mtc.2013.6555566-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001650_i2mtc.2013.6555566-Figure2-1.png", "caption": "Figure 2. Welded plates with dimensions and the representation of a tensile specimen.", "texts": [ "404A With this formula, it is possible to find the multiplication factor A for each one of the welding experiments that were performed. The values for each experiment performed in this study are given in the Table IV. Table V shows the thermal nodal load for each one of the welding experiments. . (7) 1, 2 and 3 are (8) \u03b7 is the heat efficiency of the arc. The results of this model were validated by experiments measuring the temperature in different zones, transverse to the heat source as shown in Fig. 2. were obtained by measuring the temperature of welded sheets using contact thermocouples. Thermocouples were placed on a line perpendicular to the weld, to cover a distance equal to the width of the heat affected zone. The thermocouples that were used are chromel type E Class 1, with a measurement range from 900\u00b0C. The thermocouple error limit was limit extension cord with a standard error of sheets of AA6061\u2212T6 4.8 mm thick were welded ER4043 electrode by the Gas Metal Arc Welding (GMAW) process using the typical simple butt To perform the welding, the Lincoln Power MIG machine was used" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000160_2013-01-1757-Figure14-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000160_2013-01-1757-Figure14-1.png", "caption": "Figure 14. Flux distribution (a) conventional magnet, (b) Dy localized diffusion magnet", "texts": [ " At the center of the magnet, coercivity was the same as before Dy local diffusion treatment, but it increased towards magnet corners and peaked at the corners, with an increase of more than 320 kA/m. This satisfied our target [Figure 13 (a)] for necessary coercivity in all areas [Figure 13 (b)]. Magnetic Flux Density and Magnetic Flux Distribution In general, adding Dy to a magnet to increase coercivity causes a drop in residual magnetic flux density(7). With the magnet prototypes in this study, local diffusion treatment of Dy was not observed to cause a decrease in residual magnetic flux density. Figure 14 shows the comparison, made with a magnet analyzer, of flux distribution between a conventional magnet and our developed magnet. The flux distribution did not differ with or without local diffusion treatment of Dy. We believe that, with a conventional magnet, Dy is spread throughout the Nd2Fe14B phase and the magnetic moments of Dy and Fe cancel each other out to decrease the overall magnetic flux density. With a magnet with locally diffused Dy, however, Dy is in grain boundary regions where the melting point is low, and magnetic flux density decreases due to suppression of increased coercivity", " We also compared torque and power characteristics between this motor and a motor using a conventional high-concentration Dy magnet. For these characteristics, measurements were made on motorinduced torque and power when a specific electric current was applied. The results are shown in Figure 16. We confirmed that those two types of motor had equal characteristics, which means local diffusion treatment of Dy did not cause a decrease in induced torque. This is consistent with the results we mentioned above where flux distribution was equal for the two types of magnets (Figure 14). Demagnetization Characteristics A motor test was conducted to compare demagnetization characteristics of a conventional magnet and a magnet with locally diffused Dy. Magnetization characteristics were measured before and after applying maximum electric current at a specific temperature to produce maximum demagnetization fields in the magnet. The results are shown in Figure 17. The developed magnet showed demagnetization characteristics more than equal to those of the conventional magnet, even though it contained 30% less Dy" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001680_ccdc.2013.6561389-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001680_ccdc.2013.6561389-Figure1-1.png", "caption": "Fig. 1: Critical collision angle of point-shaped obstacle", "texts": [ " All the obstacle boundaries in C-space consist of these close curves, which is called C-boundary. The critical collision angles are also obtained when the small mechanical arm collides with the point-shaped obstacle P, including the upper critical collision angle and the 2652978-1-4673-5534-6/13/$31.00 c\u00a92013 IEEE lower critical collision angle. The upper critical collision angle is obtained when the mechanical arm which rotates in a clockwise direction collides with the point-shaped obstacle P, as shown in Fig.1 \u03b8ta. The lower critical collision angle is obtained when the mechanical arm which rotates in an anti-clockwise direction collides with the point-shaped obstacle, as shown in Fig.1 \u03b8ba. The round-shaped obstacle can still be processed as a point-shaped obstacle because its radius can be equivalent to enlarge the width size of the mechanical arm, as shown in Fig.2. The line-shaped obstacle is seen as a set of point-shaped obstacles, whose critical angles are naturally composed of the critical angles of these point-shaped obstacles. It is assumed that the joint angles of non-collision mechanical arms are fixed and the working range of the collision mechanical arm is [-\u03c0,\u03c0], when computing the critical collision angles" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001629_amr.383-390.2963-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001629_amr.383-390.2963-Figure3-1.png", "caption": "Fig. 3. On-load operation", "texts": [ " By means of finite element software, the field-circuit coupled model is analyzed. Taking the entire circumference as the solving area, the distribution of the magnetic force lines in the case of no-load operation, on-load operation and locked-rotor operation can be obtained as shown in Fig. 2 to Fig. 4. It can be seen from Fig. 2 that the no-load motor has the regular distribution of the magnetic force lines which mainly go through the air gap vertically. When the motor is on-load operation, the magnetic force lines display with distortion mildly in Fig. 3. But in Fig. 4 the magnetic force lines display with distortion seriously under locked-rotor operation, even some lines don\u2019t loop throughout air gap. It is mainly because the magnetic field which is produced by the eddy current changes the distribution of stator and rotor magnetic field, which can make the magnetic force lines display with distortion. The distortion changes seriously with locked-rotor time duration. At this time copper loss and iron loss of the motor increase continually. This will lead to a rapid rise in the temperature of the motor" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000198_indin.2013.6622905-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000198_indin.2013.6622905-Figure2-1.png", "caption": "Fig. 2. Principle of gyroscope.", "texts": [ " A healthy wrist can do the following motions: flexion, extension, radial deviation, ulnar deviation, pronation and supination while legs can only do flexion and extension movements. This type of injury often occurs in adulthood, and mostly in women, since osteoporosis increases the fragility of bones, making them easy to be broken. In the youth, this type of fracture is mainly due to injuries in sport activities. The Powerball\u00ae is a mechanical device [1] (fig. 1), existing in the market, which can be used for wrist rehabilitation. The operation of Powerball\u00ae is based on the principle of a gyroscope [2-3] (fig. 2). Suppose that a disc is rotating around the x axis with angular velocity \u03c9x and there are some external actions (in case of Powerball\u00ae, it\u2019s the movement of the wrist, arm) that makes the disc rotating around the y axis an angle \u03b4\u03b8. The change in momentum will be: (1) The reaction moment results in a tendency to make the disc to rotate around the z axis. The structure of the Powerball\u00ae is shown in Figure 3. The main part of the Powerball\u00ae is a spinning mass (2) inside an outer shell (1). The shell almost completely covers the inside mass, with only a small round opening allowing the gyroscope to be manually started" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000532_acc.2013.6580732-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000532_acc.2013.6580732-Figure1-1.png", "caption": "Fig. 1. Schematic representation illustrating a mobile robot with a serial redundant arm (qi represent generalized coordinates)", "texts": [ " The COIN algorithm introduced in this paper accomplishes this adaptive manipulation control as a secondary task that supplements the primary objective of inverse kinematics [9]\u2013[13] and path following through redundancy resolution. COIN is capable of using a priori training data to achieve fast global convergence, and enable a mobile robot to correct its position and arm posture in real-time to offset the risk of tip-over instability during eccentric manipulation tasks. For a mobile robot with a serial redundant arm [14]\u2013[15] mounted on a base \u2013 such as in the illustration shown in Fig. 1 \u2013 the tilting moment that the n-links generate around the pivot can be written in a compact recursive form as 1 3 1 0 0 0 0 1 1 1 3 13 6 1 ( ) ( ) ( ) ( ) in i i Gi i i i ext i i g M q d q R q R q (1) where 0 3 1 Gi d denotes the vector position of the center of mass iG of link i expressed in frame 0 , and 0 3 3 Gi d the skew-symmetric matrix of vector Gi d . 0 1iR defines the rotation matrix from frame i \u2013 1 to frame 0, Mi the mass of link i , 1i ig the gravitational acceleration of frame i expressed in frame 1i , and 1i ext i the external moment vector (inertial forces, etc" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001032_s11668-011-9448-x-Figure7-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001032_s11668-011-9448-x-Figure7-1.png", "caption": "Fig. 7 As-received pulley assembly after disassembly. Note the fractured guides", "texts": [ " In addition, a reference timing belt pulley bolt was submitted for reference. The failed crankshaft and failed pulley bolt were reported to have failed approximately 5000 km after the timing belt was changed. No maintenance records are available to confirm this. The as-received timing belt assembly is shown in Fig. 4. A closer view of the failed pulley assembly is shown in Fig. 5, showing the relationship between the pulley assembly and the crankshaft nosepiece. The failed bolt is shown along with the reference bolt in Fig. 6. The pulley is shown in Fig. 7. Visual Examination The crankshaft nosepiece, failed timing pulley bolt, woodruff key, and timing belt assembly were examined optically using a stereo microscope at different magnifications. The crankshaft noise piece showed visual evidence of secondary cracking in the keyway, and in a semi-circular fashion back toward the fracture surface (Fig. 8). Opposite D. Scott MacKenzie (&) Houghton International, Inc., Valley Forge, PA 19426, USA e-mail: smackenzie@houghtonintl.com the keyway, burnishing of the end of the nosepiece was observed" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001252_peoco.2012.6230918-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001252_peoco.2012.6230918-Figure1-1.png", "caption": "Figure 1. stator current spatial phaser and rotor conn reference framework and in special framework installed spatial phasor.", "texts": [], "surrounding_texts": [ "Abstract-- In this paper, indirect rotor field oriented vector control has been studied for Squirrel Cage Induction Motor (SCIM) Drives and it has been indicated that Through vector control, flux-producing current components and torque will be separated and transient response characteristics are similar to separately excited DC machine and System with any load fluctuations or changes in reference values to same DC machine speed can be adjusted and will be the dynamic Like to DC machines.\nIndex Terms-- Vector Control; Squirrel Cage Induction Motor (SCIM) Drives.\nI. INTRODUCTION C motors are widely applied in conditions where there is increasingly variable speed performance, as their flux and\ntorques are simply controlled by field and armature current .In particular separately excited DC motor is applied increasingly for applications requiring fast response and quadrangle functions with high performance near zero speed, however, as DC motors are equipped to commutator and brushes, they suffer from some limitations, such as, periodical repairs, work constraints in high voltage and speed. These above mentioned problems may be resolved through using intermittent current motors whose structure is very simple and are very cost \u2013 effective. In comparison with DC motors, its small dimension generally allows ac motors to design for higher output to less rotating mass. There are AC different operating system, among these systems the systems carrying squirrel cage, have special and practical advantages over other systems. Squirrel cage motor is simple and resistant (strength) and is also considered one of the cheapest machines with high performance and is readily available in power. Vector control techniques. e.g fast microprocessors, are abled using induction motor drives and synchronous motor with high performance in where traditional DC drives are applied in the past [3]-[6]. In the past, these control techniques were not applicable as they are involved complex hardware and have essential software in order to solve complex control problems. Like DC machines, AC machines torque control is possible or accessible through controlling motor currents, even though, unlike DC machines, in AC machine, phase angle and current limit should be controlled, in other words, current vector should be controlled, which is so called ''vector control'' [1]-[2]. In addition, for AC\nMehdi Mohammadzadeh Rostami is with Department of Power Engineering, Science and Research Branch, Islamic Azad University, Tehran, Iran (E-mail: edu.rostami@gmail.com).\nmachines, field flux orientation and armature mmf should be kept stable, while, for ac machines, field flux line (orientation) and armature mmf need external control, and without this control, angle size among different fields will be changed for AC machines with load and as a result, we see unwanted oscillatory response. Through vector control, AC machines, flux-producing current components and torque will be separated and transient response characteristics are similar to DC machine characteristics with separately excited [4]-[5].\nII. MATHEMATICAL MODEL AND VECTOR CONTROL In a squirrel cage induction motor, descriptive equations of\nspatial phases in static reference are obtained as follows:\ndt d iRu s sss \u03c8 +=\nrr r rrr j dt\ndiRu \u03c8\u03c9\u03c8 \u2032\u2212 \u2032 +\u2032=\u2032\nsr r\nm e i L LPT \u00d7\u2032= \u03c8 2 3\nrrLoade DJTT \u03c9\u03c9 +=\u2212\nHere, we emphasize that stator current zero component is supposed zero during vector control actions. Through above equations, we can obtain a biaxial model, which enables us to implement machine state equations (8). In this procedure, spatial phasor equations are obtained in transient conditions as follows:\nsqr r\nm e i L L PT \u03c8 \u2032 \u2032 = 2 3\nm\nr sd L\ni \u03c8 \u2032 =\nmr\nrr sq LR\nLrs i \u2032 \u2032\u2032 = \u03c8\u03c9\nAnalysis of Indirect Rotor Field Oriented Vector Control for Squirrel Cage Induction Motor Drives\nMehdi Mohammadzadeh Rostami\nD\n(5)\n(4)\n(2)\n(6)\n(7)\n(1)\n(3)\n978-1-4673-0662-1/12/$31.00 \u00a92012 IEEE 505", "\u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a3\n\u23a1 \u2212 \u2212\n\u2212 = \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a6\n\u23a4\n\u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a3\n\u23a1\ns\ns mrs\nr\nry\nrx\nsy\nsx\nLL R\nL R LLL i i i i\n\u03c9\n1 2\nVector control is separate control of the then we require a procedure to compute re flux generative components and stator curren flux angle, which this procedure is perfo current breaker circuit according to above eq advantage of rotor flux control is its simp component breaker circuit ,linearity of torqu and fast response of torque .In indirect con speed of synchrony r\u03c9 is obtained by increa disjoins from current breaker circuit and m\u03c9 can be measured directly or be derived from r\n( ) dt\nds m rr \u03b8\u03c9\u03c9 += *\nFinally r\u03c9 is integrated for computing of is also used for converting of two-phase cu phase currents.\n\u23a2 \u23a2 \u23a2 \u23a2 \u23a3\n\u23a1\n\u2212\u2212\u2212\u2212\n\u2212\u2212\u2212 \u2212\n= \u23a5 \u23a5 \u23a5 \u23a6\n\u23a4\n\u23a2 \u23a2 \u23a2 \u23a3\n\u23a1\n*\n*\n*\nsin()cos(\nsin()cos( sincos\nrr\nrr\nr\nc\nb\na r i i i\n2 3 2\n2 3 2\n\u03c0\u03c1\u03c0\u03c1\n\u03c0\u03c1\u03c0\u03c1 \u03c1\u03c1\n\u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a6\n\u23a4\n\u2212\u2212\u2212 +++\n\u2212\u2212\u2212 +\u2212\u2212\n+++\nrrxsyryxsm\nrysrrxrrssymssxrm\nryrrsrxsrsyrmssxm\nrymrrxrrmsyrssxrm\nryrrmrxmrsyrmsxrs DiiiiJPL iLRiLLiLRi iLLiLRiLLiL iLRiLLiLRi iLLiLRiLiL \u03c9 \u03c9\u03c9 \u03c9\u03c9 \u03c9\u03c9 \u03c9\u03c9 )(2/3 2\n2\nective flux for static in rotor connective\nflux and torque, ference values of t torque and rotor rmed through a uations. The real licity for current e character-speed trol method, the sing ( )*rs\u03c9 which is obtained that\notor position:\nrotor flux which rrents into three-\n(10) \u23a5 \u23a5 \u23a6\n\u23a4\n\u23a2 \u23a2 \u23a3\n\u23a1 \u23a5 \u23a5 \u23a5 \u23a5 \u23a6\n\u23a4\n*\n*\n)\n) q\nd i i\n3\n3\nIII. SIMULATION AND\nPosition controllers and speed de stable as base speed refers to the sy that, as the speed increase inv correspondingly. The results of sim motor with the specifications given i\n(9)\n\u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a5 \u23a6\n\u23a4\n\u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a2 \u23a3\n\u23a1\n\u2212 \u2212 \u2212+\nload\nsym\nsxm\nsyr\nsxr\nT uL uL\nuL uL\nITS RESULTS rive from PI type. Flux is nchronous speed and after ersely, it will decrease\nulation of an induction n Table, 1.\nof simulation.\nrol independently of each other.\n(8)", "Figure 4. Three-phase stator currents during the tran performance.\nFigure 5. Vertical and direct component for stator refe\nsient to steady state\nrence current.\nFigure 6. stator current vertical compon changes proportionate\nFigure 8. The response of motor into po degree angle\nThe settling time and percen response and motor speed are deter of position and speed.\nIV. CONCLUS Using vector control, can obtain\nexcited DC motor from AC mac Disadvantages of DC motor such as work constraints in speed\nent after transient state which for torque.\nltage to the motor stator\nsition input for rotation in 120 .\nfor reaching a desired position.\nt overshoot for position mined through controllers\nION the response of separately hine, while we see no requiring periodic repair,\nand high voltage." ] }, { "image_filename": "designv11_101_0003445_icgt.2012.6477977-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003445_icgt.2012.6477977-Figure2-1.png", "caption": "Fig. 2 DC equivalent circuit of source, motor, and dc injection [22]", "texts": [ " The equations are modified and expressed in the form required for the observer, as p \u03bb\u03bb = VV \u2212 R + \u03c3L p 00 R + \u03c3L p ii (11) p \u03bb\u03bb = \u22121 T \u2212\u2375\u2212\u2375 \u22121 T \u03bb\u03bb + ii (12) Based on (11) and (12) speed observer is being developed which calculates the speed by (13) \u2375 = \u2212 ( ) \u2212 \u2212 (13) The stator resistance is estimated from the input line voltages and phase currents using an on-line DC signal injection method. The dc signal injection method is practical; in particular the use of a simple MOSFET controlled circuit to intermittently inject a controllable dc bias into the motor. The impedance of an induction motor when a dc input is injected in steady state is Rs. The circuit structure is shown in Fig. 2. This method has low power dissipation and torque distortion, and is capable of providing accurate stator resistance estimate under motor startup, load variation, and abnormal cooling conditions. The installation of the injection circuit can be easily done in the motor control center for mains-fed machines. The losses are estimated from empirical values using only motor nameplate data. The friction and windage loss is 1.2% of the rated output power; and the stray-load loss is estimated from the recommended values in IEEE standard 112" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001636_icmtma.2011.837-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001636_icmtma.2011.837-Figure1-1.png", "caption": "Figure 1: The possible architectures of fully decoupled spherical parallel mechanisms", "texts": [ " According to the number of constraint, the feasible kinematic chains of spherical parallel mechanism could be divided into four types: the 0-constraint kinematic chain, the single-constraint kinematic chain, the two-constraint kinematic chain and the three-constraint kinematic chain. The combination of 3-Legged parallel wrist mechanism and legs for parallel wrist mechanism are shown in Table1 and Table 2. Where 'R ''R denotes the revolute joints which are parallel to the revolute joint R . The fully decoupled SPM which limbs synthesized perpendicularly are obtained shown in Figure 1. The SPM in figure 1-a, 1-b, 1-c are symmetrical and SPM in figure 1- d, 1-e are non-symmetrical. Since each RRR limb affords three constraints 1 2 3, ,f f f , the 3 RRR\u2212 parallel mechanism (figure 1-a) is over-constrained, whereas the spherical parallel mechanism in figure 1-b, 1-c are also overconstrained, as shown in Table 3 and Table 4. IV. CONCLUSIONS In this paper, a family of fully decoupled spherical parallel mechanisms is presented. The parallel mechanisms have three independent rotational DOFs. The manipulator Table 3 Generated architectures of 3-Legged fully decoupled parallel wrist mechanism Architect ures Constraints Typesymmetr ical TypeNonconstrai ned The 1st kinema tic chain The 2st kinema tic chain The 3st kinema tic chain 1-a 1 2 3, ,f f f 1 2 3, ,f f f 1 2 3, ,f f f 1-b 1f 2f 3f 1-c 1f 2f 3f 1-d 0 0 1 2 3, ,f f f 1-e 1f 2f 3f V" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001318_kem.572.351-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001318_kem.572.351-Figure2-1.png", "caption": "Fig. 2 Concept of hypoid gear hobbing", "texts": [ " Appropriate modifications of existing basic manufacture parameters can significantly enhance the performance characteristics of a gear drive. For this reason, the following manufacture parameters are taken as the basis of the proposed optimization formulation: the radii of the head-cutter blade profile ( 1profr and 2profr , Fig. 1), the difference in head-cutter radii for the manufacture of the contacting tooth flanks of the pinion and the gear ( 0tr\u2206 ), the tilt (\u03ba ) and swivel ( \u00b5 ) angles of the cutter spin- dle with respect to the cradle rotation axis (Fig. 2), the tilt distance ( dh ), the variation in the radial machine tool setting ( e\u2206 , Fig.2), and the variation in the ratio of roll in the generation of pinion tooth-surface ( 1gi\u2206 ). Therefore, the maximum tooth contact pressure and the maximum transmis- sion error depend on the 8 manufacture parameters: ( ) ( )1gd0t2prof1profmaxmax i,e,h,,,r,r,rpp \u2206\u2206\u00b5\u03ba\u2206=mp (1) ( ) ( )1gd0t2prof1profmax2max2 i,e,h,,,r,r,r \u2206\u2206\u00b5\u03ba\u2206\u03c6\u2206\u03c6\u2206 =mp Objective Function and Constraints. As pointed out earlier, the goal is to minimize tooth contact pressure and transmission errors while keeping the loaded contact pattern inside the possible contact area defined by load distribution calculation and inside the physical tooth boundaries of the pinion and the gear" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001414_cobep.2013.6785219-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001414_cobep.2013.6785219-Figure1-1.png", "caption": "Fig. 1. Proposed linear", "texts": [ " The to the same linear velocity and acce inherent in systems converting rotary to l With the linear stepper motor as propo movement along of the x-axis, in the obtained from a single traction driver, mechanical system converters like belts The mover part has a ferromagnetic massive steel 1020 U-shaper where are permanents magnetic (PM) mounted responsible by the magnetic excitation f comprehends the static ferromagnetic core where are fixed six independent c each that will receive the DC excitation system. Figure 1 shows the proposed stru 1 shows the main characteristics. Th development and (LSM), brushless, agnetic static core, lated coils, electric unted. The mover y a ferromagnetic nent magnets are llows the mover to keep the air gap rough the correctly ve control system, lectric current. A of finite element on of magnetic flux as function of DC coil. With the ssible to obtain the ive study between tor, magnetic field, ce, stepper motor. processes require ieved with stepper in the x-axis, in a pes of devices are static and a moving ugh digital control xcellent solution for id acceleration and yloads", " A traction force, due to the quantities, will be analyzed. TA Main charac Quantity Number of PM Number of coils Number of turns per coil Ferromagnetic core material Thickness of the air gap Thickness of the PM Thickness of the coil Front area of the PM Range of excitation current II. PRINCIPLES The operating principles by the movement of the secondary, is based on the machine. The permanent magnet production of the magnetic each other by a uniform air-gap over the secondary is possible tructure, where a system with unted, also presented in Fig. 1. correct movement line by the tic flux density in the air gap s function of the field from PM lso the behavior of the linear interaction between the two stepper motor and its parts. BLE I teristic of LSM Value 2 6 250 Massive steel 1020 7mm 8mm 6mm 25 x 25 mm2 0 \u2013 2 A OF OPERATION of LSM, which is characterized primary above and along the same principle of a step rotary ics are responsible for the field excitation that cross the 978-1-4799-0272-9/13/$31.00 \u00a92013 IEEE 884 coils, with N turns. According to Lorentz`s Law, when a conductor, with length L, is excited with electrical current, I, and it is immersed in a magnetic field , an electromagnetic force appeared on; this force, directly proportional by current and field is given by (1) [4, 6]" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003069_978-1-4419-9305-2_30-Figure6-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003069_978-1-4419-9305-2_30-Figure6-1.png", "caption": "Figure 6: The Finite element model of the UNSW gearbox casing", "texts": [ " This still gave a valid simulation of the gearbox for the purpose of studying its behavior for a spalled bearing (envelope signal for demodulation of a high frequency resonance) and also in studying the different interactions that exist in the system by comparing simulations with real measurements, for a variety of localized and extended faults in both gears and bearings [1, 2]. It was noticed however that localized faults gave better results than extended faults when compared to the experimental ones. More details about the LPM model and the bearing model can be found in [1]. The finite element model (FEM) of the casing (104 340 degrees of freedom) is shown in figure 6. The casing is supported by rubber pads, which are simulated using spring elements at the corners of the casing. The earlier model used only shell elements [5, 6] and was updated to the new one which has both solid and shell elements. The model has been compared with experimental modal testing and validated for the lower frequency modes [6]. In the current update the nodes on the hub of each bearing are connected to a centre node using rigid body elements. Thus one centre node is formed at the centre of each bearing, which will eventually capture the flexibility of the casing" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001734_iros.2011.6094411-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001734_iros.2011.6094411-Figure2-1.png", "caption": "Fig. 2: Principle of the Mechanism", "texts": [ "00 \u00a92011 IEEE 4048 CONTINUOUSLY VARIABLE TRANSMISSION MECHANISMREVIEW STAGE The continuously variable transmission (CVT) mechanism proposed in this study comprise of (a) spherical drive, (b) drive axis to rotate it, (c) variable motor housing with the passive rotational axis to vary the inclination angle of the active rotational axis, (d) fixed bracket to support these and (e) linear sliding plate as shown in Figure 1. The passive rotational axis for the variable motor housing goes through the center of the spherical drive. Rotary motion is converted into linear motion as the spherical drive rotates and pushes out the linear sliding plate. Next, the principle for CVT is described. Figure 2 shows the cross-section drawing of a transmitter. Defining that the radius of the spherical drive is r and the inclination angle for the active rotational axis is \u03b8, suppose that the \u03b8 value when the active rotational axis is in horizontal position against the linear sliding plate is 0 [deg]. When \u03b8 = 0 [deg] (Figure 2(a)), the relative radius r\u2019 value becomes maximum. When \u03b8 > 0 [deg] (Figure 2(b)), the relationship between relative radius r\u2019 and r becomes r\u2019 = r cos \u03b8. Here, when torque T [Nm] is input into the active rotational axis, the linear sliding force on the linear sliding plane F [N] is expressed as follows; F = = T T r\u2019 rcos (1) That is, the linear force F becomes smallest when \u03b8 = 0 [deg] and infinite when \u03b8 = 90 [deg], and CVT becomes possible by varying the inclination angle of the active rotational axis. are sent to the corresponding author only. To enable CVT in the proposed mechanism, the inclination angle \u03b8 for the active rotational axis of the spherical drive needs to be varied" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002317_978-94-007-4201-7_5-Figure5.6-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002317_978-94-007-4201-7_5-Figure5.6-1.png", "caption": "Fig. 5.6 A 3-R(CRR) PM (a) 3-R(CRR) PM (b) a R(CRR) limb", "texts": [ "97) \u20acq \u00bc GH q h i AH _qT HH q h i _q (5.98) where HH q h i\u00f0i\u00de 2 R6 6 6 denotes the virtual second-order synthesis KIC matrix of PM. Hq H\u00bd 2 R6 6 6 is similar to Hq H\u00bd . If function y \u00bc f \u00f0x\u00de is correct, then its inverse function x \u00bc f 1\u00f0y\u00de is also correct as long as the inverse function exists. Therefore, if Eq. (5.91) is correct and matrix GH \u2019 h i is reversible, then Eq. (5.92) is undoubtedly correct. This rule also applies in Eqs. (5.93), (5.94), (5.95), (5.96), (5.97), and (5.98). Example 5.1. As shown in Fig. 5.6a, the 3-R(CRR) chain is a 5-DOF PM with three revolute and two translate freedoms. The base and moving platforms of the 3-R(CRR) mechanism are connected by three limbs, each with three revolute joints and one cylindrical pair. Both the upper and lower platforms are equilateral triangles. Each RCRR limb can be represented by five single-DOF pairs, as shown in Fig. 5.6b. The cylindrical pair can be replaced by one revolute joint and a coaxial prismatic pair. The first joints in the three limbs are perpendicular to the base platform. All other axes of the pairs intersect at one point called rotation center O. As a lower-mobility PM, the Jacobian andHessianmatrices for the 3-RCRRare not square or cubic, which consequently adds obstacles in achieving the reverse kinematic modeling. To obtain the accurate solutions, VMP is adopted in the present example for forward/reverse velocity and acceleration analyses" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001032_s11668-011-9448-x-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001032_s11668-011-9448-x-Figure3-1.png", "caption": "Fig. 3 Schematic of the crankshaft timing belt assembly, timing belt pulley bolt, and crankshaft", "texts": [ " This problem has been associated with 1990 or 1991 Miata model years with a VIN of 209446 or less. The Mazda MiataTM is a two-seat convertible, utilizing a 1.6-liter double-overhead cam engine. A schematic of the engine is shown in Fig. 1, showing the location of the crankshaft nosepiece, timing belt, and timing belt assembly. An example of the as-installed appearance in the car is shown in Fig. 2. The schematic of the crankshaft timing belt assembly, timing belt pulley bolt, and crankshaft is shown in Fig. 3. A failed crankshaft nosepiece and failed timing belt pulley bolt assembly were submitted for analysis. In addition, a reference timing belt pulley bolt was submitted for reference. The failed crankshaft and failed pulley bolt were reported to have failed approximately 5000 km after the timing belt was changed. No maintenance records are available to confirm this. The as-received timing belt assembly is shown in Fig. 4. A closer view of the failed pulley assembly is shown in Fig. 5, showing the relationship between the pulley assembly and the crankshaft nosepiece" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000195_9780857094537.13.799-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000195_9780857094537.13.799-Figure1-1.png", "caption": "Fig. 1 Arrangement of typical geared compressor train", "texts": [ " The results show that reduced model properly inherits selected vibration modes from original DOF model. For energy-saving purpose, the growing number of conventional turbine-driven (both steam turbines) rotating machineries are replaced by electrical motor-driven ones in recent years. As AC motors are the mainstream of electrical motors in industrial applications and they are used up to or less than 3600 rpm in many cases, load machineries such as compressors are most likely combined with speed increaser gears, as shown in Fig. 1. As widely understood, in geared rotor systems, lateral and torsional vibrations are strongly coupled and the coupled vibrations could lead to problems. Since there were many cases of problems of torsional vibrations in actual application detected by lateral vibration of gear shafts, investigations for such problems were made by measurements and calculations consistently for many years. (1) For practical purpose, accurate prediction of lateral-torsional coupled vibration for geared rotor systems is of much importance" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001431_amr.694-697.512-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001431_amr.694-697.512-Figure3-1.png", "caption": "Fig. 3 End section of involute worm", "texts": [ " 2, it is the coordinate system of laevo rotatory ZI-type cylinder worm. For any point N on the tooth surface, it satisfies the following equations [4] 1 1 + +O N O K KM MN= (2) ( ) ( ) 1 1 1 1 1 1 1 cos -sin , cos sin +cos - sin b b b O K r i j KM p k MN u i j u k \u03b8 \u03b8 \u03b8 \u03bb \u03b8 \u03b8 \u03bb = = = (3) In the equation (3), br is the cylindrical worm base circle radius, b\u03bb is the lead angle of the helix, =r tanb bp \u03bb\u2217 is the helical parameter, variables u and \u03b8 are the parameters of the tooth surface. Shown in Fig. 3, in the end section of 1=0z , axis 1x is the symmetry axis. It can be drawn from the Fig. 3 that = - 2 t t p inv r \u03c9 \u00b5 \u03b1 .where, t\u03c9 is the alveolar width on the pitch cylinder as well as within the end section. t\u03b1 is the profile angle of the end section: = arccos b t p r r \u03b1 . From involute trigonometric relationship, = tan -t t tinv\u03b1 \u03b1 \u03b1 .By calculating, the final expression of laevo rotatory cylindrical worm flank I: 11 11 1 11 1 11 1 11 11 11 =x +y +z x = cos ( + )+ cos sin ( + ) y =- sin ( + )+ cos cos ( + ) z =- sin +p b b b b b r i j k r u r u u \u03b8 \u00b5 \u03bb \u03b8 \u00b5 \u03b8 \u00b5 \u03bb \u03b8 \u00b5 \u03bb \u03b8 (4) For the same reason, the expression of flank : 12 12 1 12 1 12 1 12 12 12 =x +y +z x = cos ( + )+ cos sin ( + ) y = sin ( + )- cos cos ( + ) z = sin -p b b b b b r i j k r u r u u \u03b8 \u00b5 \u03bb \u03b8 \u00b5 \u03b8 \u00b5 \u03bb \u03b8 \u00b5 \u03bb \u03b8 (5) Unit normal vector equation of any point on the ZI-type worm tooth surface" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000451_j.proeng.2011.08.545-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000451_j.proeng.2011.08.545-Figure2-1.png", "caption": "Fig. 2. The angle between the directions of dF and the horizontal plane", "texts": [], "surrounding_texts": [ "The first problem was the direction of the ampere force of the current-element ld2I . The second problem was the projection of the ampere force in the horizontal plane of ACGHA. Because the coil plane KCH was perpendicular to the horizontal plane of ACGHA, the angle between the projection in the horizontal plane of the current-element ld2I and the direction of its magnetic induction intensity was 1\u03b1 in the top view shown in figure 1(c) and \u03b8\u03b2\u03b1 --2/ 11 \u03c0= . The difficult was how to get the angle 2\u03b1 . The angle 2\u03b1 was the angle between the direction of the ampere force and the horizontal plane. The method was that the current-element was as an origin point. The extension line of AH was as x-axis. The direction of the magnetic induction intensity of the current-element was as y-axis. The upward direction was as z-axis. The result was \u03d5\u03d5 cos/sin 2RCK = and \u03d5sinRCH = . The projection of CH in the direction of x-axis, the result was as follow. )cos(sin)(sin 11 \u03b8\u03b2\u03d5\u03b1 +=\u22c5 RCH The projection of CH in the direction of y-axis, the result was as follow. )sin(sin)cos( 11 \u03b8\u03b2\u03d5\u03b1 +\u2212=\u2212\u22c5 RCH The results were )0,,0( 1bHG = and )cos/sin),sin(sin),cos(sin( 2 11 \u03d5\u03d5\u03b8\u03b2\u03d5\u03b8\u03b2\u03d5 RRRHK +\u2212+= The vector direction of dF was as ),,( 111 zyxn = 9 , then the relations were 0=\u22c5 nHG 9 , 0=\u22c5 nHK 9 and \u23aa\u23a9 \u23aa \u23a8 \u23a7 =\u22c5++\u2212\u22c5++\u22c5 =\u22c5 0 cos sin ))sin(sin()cos(sin 0 2 11111 11 \u03d5 \u03d5\u03b8\u03b2\u03d5\u03b8\u03b2\u03d5 R zRyRx by The results were 01 =y and 0cos/sin)cos(sin 2 111 =\u22c5++\u22c5 \u03d5\u03d5\u03b8\u03b2\u03d5 RzRx . It was assumed 11 =x , then \u03d5 \u03b8\u03b2\u03d5 sin )cos(cos 1 1 +\u2212=z . The result was obtained as follow. \u03d5 \u03b8\u03b2\u03d5\u03b1 sin )cos(cos tan 1 2 += (3) The element ampere force of the current-element ld2I and its contribution in the horizontal direction and along the HA direction was as formula (4). \u03b1\u03bc sin 2 dd 1 10 2 \u22c5 \u03c0 \u22c5= r I lIF , 2 1 10 2 cossin 2 dd \u03b1\u03b1\u03bc \u22c5\u22c5 \u03c0 \u22c5= r I lIFh (4) The results were from equation (1). \u03b8\u03d5\u03d5 \u03b8\u03b8\u03b2 cossin2sin sin )sin( 222 1 bRRb b \u2212+ =+ , \u03b8\u03d5\u03d5 \u03d5\u03b8\u03d5\u03b8\u03b1 cossin2sin sinsin)sincos( sin 222 2222 bRRb bRb \u2212+ +\u2212= , \u03d5\u03b8\u03d5\u03b8 \u03b8\u03d5\u03d5\u03d5\u03b1 2222 222 2 sinsin)sincos( cossin2sinsin cos bRb bRRb +\u2212 \u2212+ = . Where it was as \u03d5dd Rl = and \u03b8\u03d5\u03d5 cossin2sin222 1 bRRbr \u2212+= . The total force in horizontal direction and the total torque were as follow. \u222b \u03c0 \u2212+\u03c0 = 2 0 222 10 21 cossin2sin sin 2 d \u03b8\u03d5\u03d5 \u03d5\u03bc\u03d5 bRRb I RIF h (5) \u222b \u03c0 \u2212+\u03c0 = 2 0 222 22 210 d )cossin2sin(2 sinsin \u03d5 \u03b8\u03d5\u03d5 \u03b8\u03d5\u03bc bRRb bRII M (6) Because the direction of torque was opposite to the direction of the angle \u03b8 increase, the work of the ampere force when \u03b8 was from 0 to 2/\u03c0 as formula (7). \u222b \u03c0/ \u2212+ + \u03c0 \u2212= 2 0 222 222 210 d)sin sin2sin sin (ln 4 \u03d5\u03d5 \u03d5\u03d5 \u03d5\u03bc bRRb RbRII A (7)" ] }, { "image_filename": "designv11_101_0000649_amr.631-632.971-Figure9-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000649_amr.631-632.971-Figure9-1.png", "caption": "Fig. 9 Contact Stress Vector diagram of Meshing Teeth Pairs\uff08 1z =99\uff0c 2z =100\uff09", "texts": [ " As the meshing continues, the meshing point will move toward the tooth root of external gear and the tooth top of internal gear until the tooth top of external gear contacts with the tooth root of internal gear and then quits from meshing. 2. At meshing-in end and meshing-out end, the tooth top and tooth root of gear meshes, so great contact stress will be formed, which causes the contact stress of two pairs of teeth to reach their peak value. This phenomenon can be clearly understood through observing the vector diagram of contact stress, as shown in Figure 9. Due to great contact stress, the contact stress of external teeth pairs of the two pairs of teeth is very small, indicating only slight contact occurs. 3. At meshing-in end and meshing-out end, the tooth top and tooth root of gear meshes, the internal gear of meshing-in end and external gear of meshing-out end will also bear great bending stress (usually not the maximum value). This phenomenon can be clearly understood through observing bending stress could chart and bending deformation graph, as shown in Fig.10 and Fig.11 respectively. Analysis of the Number of Meshing Teeth Pairs of Internal Gear Pairs with Small Tooth Number Difference of stub gear Make a comprehensive analysis by combining three kinds of stress broken line graphs, contact stress vector diagram, bending stress cloud chart, bending stress deformation graph, the number of apparently deformed meshing teeth pairs can be determined. For example, observing FIg. 3 (b) , FIg. 9,FIg. 10 and FIg.11, the number of apparently deformed meshing teeth pairs can be determined as 6 pairs when 1z =99, 2z =100. This article names the apparently deformed teeth pairs undertaking main load as effective meshing teeth pairs. According to calculation, when the tooth number of internal gear 2z =40-150 and tooth number difference 12 zz \u2212 =1, the number of effective meshing teeth pairs is 2-10 pairs; as shown in Table 2. when the tooth number difference 12 zz \u2212 in increased to 5, the number of effective meshing teeth pairs will reduce accordingly and changed into 2-8 pairs, as shown in Table 3" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001676_esda2012-82491-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001676_esda2012-82491-Figure3-1.png", "caption": "Figure 3. CYLINDRICAL ROLLERS.", "texts": [ " Recently, Chaise and Nelias [20] found that the way to attain the final state may be different when the load is moving compared to purely vertical loading. ELASTIC ANALYSIS Cylindrical roller bearing A NJ2232 roller bearing has been selected for this study: - inner ring raceway diameter di=228 mm, - outer ring raceway diameter de=308 mm, - roller diameter Dw=40 mm, - total roller length Lw=65 mm, - roller\u2019s end chamfer R2=1.2 mm, - number of cylindrical roller Z=17, - internal clearance S=0.15 mm. Two roller profile designs were involved: - straight line profile with end chamfer, Fig. 3a, and - cylindrical-crowned profile with end chamfer, Fig. 3b. Class I discontinuities exist at the intersection points of roller profile: points A and B in Fig. 3a and Fig. 3b, respectively. A radial load of 450 kN, a inner ring rotation speed of 1000 rpm and bath oil lubrication were considered as main working conditions. The pressure distributions obtained for pure elastic material, exemplified in Fig. 4 and Fig. 5, exhibit sharp peaks where discontinuities exist. If considered in life calculations these peaks make smaller the basic reference rating life of the bearing, [1], [3]. On the other hand these sharp pressure peaks may overcome, locally, the yield condition and cause plastic deformations able to attenuate the detrimental effect of geometrical discontinuities" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002915_6.2011-6350-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002915_6.2011-6350-Figure1-1.png", "caption": "Figure 1. Geometry of reference frames", "texts": [ " In Section II, the rotational equations of motion of a satellite are derived along with a detailed derivation of a state space model. The control system design is discussed in Section III and the parameters of a typical satellite mission are listed in Section IV to validate the controllability and stability of the designed controller. Further, results of numerical simulations are discussed with plots of settling times for comparison with the pyramid con guration. Section V presents the conclusions. II. System Model and Equations of Motion An illustration of the reference frames used in this work is shown in Fig. 1. We consider a micro-satellite in a near-circular low-Earth orbit. The orbital reference frame is de ned such that, ZO points towards the Earth\u2019s center and XO is parallel to the tangential velocity vector and YO completes the right-hand system. The satellite body axis is xed with respect to the inertial axis, the origin of which is assumed to conincide with the geometric center of the satellite. Two actuator con gurations - pyramid mounted with 2 of 12 American Institute of Aeronautics and Astronautics D ow nl oa de d by N A N Y A N G T E C H N IC A L U N IV E R SI T Y o n O ct ob er 1 5, 2 01 5 | h ttp :// ar c" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000547_embc.2013.6611225-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000547_embc.2013.6611225-Figure1-1.png", "caption": "Fig. 1. SLIP model. A single step is defined by a sequence of single a double support phases and it begins when the CoM position is at the apex height. The step ends when the same condition is reached.", "texts": [ " This definition could lead to different results if different definition of the leg\u2019s length are used [8]. This paper presents a different approach of calculating leg stiffness using an extended kalman filter (EKF). The formulation is based on the equations of motion of the SLIP model. The CoM trajectory is considered as the measurement information. Simulation results for different walking solutions using different stiffness values are presented in order to validate the filter performance. According to the SLIP model (see Figure 1), a walking pattern can be generated by two massless legs represented by two springs with fixed rest length L0 and stiffness kleg. The body is represented by a point mass m located in the body center of mass r = [ xCoM yCoM ]T . During stance, the forces F1,2 are generated from the springs directed from the foot points rFP = [ xFP yFP ]T to the body mass m. The swing leg doesn\u2019t affect the dynamics of the whole system since it\u2019s massless and generates no force. A single step is defined by a single support of one leg, followed by a double support phase, and finally a single support of the leg which was initially in a swing phase" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003715_1.5062939-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003715_1.5062939-Figure1-1.png", "caption": "Figure 1 Both distinct phases of a laser welding: the interaction phase and the cooling stage", "texts": [ " Consequently, Lagrangian or semi-Lagrangian After a short description of the main physical phenomena, physical and numerical aspects are described. Results obtained by both approaches are then discussed and concluding remarks are exposed. During the interaction between the laser and the material, temperature greatly increases in the welding area and a melted zone is formed. When the boiling point is reached, the ejected vapor induces a pressure called the recoil pressure which inserts the liquid-gas interface as a piston: it tends to form a deep and narrow cavity called the \u201ckeyhole\u201d (Figure 1). Vapor and ambient gas coexist in a complex media influencing the energy distribution. At the end of the interaction, the vaporization process stops and surface tension forces provoke the keyhole collapse (Figure 1). For significant interface deformations gas bubbles are trapped into the melting pool. If the solidification time is insufficient these bubbles give birth to residual porosities [7]. In our study, we focus our attention on an isolated impact made on a Ti6Al4V sheet. The numerical results will be compared with experimental tests [8], especially metallographic data obtained with a fixed power and different pulse durations (Figure 2). Main process parameters are reported in Table 1. The properties of the alloy have been previously characterized by many authors [9,10]" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001710_ipemc.2012.6259255-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001710_ipemc.2012.6259255-Figure3-1.png", "caption": "Figure 3 Phasor diagram of VVVF control in steady status [V y ] =", "texts": [ "The improved VVVF control can keep the current reference constant with light loads and dynamically adjust the 978-1-4577-2088-8/11/$26.00 mOl21EEE reference value s according to working conditions with heavy loads. It can also provide a good solution to the compensation of voltage drop for stator resistors to keep the magnetic field constant. [R + pLy meLii ] [ \ufffdy ] + [1f/1me c\ufffds rp] (1) Vii -meLy R+pLii Iii If/lmesm rp When in steady status, the VVVF control can be analyzed in phasor diagram as shown in Figure 3. VVVF control is located in o-y coordinate system with the phase difference

0 Limit switches are common types of mechanical switches used in mechatronic applications. These switches can be pushed by a button or lever operated mechanism. Figure 3.92 shows the schematic diagram of single pole single through (SPST) device that opens or closes a single connection. In addition, a switch locking mechanism may be used. A pole is a moving element in the switch that makes or breaks the connection. The through is a contact point for the pole. Figure 3.93 shows the bond graph for the SPST switch. Here Df element is current detector which detects the current once the circuit is closed. The constitutive relation for the R element in the bond graph is given as R = { Rhigh, if x < xlim Rlow, if x xlim (3.97) where Rhigh represents the resistance of the air gap. I:m A proximity sensor consists of an element that changes the state of its own or of an analogue signal when it is in proximity of an object. A photo emitter (LED), photo detector (photo resistor) pair can be used as a proximity sensor" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002083_ifost.2011.6021074-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002083_ifost.2011.6021074-Figure1-1.png", "caption": "Figure 1. The control vector diagram by the side of motor.", "texts": [], "surrounding_texts": [ "suitable for transmitting power to the grid, is proposed in\nallusion to the unstable characteristic of the output power. This\ncontrol strategy captures wind power from a vertical axis wind turbine(V A WT), and a back-to-back PWM converter is applied\nto control the output power from permanent magnet\nsynchronous motor(PMSM) in order to make sure the wind\npower can be transmitted to the grid safely. Due to the characteristic of the traditional control method that the power\nfactor is not the optimal by the side of motor, this paper presents\na control method of unit power factor, so that the voltage and the\ncurrent are controlled to the same phase by the side of motor as well as the voltage and the current are controlled to the opposite\nphase by the side of power grid in order to ensure the output\nreactive power is zero. The simulation results based on Simulink\nvalidates the effectiveness of the proposed control strategy.\nKeywords-wind energy; permanent magnet synchronous motor; Vertical axis wind turbine; control strategy.\nI. INTRODUCTION\nThe quantity of coal, petroleum and other non-renewable resources is less and less in the modern world, so that there needs to be some research on the new sources of energy to replace them as soon as possible. As a kind of clean resources, wind energy does not pollute the environment which attracted more and more attention. Although the development of horizontal axis wind turbine is used wider now, the V A WT has many advantages such as a simple structure and repairing conveniently which can be applied in many places. Comparing with the asynchronous machine, the PMSM as generator has many advantages of without a growth gearbox and field winding, which has good prospects for development.\nHowever, the output electric energy of the generator, which is not stable, can't connect with the power grid directly. So the electric energy must be regulated in order to meet the conditions of power transmitted to the grid. Ref.[l] does researches on the power transmitted to the grid control strategy of a direct-driven permanent magnet synchronous wind power generator. It uses a method of rotor flux oriented vector control to regulate the PMSM, and the flux direction of the generator is seen as the direction of d-axis. And then, it controls the d-axis current to 0, so that the optimal wind energy can be captured by controlling the q-axis current. In this case, the power factor of the PMSM is less than 1 as well as the d-axis voltage is not O. That means this control strategy will produce reactive power and increase the motor capacity. Ref.[2] applies a control method of uncontrollable rectifier and\n978-1-4577-0399-7/11/$26.00 \u00a92011lEEE 508\ncontrollable inverter as the circuit of power transmitted to the grid which has a lower cost of the system and a simpler control algorithms. However it can not regulate the torque of the generator directly, meanwhile it will increase the stator harmonic currents of the generator.\nThis paper introduces a control strategy by the side of the PMSM and by the side of the grid in detail. Meanwhile, the VA WT is seen as the object of studying instead of the traditional horizontal axis wind turbine. The simulation results show the feasibility of this control strategy.\nII. THE CONTROL STRATEGY BY THE SIDE OF MOTOR\nFrom the aerodynamic characteristics of VA WT, it can be known that the mechanical output power is given by\nI 3 P,,, = 2. pSv Cp (A) (1)\nwhere p is air density, S is the swept effecti ve area of VA WT, v is the wind speed, Cp(A) is wind-power utilization coefficient and A is tip speed ratio. In the case of the pitch angle does not change, the wind-power utilization coefficient is a function of tip speed ratio.\nAnd\nS=2RH (2) where R is blade radius, H is the height of wind turbine.\nThe mechanical angular velocity of the wind turbine IS\ngiven by\nAV m= - (3)\nR Wind turbine mechanical angular velocity is equal to the\nmechanical angular velocity of generator rotor when the direct-driven permanent magnet synchronous generator is used. When the wind speed is changing, the wind turbine rotate speed is regulated by controlling the generator rotate speed so that the wind turbine has the optimal tip speed ratio and the maximum power coefficient. This regulation can ensure the wind turbine output power is maximum.\nThis paper uses a back-to-back PWM control strategy, and the system is composed by a V AWT, a PMSM, a convertor by the side of motor, DC capacitor, a convertor by the side of grid and the grid. The convertor by the side of motor is controlled by the SVPWM pulse output, which is used for controlling to\nAugust 22-24, 2011", "output the maximum power. The traditional control method supposes the direction of the rotor flux as the direction of d axis, and the q-axis is ahead of the d-axis 90 degrees. The voltage equation of the generator is given by\nU,d = R,id + Ld dt - (f),Lqiq 1 . d i d .\ndi (4)\nU = R i + L - q\n+ (f) L)d + (f) IjI sq s q q dt e e\nwhere Usd is d-axis component of stator voltage, usq is q-axis component of stator voltage, Rs is stator resistance, Ld is equivalent d-axis inductance, Lq is equivalent q-axis inductance, id is d-axis component of stator current, iq is q-axis component of stator current, We is synchronous electrical angular velocity, and If! is the rotor flux of the permanent magnet.\nThe electromagnetic torque equation of the motor is given by\nr; =1.5p[(Ld -Lq)i )q + ljIi q] (5)\nwhere p is the motor pole pairs.\nWhen id is supposed as 0, the equation can be simplified as\nr; = l.5 PIjI i q The mechanical motion equation is given by\ndw J - =T -T\ndt m e\n7;\" =\ufffd\" / (f)\n(6)\n(7)\n(8)\nwhere J is the moment of inertia of the generator and Till is the VA WT mechanical torque.\nIt can be seen that the generator rotate speed is directly related to the electromagnetic torque, and the electromagnetic torque is proportional to the q-axis current. So we can control the motor speed by controlling the q-axis current in order to get the optimal wind energy.\nHowever, this control method has a deficiency that the d axis voltage and q-axis voltage are not O. The generator will produce active power, and produce reactive power at the same time. The reactive power will increase with the increase of active power, which will increase the cost of production and operation of the generator. This paper presents a new control method to do with this problem.\nIn order to make sure the output reactive power is 0, you need to keep the stator voltage integrated vector and the stator current integrated vector in the same phase. This phase can be defined as the direction of q-axis, and the d-axis lags behind q axis 90 degrees. In this case, there is an angle f}.e between the direction of d-axis and the direction of flux linkage. The electrical angle between the direction of flux linkage and the stator phase A can be defined as e. So the angle between the direction of d-axis and the stator phase A is given by\nB' = B - f}.l) (9)\nThe control vector diagram by the side of motor is shown in Fig. I.\nThe e can be obtained by the position of generator rotor, and f}.e is given by\nLi f}. ()= arcsin \ufffd\nIjI (10)\nwhere is is the value of stator current integrated vector. The position of stator current integrated vector is in the q-axis, so it can be known that is=iq\u2022\nBecause of the changing position of the d-axis, the voltage equation of the generator needs to be changed.\n1 di U , = R i d + L, _ d -(f) L i - (f) 11/ sin f}.()\nSt s ( dt e q q e't'\ndi (11)\nU = R i + L - q\n+ (f) Ldi d + (f) 11/ cos f}.() sq s q q dt e e'r\nTherefore, the control strategy by the side of motor is controlling the rotor speed in the outer ring and controlling the current in the inner ring. The outer ring control makes sure the use of wind energy is maximum. Meanwhile, the inner ring control can regulate the electromagnetic torque directly.\nWe set the given rotor speed is W *, the given q-axis component of stator current is iq *.", "III. THE CONTROL STRATEGY BY THE SIDE OF GRID\nThe controlling goal by the side of grid is making sure that the voltage of the DC capacitance is constant. It also needs to ensure that the output reactive power is O. Therefore, the direction of grid voltage integrated vector is taken as the d axis, and the q-axis is ahead of d-axis 90 degrees.\nThe control vector diagram by the side of grid is shown in Fig.3.\ned is d-axis component of the grid voltage, igd is d-axis component of the grid current, and Bg is the angle between the d-axis and the phase A, which is given by\ne () = arccos-'!...\ng e, (12)\nwhere ea is the phase A voltage of the grid, and es is the value of the grid voltage integrated vector. The grid voltage can be known from the system, so that the value of the grid voltage integrated vector is given by\n(13)\nIn this case, we can know that eres, eq=O, and the power equation by the side of grid can be simplified as\n{p = \ufffd (edigd +eqigq) = \ufffdeJgd (14)\nQ 3( . . ) 3 . = \"2 eqlgd -edlgq = -\"2e,lgq where P is the output active power by the side of grid, Q is the output reactive power by the side of grid, eq is q-axis component of the grid voltage, and igq is q-axis component of the grid current.\nIn order to ensure the output reactive power is 0, it needs to make sure that igq=O. It can be found that the direction of the grid current integrated vector is in the d-axis too. So the direction of the grid voltage integrated vector is opposite to the direction of the grid current integrated vector.\nThe voltage equation by the side of grid is given by\nt di u =-Ri -L\ufffd+OJLi +e gd ggd g dt g ggq d\ndi (15)\nu =-Ri -L \ufffd-OJLi gq ggq g dt g ggd\nwhere Ugd is d-axis component of the convertor voltage by the side of grid, Ugq is q-axis component of the convertor voltage by the side of grid, Rg is the resistance by the side of grid, Lg is the inductance by the side of grid, OJg is the grid synchronous electrical angular velocity.\nThe control strategy by the side of grid is controlling the DC voltage in the outer ring and controlling the current in the inner ring. The outer ring control makes sure that the voltage of the DC capacitance is constant. Meanwhile, the inner ring control can ensure the output reactive power is O. The convertor by the side of grid is also controlled by the SYPWM pulse output.\nWe set the given DC voltage is Udc*, and the given d-axis component of the grid current is igd*.\nIV. THE SIMULATION OF THE CONTROLLING SYSTEM\nAccording to the control strategy above, the software of Sirnulink can be used for simulation and research. It is supposed that the air density is 1.225kglm3, and the blade radius is 1.5m. The height of wind turbine is 5m, and the pitch angle is zero. The PMSM parameters are set below, the pole pairs is 2. The stator resistance is 2.8758Q. The stator inductance is 8.5mH. The flux linkage is O.l75Y\u00b7S. The moment of inertia is 0.0006kg\u00b7m2. The inductance by the side" ] }, { "image_filename": "designv11_101_0002010_icmtma.2011.161-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002010_icmtma.2011.161-Figure1-1.png", "caption": "Figure 1. Reduced finite element models of crankshaft and main bearing", "texts": [ " Combined with AVL Excite nonlinear multi-body dynamics software, based on orthogonal design method, the minimum oil film thickness of main bearings under rated working conditions were selected as optimization objective, the main bearing structure parameters and lubricant parameters were optimized and matched. II. MULTI-BODY DYNAMICS MODELLING Elastic crankshaft and main bearing solid models were created in Pro/E software and then entered into ANSYS finite element software for pretreatments. Crankshaft and main bearing were meshed, and oil holes and journal fillets were further remeshed to reduce errors. Substructure method was employed and reduced finite element models were shown in figure 1. Five main degree of freedom (DOF) nodes were defined on crankshaft journal, distributed uniformly in the bearing width and coupled with main DOF nodes of main bearings. Two additional main DOF nodes were on the 4th main journal for thrust bearings. One main DOF node was on crankpin, two on each end of crankshaft and two on shock absorber. A total of 35 main DOF nodes and each node had six degrees of freedom (UX, UY, UZ, ROTX, ROTY, ROTZ). Corresponding to main journal, each of the main bearing shells was defined 5 rows of axial main DOF nodes and each row had 48 circumferential main DOF nodes, a total of 240 main DOF nodes" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001312_imece2011-64349-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001312_imece2011-64349-Figure1-1.png", "caption": "Figure 1. CIRCULAR CYLINDRICAL SHELL: (A) COORDINATE SYSTEM AND DIMENSIONS; (B) 2-D PROJECTION OF THE PANEL MIDDLE SURFACE.", "texts": [ " Different multimodal expansions with up to 37 generalized coordinates associated with natural modes of the panel are used to study the convergence of the solution. This is presented for the first time in studies, where the R-function method involved. The pseudoarclength continuation method and bifurcation analysis are applied to study the nonlinear equations of motion. Numerical responses are obtained in the spectral neighbourhood of the lowest natural frequency. Results are compared to those available in the literature. Internal resonances are also identified and discussed. A circular panel of uniform thickness h with complex base shown in Fig. 1 is considered. The principal lines of curvature of the middle surface coincide with the coordinates x, y of the Cartesian coordinate system, and z is directed along the normal to the middle surface of the shell, as shown in Fig. 1; the origin is placed at the middle of the panel base in order to take advantage of the symmetry of the structure. The displacements of an arbitrary point of coordinates (x, y) on the middle surface of the panel in the axial, circumferential and radial are direction are denoted by u, v and w, respectively; it is assumed that the deflection of the middle surface of the panel is of the same order of the thickness and is taken positive outwards. In the present study, two different nonlinear shell theories, namely the Donnell and the Sanders-Koiter theories retaining in-plane inertia are used", " (10a-c) are inserted into mode shapes (9) and expanded in double series in terms of Chebyshev polynomials, where a single\u2013index sequence of re-ordered unknown coefficients ,i j u , ,i j v and ,i j w becomes following { }31 1 2 2 ( )( ) ( 1) ( ) ( 1)(1) (2) 0,0 0,1 , 1 0,0 , 1 0,0 , 1 , ,.., , , , , , , w w u u v v NN N N N M M M M M M w w w u u v v + + \u2212 \u2212 \u2212= \u2026 \u2026q . (11) In Eqn. (9) ( ), 0x y\u2126 = is the equation of the panel boundary [15, 16]. The complete procedure of building such equation of the domain for the panel shown in Fig. 1 is addressed in references [4, 18] in details. Equations (8, 9, and 10a-c) are to be inserted into the linear expression of the strain energy, Eq. (3), and the kinetic energy (4), respectively. Lagrange equations for free vibrations are 3 d 0, 1,2, , d i i L L i N t q q \u2202 \u2202 \u2212 = = \u2026 \u2202 \u2202 , (12) where s s L T U= \u2212 ; ( ) 3, 1,..., i q f t i N= are the general coordinates, which correspond to the reordered sequences, see Eq. (11), of the unknown coefficients from equations (10a-c). Then, for harmonic linear vibration, i", " In particular, the shell response under harmonic excitation has been studied by using an analysis in two steps: (i) first the excitation frequency has been fixed far enough from resonance and the magnitude of the excitation has been used as bifurcation parameter; the solution has been started at zero force where the solution was the trivial undisturbed configuration of the shell and has been continued up to reach the desired force magnitude; and (ii) when the desired magnitude of excitation has been reached, the solution has been continued by using the excitation frequency as bifurcation parameter. Numerical calculations are performed for the clamped cylindrical panel with the complex base shown in Fig. 1, with the following dimensions and material properties: overall length 0.199 b = m, curvilinear width 0.132 a = m, length of the cut 0.041c = m, curvilinear width at the cut 0.092d = m, radius of curvature 2R = m, thickness 0.00028 h = m, Young modulus 195 E = GPa, mass density 7800 \u03c1 = kg/m 3 and Poisson\u2019s ratio 0.3\u03bd = . The same shell was considered in the reference [18]. The frequency range around the lowest natural frequency is investigated. Figure 2 presents the fundamental frequency \u03c91,1 and the other natural frequencies of interest in this study and corresponding mode shapes of the panel obtained for the Donnell shell theory" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003098_s11771-013-1828-9-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003098_s11771-013-1828-9-Figure1-1.png", "caption": "Fig. 1 Principle and structure of target tracking control system", "texts": [ " Considering the target tracking control system as a typical multi-rate control system, analyze the impact of the low-rate sampling of the outer loop on the tracking performance and put forward the controller design method based on different sampling rates. Furthermore, a sampling rate converter performs rate matching, to ensure that the effective control signal can be output at the non-sampling time. The design method is expected to meet the same performance as a high sampling rate controller. We discussed the control design process in details and performed the simulation and experiment verification. The typical target tracking control system possesses a two-axis gimbal shown in the Fig. 1, an inertia sensor gyro, an optical detector, and other loads are all located on the inner pitching framework. In addition, its two axes are both equipped with the torque motor, encoder, and other assemblies. The machine has two independent motion freedoms. The composition and logic relations of the azimuth and elevation movement control are shown in Fig. 2. In order to improve the anti-jamming capability and response rapidity, the system includes three loops: current loop, inertia stabilization loop, and tracking loop" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001690_icra.2011.5979627-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001690_icra.2011.5979627-Figure1-1.png", "caption": "Fig. 1. Model of telescopic-legged rimless wheel", "texts": [ " Suguro are with the School of Information Science, Japan Advanced Institute of Science and Technology, 1-1 Asahidai, Nomi, Ishikawa 923-1292, Japan {fasano,suguro}@jaist.ac.jp for the high-speed stance-leg extension. The robot should change the gait from walking to running in that case. In this paper, we achieve active dynamic running as an extension of the high-speed level walking. By accelerating the stanceleg actuation, a running gait is naturally generated. Although there are many criteria for running gait [10], we evaluate the gait efficiency in terms of the walking speed and the specific resistance. Fig. 1 shows a model of a planar telescopic-legged rimless wheel. This robot consists of eight telescopic legs whose mass is m [kg], and has a hip mass of mH [kg] at the central position. Each leg has a control force for the telescopic-leg actuation, but only those of the stance-leg, u1 and of the previous one, u2, are treated. Let q \u2208 R 5 be the generalized coordinate vector defined as qT = [ x z \u03b8 L1 L2 ] . (1) (x, z) is the tip position of the stance leg, \u03b8 [rad] is the angular position of the stance leg with respect to vertical, L1 [m] and L2 [m] are the lengths of the stance-leg and the previous stance-leg" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000268_amm.473.39-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000268_amm.473.39-Figure2-1.png", "caption": "Fig 2 Schematic of bidirectional impact of Spring System", "texts": [ " The armature would be attracted to move closer to the iron core by the strong magnetic force\uff0cat the same time, the spring system would rotate around the endpoint of the yoke, then making a circuit. Finally the target of control and change the circuit is achieved. When the relay is exposed to an impact, the contacting- spring would feel an inertia force along the negative direction of the impaction. Upward impact forces the contact spring system downward movement relatively; downward impact causes a trend upward movement of the contact spring system, as shown in Fig 2. Strong impact could cause transient or permanent deformation of the contact spring system. So that strong impact is the key factor of uncontrollable and false triggering of electromagnetic relay in engineering applications. In the launch system of an experiment, the strong impact caused the uncontrollable of electric control system\uff0cand result the virtual connection in electromagnetic relay of the electric control system, which lead to the control system failure. Fig 3 shows the contrast of internal structure of SRA-05VDC-CL electromagnetic relay before and after mechanical impact" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000524_amr.301-303.573-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000524_amr.301-303.573-Figure1-1.png", "caption": "Fig 1 Schematic diagram of the structure of rotary seal", "texts": [ " Basic on the principle of hydraulic resistance network, via establishing mathematical model of each section belonged to flow system, a leakage model of flow system is set up. Using the analytical model and the experimental methods, the factors of leakage are discussed. The results indicate that the characteristics of leakage depend on the working conditions, structural parameters of rotary seal and viscosity-temperature characteristics of lubricating medium. Introduction The rotary seal has been widely applied in vehicle transmission, such as wet clutch [1-2], the geometric model of flow field distribution of which is shown in Fig.1. When the rotary seal is working, there exists frictions between the static and the still ends, therefore, heat is generated and causes temperature raise in film and sealing ring. If the working temperature keeps rising, the lubricating film will vaporize partly and the sealing ring will deform greatly, result in rougher friction and abrasion and shorter longevity [3]. The friction heat is mainly brought away by the lubricating media, and in this system the leakage is directly related to leak power loss. Thereby, a proper design for the flow of the rotary seal system is pretty vital. The geometrical model of the flow field in rotary seal is shown in Fig.1. The flow field is divided into five sections (0, 1, 2, 3, 4) according to the structure. According to the hydraulic resistance theory, every portion of a fluid system can be seen as one hydraulic resistance, which accords to the formula: /R p q= \u2206 (1) In the formula above, q is the flow passing through the portion; \u2206p is the pressure drop of passing before and after. According to the structure, hydraulic resistance network model is shown in Fig.2. p0 is the entering pressure; \u2206p is the pressure drop; q is flow; \u2206p0, \u2206p1, \u2206p2, \u2206p3, \u2206p4 are the pressure drops of each hydraulic resistance; R0, R1, R2, R3, R4 are the hydraulic resistances", "tank 2.temperature sensor 3.loading motor 4.gear pump 5.filter 6.pressure sensor (spillover valve) 7. frequent conversion motor 8.speed and torque sensor 9.input shaft 10.gear transmission 11.pressure sensor (entrance of sealing system) 12.entrance where pressed oil enters test box 13.inlet bushing 14.distributing bushing 15.the sealing test segment 16.rotary sealing ring 17.main shaft 18.exit where leaking oil outflows 19.flowmeter 20.gear accelerating transmission segment 21. Spillover valve With Fig.1, parameters of sealing ring are shown below as Table.1. Tab 1 Structural parameters of sealing ring l0/mm l2/mm h1/mm r0/mm r1/mm r2/mm r3/mm h3/mm S3/mm 3 3.4 0.3 45 47.5 49.7 50 0.5 2 The rotary seal applies 15W/40CD lubricating oil, with viscosity of 0.09Pa\u00b7s when 40\u00b0C and viscosity of 0.02 Pa\u00b7s when 80\u00b0C With the temperature of 80\u00b0C, the testing results of the leakage of the rotary seal compare with the theoretical calculation and simulation, shown in Fig.5 as below. From Fig.5, the results in test and theory accord well, with a appropriate error, which demonstrates that the model is correct" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000174_978-3-642-28768-8_13-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000174_978-3-642-28768-8_13-Figure1-1.png", "caption": "Fig. 1 Combustion engine", "texts": [ " Another parameter that should be taken into account is the transient behaviour of a gear transmission caused by the fluctuation of the rotational speed. Bouchaala N et al (2011), Khabou M T et al (2011) showed that the transient regime influences significantly the dynamic behaviour of the transmission. Therefore the aim of this paper is to use the numerical simulation based on the Newmark integration method, to study the dynamic behavior of spur gear system in the transient regime powered by four strokes four cylinders inline diesel engine. The combustion engine (figure 1) is subject to two types of solicitation, those from the combustion gases and those related to the phenomena of inertia. In a cylinder gases produce a variable pressure in the cycle, which at the moment of combustion, generates an effort called \"effort gases\" on the piston. This effort is the cause of the piston displacement and traction engine. Similarly, the moving parts (piston, rod, ...) undergo efforts called \" inertia effort \". Generalized efforts, which represent the resultant forces and moments when applied to the engine, assess the movements of the latter on his pads and the rotation of the crankshaft" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001128_j.proeng.2012.04.134-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001128_j.proeng.2012.04.134-Figure1-1.png", "caption": "Fig. 1. (a) Definition of the coordinate system for analysis; (b) Definition of the double pendulum model", "texts": [ " The location of the shoulder joint center was determined as the point of bisection of the anterior and the posterior markers, which was orthogonal to a line dropped from the acromion marker. The elbow joint center (x0) was the midpoint between EL and EM. The wrist joint center (x1) was the midpoint between RS and US. Prior to the test, we determined the third metacarpal joint center x2 using vector algebra and measured the hand height (28 mm) at the head of the third metacarpal where the hand marker was placed. It should be noted that 1 is the length of the central forearm (link 1) from x0 to x1; l1 is the reference frame vector normalized to unit length (Fig. 1a). ua is a vector from x0 to the shoulder joint center. 1 is the cross-product of l1 and ua; 1 is the cross-product of 1 and l1. Similarly, 2 is the length of the hand (link 2) from x1 to x2; l2 is the reference frame vector normalized to unit length. 2 represents the cross-product of l2 and a vector from the US to RS, and 2, the cross-product of 2 and l2. 1 and 1 do not necessarily coincide with the velocity vector of x2, because the hand segment rotates about the long axis. Therefore, we determined an additional local reference frame ( q2 ) for analysis. represents l2 l2 , where l2 is the velocity of l2 . The direction of is perpendicular to that of l2. t2 is defined as the cross-product of l2 and . x 1 and x 2 denote the center of masses of link 1 and link 2, respectively (Fig. 1b). Further, m and m denote the masses of the two links, respectively. J1 and J denote the inertia matrix of the two links, respectively. In this study, the ball was regarded as the mass that was applied to link 2. 1 and 2 denote the length between each pivot and the center of mass. Using the measured data and the body segment parameters for Japanese populations reported by Ae et al. [7], we calculated the mass and the moment of inertia of each segment. From the second derivatives of each position vector x 1, x1, and x 2, accelerations 1, 1, and 2 are given by 1 0 1 l1 1 l1 (1) 1 0 1 l1 1 l1 , (2) 2 1 2 l2 2 l2 ", " By differentiating the mechanical energy of the hand, E2 T2 U2, and by substituting (10), we obtain another equation for the power of the hand: 2 1 . (11) It should be noted that the first term of Eq. (11) denotes the power transferred via the internal force, and the second term denotes the power produced/absorbed by the muscle (external force). Eqs. (3), (4), and (11) can be used to obtain the rate of change of the mechanical energy for the hand, 2: 2 t2 l2 q2 , (12) where the internal force 2 and velocity 1 are defined as 2 , , and 1 t2, l2, q2 in a coordinate frame q2 (Fig. 1.) The velocity of the ball at REL (t=1.000 s) was 39.0 m/s. The maximal shoulder external rotation (MER) occurred at t=0.970 s. The maximum value of the relative angle of link 1 and link 2 was obtained at t=0.960 s (44.8\u00b0). This value then gradually decreased, and subsequently, was 14.3\u00b0 at REL. The angular velocity of link 1 ( l1 ) gradually increased to the REL. In contrast, the angular velocity of link 2 ( l2 ) attained a minimum value just before the MER (25.7 rad/s), following which it dramatically increased to the REL" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000636_iceee.2011.6106667-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000636_iceee.2011.6106667-Figure2-1.png", "caption": "Figure 2. High-performance permanent-magnet radial and axial topology synchronous generators. Windings are on the stator, while permanent magnets are on the rotor. This allows one to design turbines with permanent magnets.", "texts": [ " The differential equations of a round-rotor permanent- magnet synchronous generator are [1, 2] dt d dt d dt d mabcs sabcss abcs abcss \u03c8i Lir \u03c8 ir ++=+=0 , ( ) ( ) + \u2212+ +\u2212\u2212 \u2212+\u2212 \u2212\u2212+ = = \u2211 \u2211 \u2211 \u221e = \u2212 \u221e = \u2212 \u221e = \u2212 1 3 212 1 3 212 1 12 2 1 2 1 2 1 2 1 2 1 2 1 sin sin sin n r n n n r n n n r n n m cs bs as mlsmm mmlsm mmmls cs bs as abcs a a a i i i LLLL LLLL LLLL \u03c0\u03b8 \u03c0\u03b8 \u03b8 \u03c8 \u03c8 \u03c8 \u03c8 \u03c8 ( ) ( ) ( ) ( ) , 2 sincos)12( sincos)12( sincos)12( 4 1 3 222 3 2 1 3 222 3 2 22 1 2 PMr m n r n rncs n r n rnbs r n n rnas mr T J P J B ani ani ani J P dt d +\u2212 ++\u2212+ \u2212\u2212\u2212+ \u2212\u2212= \u2211 \u2211 \u2211 \u221e = \u2212 \u221e = \u2212 \u2212 \u221e = \u03c9\u03c0\u03b8\u03c0\u03b8 \u03c0\u03b8\u03c0\u03b8 \u03b8\u03b8 \u03c8\u03c9 d dt r r \u03b8 \u03c9= , (1) where iabcs=[ias ibs ics] T and \u03c8abcs=[\u03c8as \u03c8bs \u03c8cs] T are the currents and flux linkages; d\u03c8m/dt is the motional emf which represents the induced phase voltages uabcs=[uas ubs ucs] T ; \u03c9r and \u03b8r are the electrical angular velocity and displacement; rs is the diagonal matrix of phase resistances; Ls is the symmetric matrix of inductances; \u03c8m is the vector of flux linkages due to phase windings and permanent magnet timevarying and displacement-dependent magnetic coupling; \u03c8m and an are the constants; Bm is the friction coefficient. The images of some examined high-performance axial and radial topology generators are documented in Figure 2 [2]. 2. 2. Power Electronic Module Two- and three-phase ac to dc controlled and uncontrolled rectifiers with a LC output filter are used to rectify the induced ac phase voltages to a dc voltage. To charge a rechargeable battery or super-capacitor, specialized chargers must be used. The input voltage to the charger must be stabilized using a controllable PWM buck-boost converter which are controlled using PI and soft-switching control laws. The voltage can be stabilized in the full operating envelope" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000476_amm.419.795-Figure6-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000476_amm.419.795-Figure6-1.png", "caption": "Fig. 6 Analysis and determination between pattern III and IV", "texts": [ " (4) where \u03b3 is the angle formed by the line crossing from stair vertex to driving wheel's center and the line perpendicular to stair slope line, and \u03b3cr is the angle formed by the line crossing from the tangent point of track and driving wheel to driving wheel's center and the line perpendicular to stair slope line. Pattern II: The stair vertex with the smaller label shown as right subfigure's i in Fig. 5 exert a force on the track while not on the driving wheel directly. The criterion of pattern II can be obtained as: 2 2 2 2 ( )E A cr r r h b i h b l h \u03b3 \u03b3 \u2212 + + + > > . (5) Analysis and determination between pattern III and Pattern IV When 2 2 2 2 ( )E Ar r h b i h b l h \u2212 ++ + \u2264 , there are also two patterns, as shown in Fig. 6. Pattern III: The stair vertex with the larger label shown as the left subfigure's i+1 in Fig. 6 keeps in no contact with the track, and exerts no force on the driving wheel directly. While the stair vertex with the smaller label shown as the left subfigure's i in Fig. 6 exerts a force on the track while not on the driving wheel directly. The criterion of pattern III can be obtained as: 2 2 2 2 ( ) 2 E A E r r h b i h b l h r \u03c0\u03b7 \u03b6 \u2212 ++ + \u2264 > < and . (6) where \u03b7 is the distance from driving wheel's center to stair vertex with the larger label, and \u03b6 is the angle formed by stair slope line and the line crossing from driving wheel's center to stair vertex with the larger label. Pattern IV: The stair vertex with the larger label shown as the right subfigure's i+1 in Fig. 6 keeps in contact with the track, and exerts a force on the driving wheel. While the stair vertex with the smaller label shown as the right subfigure's i in Fig. 6 keeps in no contact with the track. The criterion of pattern IV can be obtained as: 2 2 2 2 ( ) 2 E A E r r h b i h b l h r \u03c0\u03b7 \u03b6 \u2212 ++ + \u2264 \u2264 \u2265 or . (7) If the expression 2 \u03c0\u03b6 < is not satisfied simultaneously for the criterion of pattern III, as is shown in Fig. 7, pattern IV might be mistaken for pattern III. STATEFLOW [4] module in MATLAB is used to simulate the whole process of WT wheelchair robot' stair-climbing, as is shown in Fig. 8, which can realize the function similar to a finite state machine and make a determination for wheelchair's pattern in the stair-climbing process" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001846_amr.538-541.1874-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001846_amr.538-541.1874-Figure1-1.png", "caption": "Fig. 1. The finite element model", "texts": [ " 3 2 (1 1 1 1 ln 1s s sr r f f s f s f f f k k kk kk k k k k k k k k \u03c6 \u03c6 \u03c6 = \u2212 \u2212 + + \u2212 \u2212 + \u2212 \u2212 \uff09 (3) Where \u03c6 is the fractional porosity of the powder bed, kf is the thermal conductivity of the air, ks is the thermal conductivity of the solid and kr is the thermal conductivity portion of the powder bed due to radiation among particles. To simulate the laser cladding processing, a finite element model is developed with ANSYS software using APDL (Ansys Parametric Design Language). \u2018Birth and Death\u2019 Element method has been used for the programming. Calculations were carried out on a block substrate with dimension of 14\u00d78\u00d75mm and a coating thickness 0.5mm. The model meshed in Solid70 element is showed in Fig.1. To guarantee the computational efficiency and simulation accuracy, a fine mesh was used in the area along the centerline and a coarse mesh for the rest of the entity [9]. For the Ni60 powder with a melting point of 950 0 C was selected as the cladding material and 45# steel has been selected as the substrate. The specific heat enthalpy and thermal conductivity coefficient of Ni60 is time-dependent, and given in Table 1 Table.1 The enthalpy and thermal conductivity coefficient Temperature/k Enthalpy/(10 9 kJ/mol) Conductivity 273 0 0" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000900_amr.562-564.654-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000900_amr.562-564.654-Figure1-1.png", "caption": "Fig. 1 Ten-bar linkage", "texts": [ " To overcome those shortcomings, a simple and general formula for mobility calculation of planar mechanism which is expressed by virtual loop constraint is given. The concept of virtual loop[3] will be introduced next. For close loop mechanism, a link group is regarded as the combination of non-coincident links between any independent loop and its adjacent loop. The mobility of a link group can be equal to 0, or less than 0, or more than 0. The links of the link group can be driven or motive. The link group defined here is a generalized group, not the Assur group. For the planar 10-bar linkage shown in Fig. 1, it contains five independent loops, and has five link groups, ABCD, EFG, HK, RSP and JM, as shown in Fig. 2. All rights reserved. No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of Trans Tech Publications, www.ttp.net. (ID: 128.119.168.112, Univ of Massachusetts Library, Amherst, USA-11/07/15,07:43:21) To describe the motion transmission manner of two adjacent loops by a terminology, we define the concept of virtual kinematic pair" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003458_9781118516072.ch5-Figure5.14-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003458_9781118516072.ch5-Figure5.14-1.png", "caption": "Figure 5.14. Representation of the synchronousmachine for zero sequence: (a) currents paths; (b)", "texts": [ " No emf is represented, since the synchronous machine produces positive-sequence voltages only. The reference bus of the equivalent circuit is the neutral of the generator. When zero-sequence currents flow through the armature, their instantaneous values in the three phases are equal and the resultant air gap flux is zero. Therefore, the zero-sequence reactance is approximately equal to the leakage reactance. The flow of the zero-sequence currents in the three phases and the neutral of the synchronous machine is shown in Figure 5.14a. The zero-sequence equivalent circuit is illustrated in Figure 5.14b. machine. The zero-sequence currents flowing through the three phases return through the grounding impedance Z n as a sum I n \u00bc I0 a \u00fe I0 b \u00fe I0 c \u00bc 3I0a : This is why a term 3Z n is introduced in the equivalent circuit of Figure 5.14b. If the neutral is isolated from the ground, then Zn \u00bc 1 and In \u00bc 0: The grounding impedance of the synchronous machine is not a part of either the positive- or negativesequence network because neither positive- nor negative-sequence current can flow through this impedance [6]. The reference bus of the equivalent circuit is the ground at the generator. Since all the neutral points of a symmetrical three-phase system through which balanced currents flow are at the same potential, all the neutral points have the same potential when either positive-sequence or negative-sequence currents flow [9]" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002155_icma.2013.6617946-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002155_icma.2013.6617946-Figure1-1.png", "caption": "Fig. 1 Geometry of the hypersonic vehicle model. The slender geometries and light structures of FAHV cause significant flexibility effects that severely affect the aerodynamics of the aircraft. Three factors that determine the ith flexible mode effects are the frequency i , damping ratio", "texts": [ " Vehicle Model The vehicle studied in this paper is the model developed by Bolender and Doman [1] for the longitudinal dynamics of a FAHV. Flexibility effects are included by modeling the fuselage as two cantilever beams clamped at the center of gravity, rather than a single free-free beam as done in [3]. This vibrational model captures the inertial coupling between the rigid-body states and the flexible states, resulting in a system that is more complex to control [1]. Assuming a flat Earth and normalizing the vehicle to unit depth, the longitudinal sketch of the vehicle is illustrated in Fig. 1, and the equations of motion are written as [5] ( cos ) / sinV T D m g (1) ( sin ) / ( ) cos /L T mV g V (2) sinh V (3) Q (4) yy f f a aI Q M (5) 22 / /f f f f f f f f f yy f a a yyk N M I I (6) 22 / /a a a a a a a a a yy a f f yyk N M I I (7) This model is composed of five rigid-body state variables [ , , , , ]Tx V h Q , where , , , ,V h Q are the velocity, flight- path angle, altitude, angle of attack, and pitch rate, respectively. It also includes four flexible states [ , , , ]T f f a a which correspond to the first generalized elastic deformations and their derivatives of the forebody (denoted with subscript f ) and aftbody (denoted with subscript a )" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002253_978-1-4614-2119-1_20-Figure20.1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002253_978-1-4614-2119-1_20-Figure20.1-1.png", "caption": "Fig. 20.1 Schematic of the bacterial flagellum. The substructures and components of the flagellum of E. coli. HC/NaC influx through the stator complex is believed to generate torque at the interface between the stator component of MotA and the rotor component of FliG. OM outer membrane, PG peptidoglycan, IM inner membrane", "texts": [ " We demonstrate transient-state control of NaC-driven flagellar motor rotational speed by switching local discharges between NaC-containing and -free solutions, and steady-state control by simultaneous local discharges of the solutions with controlling discharge velocities independently. Many bacteria swim actively in liquid by rotating their flagella and migrate toward favorable directions. The bacterial flagellum has a helical filament that acts as a propeller. Each flagellum consists of a helical filament extending from the cell body, a basal body embedded in the cell surface, and a flexible hook that connects them [20\u201323] (Fig. 20.1). The bacterial flagellar motor is a molecular machine that converts ion-motive force (IMF) into the mechanical force; the energy source of the rotation is the electrochemical gradient of HC or NaC ion across the cytoplasmic membrane. The rotational speed of the flagellar motor depends on IMF [24, 25]. In most situations, it can be considered that the concentration of the coupling ion inside the membrane is constant, because of the homeostasis of the bacteria. Therefore, the rotational speed of the flagellar motor mainly depends on the ion concentration in the external environment" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000654_sii.2012.6427383-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000654_sii.2012.6427383-Figure2-1.png", "caption": "Figure 2. Sectional view of tactile sensor", "texts": [ " However, when using the light of two layers, fixed thickness is needed for a sensor. It becomes a problem when this thickness makes a sensor smaller. Then, this research uses the light of one layer and aims at realization of the further downsizing of a sensor. And as compared with the sensor which uses the light of two layers, the detection accuracy of the force aims at an equivalent sensor. The small tactile sensor photograph used by this research is shown in Fig. 1. Moreover, the sectional view is shown in Fig. 2. This tactile sensor consists of three big parts, a camera, a force detection elastic body part, and a cylinder case part. The parts of previous research [6] were used for the parts of this sensor. In the future, if the algorithm of this research is adopted, it will become possible to carry out small size of the cylinder case part. The small CMOS camera (ACB-U 04v2) used for mobile terminals, such as a mobile phone, was used for the camera. The conceptual diagram of a force detection elastic body part is shown in Fig" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001260_s11465-011-0225-z-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001260_s11465-011-0225-z-Figure2-1.png", "caption": "Fig. 2 A segment erector consisting of the circumferential, radial, and pose adjustment mechanisms", "texts": [ " When the shield machine is cutting and thrusting, a segment erector is adopted to place the pre-cast concrete segments to the required locations of the tunnel inwall in a safe, efficient, and precise manner. The classical segment erector has three degrees of freedom with the ability to implement the necessary erecting actions, although it cannot adjust the pose of the segment [1]. To improve the quality and efficiency of the segment erection, the modern 6-DOF (degree of freedom) erector is invented [2]. A segment erector consists of three components, including the circumferential, radial, and poses adjustment mechanisms (Fig. 2). The radial mechanism in the classical erector only has one degree of freedom, and the two driving hydraulic cylinders move synchronously. Two degrees of freedom exist in the radial mechanism of a 6-DOF erector, and the pair of hydraulic cylinders can move both synchronously and differentially. The errors and disturbances in the control system create difficulties in implementation and in achieving precise synchronization control for the 2-DOF mechanism. Hence, this present paper proposes an improved 2-DOF radial mechanism for the erectors", " As can be seen, only a single degree of freedom exists in the radial mechanism, but there are two driving hydraulic cylinders in the PRRP mechanism. Here, the letters P and R denote the prismatic and revolute joints, respectively. Obviously, these two cylinders should move synchronously. Notably, the synchronization of two cylinders is not only implemented by the hydraulic control system, but is also ensured by the mechanical structure, such as link C in Fig. 3, showing an especially rotating angle for the circumferential mechanism not drawn in this figure. 2.2 A PRRRP mechanism in the existing 6-DOF erectors The PRRRP mechanism is adopted in the existing 6-DOF erector to adjust the segment pose (Fig. 4) where the circumferential mechanism and the pose adjustment mechanism are not drawn. There are two degrees of freedom in this radial mechanism. The two hydraulic cylinders can implement synchronous lifting and pose adjustment when the cylinders move synchronously and differentially, respectively. The synchronization of two cylinders depends entirely on the hydraulic control system" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002682_msec2011-50018-Figure7-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002682_msec2011-50018-Figure7-1.png", "caption": "Fig. 7: Tested telescopic cover (top view)", "texts": [], "surrounding_texts": [ "As mentioned above, a cover is an additional multi-mass system that interacts dynamically with the feed drive structure and its control. There are four possibilities for minimizing the disturbing forces transferred from the cover to the feed drive: A) An independent feed drive for the machine cover. The cover is driven by an independent feed drive. The dynamic forces generated by the cover are not transferred to the main structure. Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 05/24/2018 Terms of Use: http://www.asme.org/about-asme/terms-of-use 5 Copyright \u00a9 2011 by ASME B) Passive connection. Connection ensured by passive elements with constant stiffness K [N/m] and constant damping B [Ns/m]. Various behavior characteristics can be achieved by tuning these parameters. C) Semi-active connection. Connection ensured by a passive stiffness element and a semi-active damping element with a variable damping ratio. The mechanical energy can only be dissipated in this type of element, but this process is controlled in order to control the transferred force. D) Active connection. An active source of mechanical energy is used. The transferred force is controlled directly; the mechanical energy can be transferred in both directions - towards and from the mechanical system. This approach provides a wider range of force control boundaries compared to previous variant). moving structure and telescopic cover. The B) option is the simplest and most cost-effective. Therefore, it is the most relevant option for industrial application. This chapter describes the building of a dynamic model of the cover and finding of an optimum combination of connecting stiffness value and damping ratio. The dynamic model of the cover has been prepared in a Matlab/Simulink environment. The model contains a description of the whole cover using the submodels mentioned in the previous chapter. The connection between the cover and the feed drive is also included in the model. The connection is represented by two idealized parameters: ideal viscose damping and ideal stiffness. The connection is passive; mechanical energy can only be dissipated here, not produced. The model constitutes a basis for optimizing the connection parameters. The optimization criterion consists of various demands on the force that acts between the table and cover, and its time derivative, as well as the displacement between the table and the first segment of the cover, and its time derivative. A more detailed description of the criterion will follow. This approach aims to minimize the disturbance effects of the force acting between the table and the first cover segment. This disturbance causes table position errors and has an adverse effect on the accuracy of the manufacturing process. The properties of the model and the optimization of the connection will be shown on an example of a real cover and a real linear axis. The simulated results will be compared with experiments on a similar cover." ] }, { "image_filename": "designv11_101_0000533_s0021894412040086-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000533_s0021894412040086-Figure3-1.png", "caption": "Fig. 3. Fig. 4.", "texts": [ "5, and 4.0; z0 = 3 \u00b7 10\u221210, 4 \u00b7 10\u221210, 5 \u00b7 10\u221210, 6 \u00b7 10\u221210, and 7 \u00b7 10\u221210. Thus, the problem was solved 3\u00b7125 = 375 times. The following results were obtained: k0 = 10\u22124, G12 = 4, z0 = 4\u00b710\u221210, V\u0303c \u2223\u2223\u2223 r1=2.6 = 3.0653\u00b710\u22127, V\u0303c \u2223\u2223\u2223 r1=2.8 = 2.3295 \u00b7 10\u22127, V\u0303c \u2223\u2223\u2223 r1=3 = 1.8126 \u00b7 10\u22127, and F\u2217 = 7.0227 \u00b7 10\u22122. Below we give the results calculated for the optimal set of the parameters k0, G12, and z0. Figure 2 shows the dimensionless force G\u2032 z and drop velocity V as functions of the time t at r1 = 3. Figure 3 shows the dimensionless distance z1 covered by the drop since the instant of the cycle beginning as a function of the time t for different values of r1. The drop motion at r1 = 3 and t = 0.5 is schematically shown in Fig. 4 (the drop motion ceased at t = 0.8025). The macroscopic description of drop motion is given below. As the parameters V0 = 10\u22126 m/s and t0 = 10\u22122 s are small, it makes no sense to consider all cycles of drop motion. It is sufficient to consider a discrete set of 11 values of the cycle-averaged velocity V\u0303c at the points r1k = 3\u22120", " Solving this equation, we find t(r1) = r0 2 3\u222b r1 dr1 V\u0303c(r1) . (19) Substituting discrete values of r1k instead of r1 in Eq. (19) and calculating the integral by the method of trapezoids, we obtain a set of discrete values of tk (k = 0, 10). Figure 5 shows the experimental and calculated dependences r1(t). It is seen that these dependences differ only at the final stage of approaching of the drops (r1 < 2.25). This difference can be explained by the neglect of hydrodynamic interaction of the drops in constructing the calculated curve. Fig. 3. Dimensionless distance covered by the drop versus the dimensionless time during one cycle for r1 = 3 (solid curve) 2.8 (dotted curve), 2.6 (dashed curve), and 2.4 (dot-and-dashed curve). Fig. 4. Drop motion at r1 = 3 and t = 0.5: the solid curve is the boundary of the flow domain (the small half-circumference is the drop boundary, the straight segments are the segments on the axis of symmetry, and the large half-circumference is the external boundary of the flow domain); the dashed curve is the boundary of the liquid zone" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001172_ijmee.39.4.7-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001172_ijmee.39.4.7-Figure2-1.png", "caption": "Fig. 2 Force applied to a cylinder outside its radius.", "texts": [ "nternational Journal of Mechanical Engineering Education, Volume 39, Number 4 (October 2011), \u00a9 Manchester University Press http://dx.doi.org/10.7227/IJMEE.39.4.7 Keywords rolling cylinder; frictional force; roll expectation It is no surprise that the force applied to the cylinder (e.g. a wheel) in Fig. 1 will cause it to move to the right and roll clockwise on the fl at surface [1]. Yet a seemingly similar situation is surprising. The following situation was described by an engineering instructor who interrupted his lecture to ask a question [2]. Imagine a rod rigidly attached to a cylinder as in Fig. 2. The rod extends beyond the surface, and a force is applied to the right at the end of the rod. Which way will the cylinder roll [3]? The instructor claimed that the cylinder could move only to the right. Of course this seems logical if the cylinder slipped on the surface. But would this really happen if the cylinder was to roll? The professor proclaimed that the situation could be simply understood according to the direction of application of the force (along with Newton\u2019s laws). The applied force acted to the right, so only a force to the right would act on the cylinder\u2019s center of mass, causing the cylinder to move to the right", " The end of the rod would move to the left, since it would rotate clockwise further than the entire cylinder would translate along the surface. This suggests that if you pushed on the end of the rod with your fi ngers, and the cylinder moved to the right by rotating clockwise, the end of the rod would deadlock against your fi ngers. The forces would build up at the point of application until the cylinder was forced to slip. In other words, the cylinder could never rotate, since it would always have to slip. This seems absurd. In order to test the situation shown in Fig. 2, a prototype was devised. The prototype showed that the cylinder actually rotates counterclockwise and moves to the left. The professor was informed of the result, but did not believe it. He argued that the prototype was either built incorrectly or operated incorrectly. The professor\u2019s point of view is not isolated, and it is surely shared by some students (and others). In this note, a physics-based assessment is used to prove that the cylinder moves to the left with a counterclockwise rotation" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002256_978-1-4471-2330-9_10-Figure10.17-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002256_978-1-4471-2330-9_10-Figure10.17-1.png", "caption": "Fig. 10.17 Examine button", "texts": [ " 128 10 Animation of Crank-Slider Mechanism of a Piston Using Simulink\u00ae There are two ways to change the Viewpoint . The user can either create various Viewpoints in VRML when constructing the scene and select the desired Viewpoint before running the simulation in Simulink \u00ae or he/she can open the virtual scene before running the simulation and manipulate the view by using the mouse. In this section, the latter will be used to change the Viewpoint . Open the model created in the previous section and double-click the VR Sink . Click the Examine button (Fig. 10.17 ). Click and hold the left button of the mouse. The cursor will change to a cross (Fig. 10.18 ). Slowly drag the fi gure down to rotate the view (Fig. 10.19 ). Now you can run the simulation with the new view. 12910.5 Application Problem: Control of a Ball on a Plate Given a ball of mass = 0.008 (kg)m and radius =0 0.5 (m)r (Fig. 10.20 ). The plate is of dimensions \u00d7 \u00d7 37 0.2 7 (m ) . The plate is pivoted around the origin (0,0,0)O and has two independent actuators that can rotate it around x and z global axes by angles b and a , respectively ( 0" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003339_b978-0-08-098332-5.00002-4-Figure2.13-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003339_b978-0-08-098332-5.00002-4-Figure2.13-1.png", "caption": "Figure 2.13 Output voltage waveforms for 3-phase fully controlled thyristor converter supplying an inductive (motor) load, for various firing angles from 0 to 120 . The mean d.c. voltage is shown by the horizontal line, except for a\u00bc 90 where the mean d.c. voltage is zero.", "texts": [ " However, the power levels in most drives are such that in order to store enough energy to smooth the voltage waveform over the half-cycle of the utility supply (20 ms at 50 Hz), very bulky and expensive capacitors would be required. Fortunately, as will be seen later, it is not necessary for the voltage to be smooth as it is the current which directly determines the torque, and as already pointed out the current is always much smoother than the voltage because of inductance. The main power elements are shown in Figure 2.12. The 3-phase bridge has only two more thyristors than the single-phase bridge, but the output voltage waveform is vastly better, as shown in Figure 2.13. There are now six pulses of the output voltage per cycle, hence the description \u20186-pulse\u2019. The thyristors are again fired in pairs (one in the top half of the bridge and one \u2013 from a different leg \u2013 in the bottom half ), and each thyristor carries the output current for one-third of the time. As in the single-phase converter, the delay angle controls the output voltage, but now a\u00bc 0 corresponds to the point at which the phase voltages are equal (see Figure 2.13). The enormous improvement in the smoothness of the output voltage waveform is clear when we compare Figures 2.13 and 2.10, and it underlines the benefit of choosing a 3-phase converter whenever possible. The very much better voltage waveform also means that the desirable \u2018continuous current\u2019 condition is much more likely to be met, and the waveforms in Figure 2.13 have therefore been drawn with the assumption that the load current is in fact continuous. Occasionally, even a 6-pulse waveform is not sufficiently smooth, and some very large drive converters therefore consist of two 6-pulse converters with their outputs in series. A phase-shifting transformer is used to insert a 30 shift between the a.c. supplies to the two 3-phase bridges. The resultant ripple voltage is then 12-pulse. Returning to the 6-pulse converter, the mean output voltage can be shown to be given by Vdc \u00bc Vdocosa \u00bc 3 p ffiffiffi 2 p Vrmscosa (2", " It is probably a good idea at this point to remind the reader that, in the context of this book, our first application of the controlled rectifier will be to supply a d.c. motor. When we examine the d.c. motor drive in Chapter 4, we will see that it is the average or mean value of the output voltage from the controlled rectifier that determines the speed, and it is this mean voltage that we refer to when we talk of \u2018the\u2019 voltage from the converter. We must not forget the unwanted a.c. or ripple element, however, as this can be large. For example, we see from Figure 2.13 that to obtain a very low d.c. voltage (to make the motor run very slowly) a will be close to 90 ; but if we were to connect an a.c. voltmeter to the output terminals it could register several hundred volts, depending on the incoming supply voltage! In Chapter 4 we will discuss the use of the fully controlled converter to drive a d.c. motor, so it is appropriate at this stage to look briefly at the typical voltages we can expect. Utility supply voltages vary, but single-phase supplies are usually 220\u2013240 V, and we see from equation (2" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003001_detc2013-12231-Figure17-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003001_detc2013-12231-Figure17-1.png", "caption": "Fig. 17 Profile modification with the quantity variable along the contact line", "texts": [ " The instantaneous contact load distributions in a same meshing position for the four cases are shown in Fig. 16. It can be seen that case 4 generates the smallest maximum contact pressure and the edge contact phenomenon on the toe is completely removed, which shows the optimum performances. 5 Copyright \u00a9 2013 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/conferences/asmep/77583/ on 02/16/2017 Terms of Use: http://www.asme.org/about-asme/terms-of-use makes the analysis of the effect on the meshing performance easily, as shown in Fig. 17. The modification equation derived by Terauchi is applied here , 2.1 max L xCCm (4) Where x means the relative location along the meshing line of the meshing point, the original is located at the switch point from the two-pair-meshing phase to three-gear-pair meshing phase, L is the modification length which can be calculated by 1 / 2btL P , btP is the base pitch and is contact ratio. mC is the modification quantity at the x position, maxC is the maximum modification quantity, bfC max " ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003339_b978-0-08-098332-5.00002-4-Figure2.16-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003339_b978-0-08-098332-5.00002-4-Figure2.16-1.png", "caption": "Figure 2.16 Inverter output voltage and frequency control with pulse-width modulation.", "texts": [ " motor drives (see Chapter 7) we can arrange for the link voltage to track the output frequency of the inverter, so that at high output frequency we obtain a high output voltage and vice versa. This method of voltage control results in a simple inverter, but requires a controlled (and thus relatively expensive) rectifier for the d.c. link. The second method, which predominates in all sizes, achieves voltage control by PWM within the inverter itself. A cheaper uncontrolled rectifier can then be used to provide a constant-voltage d.c. link. The principle of voltage control by PWM is illustrated in Figure 2.16. At low output frequencies, a low output voltage is usually required, so one of each pair of devices is used to chop the voltage, the mark/space ratio being varied to achieve the desired voltage at the output. The low fundamental voltage component at low frequency is shown as a broken line in Figure 2.16(a). At a higher frequency a higher voltage is needed, so the chopping device is allowed to conduct for a longer fraction of each cycle, giving the higher fundamental output shown in Figure 2.16(b). As the frequency is raised still higher, the separate \u2018on\u2019 periods eventually merge, giving the waveform shown in Figure 2.16(c). Any further increase in frequency takes place without further increase in the output voltage, as shown in Figure 2.16(d). When we study a.c. drives later, we will see that the range of frequencies over which the voltage/frequency ratio can be kept constant is sometimes known as the \u2018constant-torque\u2019 region, and the upper limit of the range is usually taken to define the \u2018base speed\u2019 of the motor. Above this frequency, the inverter can no longer match voltage to frequency, the inverter effectively having run out of steam as far as voltage is concerned. The maximum voltage is thus governed by the link voltage, which must therefore be sufficiently high to provide whatever fundamental voltage the motor needs at its base speed. Beyond the PWM region the voltage waveform is as shown in Figure 2.16(d): this waveform is usually referred to as \u2018quasi-square\u2019, though in the light of the overall object of the exercise (to approximate to a sinewave) a better description might be \u2018quasi-sine\u2019. When supplying an inductive motor load, fast recovery freewheel diodes are needed in parallel with each device. These may be discrete devices, or fitted in a common package with the transistor, or integrated to form a single transistor/ diode device. As mentioned earlier, the switching nature of these converter circuits results in waveforms which contain not only the required fundamental component but also unwanted harmonic voltages" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002962_gt2013-94951-Figure16-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002962_gt2013-94951-Figure16-1.png", "caption": "Figure 16. Pressure distribution for eccentricity ratio 0.6", "texts": [ " A minimum value of 2\u03bcm of fluid film thickness is assumed as a criterion of contact for reasons of numerical stability. Hence, the analysis of the bearing for this specific configuration is going to be done for a maximum eccentricity of 0.6. Figure 15 illustrates the load carrying capacity of the bearing versus the eccentricity. The curve is a 5 th degree polynomial which is derived by interpolation of the discrete points for each eccentricity. The 5 th degree polynomial that describes the fitting of Figure 15 is given by equation (22). 33.961.575.1946.2574.2352.104 0 2 0 3 0 4 0 5 0 totf (N) (22) Figure 16 illustrates the contour plot of the pressure for 0.6 eccentricity ratio. Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 05/04/2014 Terms of Use: http://asme.org/terms 8 Copyright \u00a9 2013 by ASME As for the structural problem, figure 17 illustrates the contour plot of the Von Misses stress developed in the top foil. As seen from Figure 18, as the eccentricity increases, the velocity vector along the circumferential direction increases in magnitude due to larger top foil deflection" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001012_2012-01-1936-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001012_2012-01-1936-Figure1-1.png", "caption": "Figure 1. Driving simulator", "texts": [ " Next, in order to achieve early counter steering, when the body slip angle exceeds 10 degrees and the vehicle enters the drift state, the effect of reducing the steering wheel gear ratio was confirmed to be moderately effective. In other words, a new steering control method in which the control stability is improved according to the running situation was investigated. A preferable steering control method with respect to steering and vehicle behavior characteristics was examined in relation to DSA about RAS and the variable steering wheel gear in this research. A driving simulator facilitates experiments involving sudden collision avoidance. Therefore, a drift cornering driving simulator (Figure 1) was jointly developed by Nozaki and Mitsubishi Heavy Industries, Ltd. This device is able to simulate the bodily sensations of drift cornering due to its ability to produce (1) large lateral acceleration behaviors and (2) large yawing behaviors. The yawing mechanism can be rotated infinitely. Moreover, the effects of normal lateral acceleration can be reproduced using the roll function of the simulator. In total, the simulator can reproduce a maximum lateral acceleration of \u00b10.7 G. This large lateral acceleration is simulated for actual driving, and thus enables the examination of a preferable steering control method technology for drivers" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001472_jjap.52.10mb20-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001472_jjap.52.10mb20-Figure1-1.png", "caption": "Fig. 1. Configuration of RF magnetron sputtering system.", "texts": [ " On the other hand, Gnanarajan and Lam reported that an epitaxial orthorhombic twinned Ta2O5 film with (201) orientation was fabricated by the lowpressure thermal oxidation of epitaxial tantalum films on the R-plane of sapphire (Al2O3) substrates. However, the piezoelectricity and surface acoustic wave properties were not evaluated.24) In this study, Ta2O5 thin films were deposited on Al2O3 substrates, from which single crystallization due to epitaxial growth can be expected, using the RF magnetron sputtering system with LTS cathodes, and the crystalline and R-SAW propagation properties of the thin films were evaluated. Figure 1 shows the configuration of the RF magnetron sputtering system with two LTS cathodes, namely, cathode 1 and cathode 2, used for the deposition of Ta2O5 thin films. In general, the LTS cathode can produce a thin film with a smooth surface under appropriate conditions, because the substrate is not exposed to plasma directly. The sputtering parameters used are shown in Table I and are similar to those in our previous report.23) Ta2O5 thin films were deposited on the c- and R-planes of Al2O3 (c-, R-Al2O3) substrates" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000463_icnsc.2013.6548715-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000463_icnsc.2013.6548715-Figure1-1.png", "caption": "Fig. 1. A mobile robot with three optical mice.", "texts": [ " INTRODUCTION For the localization of a mobile robot, typical sensors including encoders, ultrasonic sensors, and cameras have wheel slip, ultrasonic sensors require the line of sight, and cameras usually mandate heavy computation. To overcome those problems, new attempts have been made to use optical mice that were originally invented as an advanced computer pointing device [3-11]. In fact, an optical mouse is an inexpensive but high performance motion detection sensor with a sophisticated image processing engine inside. Optical mice which are installed at the bottom of a mobile robot, as illustrated in Fig. 1, can detect the motion of a mobile robot traveling over a plane surface. There have been several works on the velocity estimation of a mobile robot using multiple optical mice. A pair of optical mice was proposed as a simple but effective means for the mobile robot velocity estimation in the presence of wheel slip [3, 4]. Using redundant velocity measurements from two optical mice, a simple procedure was developed for the error detection and reduction of the mobile robot velocity estimation [5]" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001744_iros.2011.6095082-Figure10-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001744_iros.2011.6095082-Figure10-1.png", "caption": "Fig. 10 Synchronizing steering link", "texts": [ " SCV I uses three motors which are mini robot servo \u201cKondo KRS-4034HVICS\u201d, and two motors are used by the driving pulleys made of aluminum alloy and one motor is for steering. The power supply is lithium polymer battery which is 3 cells battery and 11.1V. The center pulleys which are black color are passive wheels that are rotating by pulling force of the flexible crawler belt. The small rollers on upper side and lower side of the center body are guide rollers and tension rollers of the flexible crawler belt. Both steering arms, at the front and at the rear, swing in a synchronized motion by one motor and steering links, as shown in Fig. 10. The motor drives only front side steering arm and the rear side steering arm is driven by the front steering arm using steering link. The steering links put on the common tangent line of the circles around axis of the front and rear steering arms, and both tangent points are connected by using rod ends. Fig. 11 is the flexible crawler belt. It should be estimated the shape after it retracts, and design the retractable ratio in each parts. Fig. 12 is the simulation results of the retraction ratio of the flexible crawler belt" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001425_ijptech.2011.038108-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001425_ijptech.2011.038108-Figure2-1.png", "caption": "Figure 2 Model creation using AutoCAD (see online version for colours)", "texts": [ " Following objectives have been set for present experimental study: 1 to study feasibility of decreasing the shell thickness from recommended one (12 mm) for statistically controlled RC solution of aluminium alloy in order to reduce the production cost and time 2 to evaluate the dimensional accuracy of the castings obtained and to check the consistency of the tolerance grades of the castings (IT grades) as per allowed IS standards for casting process 3 proof of concept, to present the concept in physical form with minimum cost by avoiding the cost of making dies and other fixtures for a new concept. In order to accomplish the above objectives, \u2018aluminium casting\u2019 has been chosen as a benchmark (Figure 1), representative of manufacturing field, where the application of RT and RC technologies is particularly relevant. The experimental procedure started with drafting/ model creation using AutoCAD software (Figure 2). For the process of RC process based on 3DP, following phases have been planned: 1 After the selection of the benchmark, the component to be built was modelled using a CAD (Figure 3). The CAD software used for the modelling was UNIGRAPHICS Ver. NX 5. 2 The upper and lower shells of the split pattern were made for different values of the thickness. The thickness values for shells thickness were 12, 9, 7, 6, 5, 4, 3, 2 and 1 mm. 3 The CAD models of upper and lower shells were converted in to STL (standard triangulation language) format also known as stereo lithography format (Figure 4)" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002317_978-94-007-4201-7_5-Figure5.1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002317_978-94-007-4201-7_5-Figure5.1-1.png", "caption": "Fig. 5.1 A 3-DOF planar linkage", "texts": [ " KIC is a powerful tool for analyzing complicated mechanisms. In this chapter, the concept of KIC is introduced. Then, first- and second-order KICs are derived, followed by a discussion of the velocity and acceleration analyses. At the last part, the lower-mobility parallel mechanisms are analyzed. Z. Huang et al., Theory of Parallel Mechanisms, Mechanisms and Machine Science 6, DOI 10.1007/978-94-007-4201-7_5, # Springer Science+Business Media Dordrecht 2013 135 First, let us discuss a planar 3-degree of freedom (DOF) mechanism, Fig. 5.1, to illustrate the KIC concepts. When three input parameters, f1 , f2 , and f3 are given for the 3-DOF linkage, all the link motions of the mechanism are determined. For the more common situation, where an N-DOF mechanism is present, we may give the common expressions for the ith link as follows: Fi \u00bc f 1\u00f0\u20191 \u20192 \u2019N\u00de Xi \u00bc f 2\u00f0\u20191 \u20192 \u2019N\u00de i \u00bc 1; 2; . . . ;m Yi \u00bc f 3\u00f0\u20191 \u20192 \u2019N\u00de (5.1) whereFi, Xi, Yi are the three configuration parameters for link i, and\u20191; \u20192; ; \u2019N are N input variables. Considering \u20191; \u20192; ; \u2019N as time variables, we have _Fi \u00bc XN n\u00bc1 @f 1 @fn _fn _Xi \u00bc XN n\u00bc1 @f 2 @fn _fn _Yi \u00bc XN n\u00bc1 @f 3 @fn _fn (5" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000747_amm.437.152-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000747_amm.437.152-Figure2-1.png", "caption": "Fig. 2 Coulomb contact points on interface between blade(left) and mortise(right)", "texts": [ " Matrix27 element whose nodes were corresponding Coulomb contact points was used to simulate the additional stiffness and damping effects caused by dry friction force [12]. Iterative program was wrote to obtain parameters and dynamic response was calculated afterwards. In order to compute the dynamic response of assembly structure considering friction damping on interface, two assemblies(blade and mortise) are modeled separately. Discrete micro-slip model is used and each node on the contact surface is the Coulomb contact point. Hence, a perfect match of meshes on the contact surface is necessary, as shown in Figure 2. Points marked are the Coulomb contact points which would be connected by Matrix27 element. Named two random corresponding nodes on blade and mortise as node A and node B. The mechanical relationship between node A and node B can be described as connecting by a normal spring, a parallel spring and a Coulomb friction element, as shown in Fig. 3.That is exactly what Matrix27 element simulates. The normal spring's stiffness kn is the vertical contact stiffness under a certain normal force; the parallel spring's stiffness kt is the shear contact stiffness" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000381_1.3606146-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000381_1.3606146-Figure2-1.png", "caption": "FIGURE 2. Elevation View of Bend Strip In literature [1-2], bending strain at the outer surface of the tape is defined as", "texts": [], "surrounding_texts": [ "Second generation 344 YBCO coated conductor are expected to be robust and flexible against electromechanical and mechanical loads. In operations bending stress/strain are unavoidable. It is, therefore, of fundamental as well as technical relevance to quantify the critical bending strain limit of the tape used. In the present work, critical current of virgin 344 YBCO tape was measured under self-field criteria (1 \u00b5V/cm) at 77 K. Then the tapes had been subjected to different bending loads and subsequently critical current was measured. The bending stresses were simulated in ANSYS and calculated analytically also. The micro-structural analysis of the tape also has been performed." ] }, { "image_filename": "designv11_101_0001712_0954406212438142-Figure8-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001712_0954406212438142-Figure8-1.png", "caption": "Figure 8. Co-ordinate system and forces acting on the rigid shaft model of a turbocharger with floating ring bearings including gyroscopic moments: la and lb are the distances between the bearing centres and the rotor centre of gravity; J! is the angular momentum of the rotor, where ! is the spin speed and J is the polar moment of inertia of the rotor; J! _ , J! _ are the gyroscopic moments about y and x axes, respectively.", "texts": [ " It is clear from this graph, that the conical mode is stabilised when \u00bc 1=2 and that for values of 51=2, the system is stable. The analysis, here, demonstrates the significance of the threshold value of \u00bc 1=2 in stabilising the conical whirl. However, the real turbocharger has an asymmetric rather than a symmetric rotor and often has floating ring bearings. Further investigation of the effects of gyroscopic moments in the system is conducted in the following section. Gyroscopic effect in an asymmetric turbocharger rotor with floating ring bearings Figure 8 shows a schematic diagram of an asymmetric rotor of a turbocharger with floating ring bearings and the forces acting on the rotor due to both the transverse at UNIV OF MICHIGAN on June 20, 2015pic.sagepub.comDownloaded from and the angular motion discussed above; la, lb are the distances of the centre of gravity of the rotor from the turbine and the compressor bearing centres, respectively. Assuming a rigid rotor of mass 2m mounted in two identical bearings with full-film, the equations of transverse motion of the turbocharger in the inner-film of the bearing is given as9 fa \u00fe fb \u00bc 2m\u20acr1c \u00f015\u00de where r1c \u00bc r1c, s1cf gT, r1c, s1c are the co-ordinates of the centre of gravity of the rotor", " Using the oil-film forces given in equation (2) for the bearings, the bearing forces are given by9 fa \u00bc 2A\u00f0_r1a _ra\u00de \u00fe A!!\u00f0r1a ra\u00de \u00f016\u00de fb \u00bc 2A\u00f0_r1b _rb\u00de \u00fe A!!\u00f0r1b rb\u00de \u00f017\u00de where ri \u00bc ri, sif gT, i \u00bc a, b, 1a, 1b, in which ra, b are the vectors of co-ordinates of the floating ring centres CB a,b in the turbine and compressor end bearings, and r1a, 1b are the corresponding vectors of co-ordinates of the journal dynamic centres CD a,b. The equation of angular at UNIV OF MICHIGAN on June 20, 2015pic.sagepub.comDownloaded from motion with the gyroscopic effect as seen in Figure 8 is given by I\u20ach\u00fe J!!_h\u00fe fTbL fTaL \u00bc 0 \u00f018\u00de where L \u00bc lb 0 0 la \" # . The linear co-ordinate transforma- tion can be done, using r1c\u00bc r1b \u00fe L l r1a r1b\u00f0 \u00de \u00f019\u00de h \u00bc r1b r1a\u00f0 \u00de l \u00f020\u00de where the derivatives of h are given by the corresponding derivatives of the right hand side of equation (20). Substituting for the forces from equations (16) and (17), and using the transformation in equations (19) and (20), the equation of motion in matrix form is given by 2m \u20acr1b 1 lb l \u00fe \u20acr1a lb l \u00fe 2A \u00f0_r1a _ra\u00de \u00fe \u00f0_r1b _rb\u00de\u00bd \u00fe A" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001430_s12206-012-1266-x-Figure4-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001430_s12206-012-1266-x-Figure4-1.png", "caption": "Fig. 4. Stresses and displacements in case that the spacer with an initial gap 1.0 mm was used.", "texts": [ " The virtual part, similar to the rigid link, in CATIA structural analysis was used to connect the center of the load to the wheel shaft. As described in the role of the spacer, the preload on the bearing depends on the gap between the spacer and the inner race. Two typical cases with an initial gap 1.0 mm and no gap were analyzed. First, it is assumed that the initial gap was 1 mm. The structural analysis result shows that both bearings were well pre- loaded as the lock nut tightening force was applied in Fig. 4(a). Next, as the dual-tire load was applied additionally, the upper part of the left bearing was the most severely compressed and the lower part of the right bearing was compressed less than that, as shown in Fig. 4(b). The contact stresses on the lock nut were roughly less than 1 MPa. The stresses were too small to initiate yielding. The lowered annular surface could not be formed by the simple compressive contact stress. It is discovered that even when the wheel shaft was supported by the well preloaded bearings, it rotated slightly in plane, not out of plane, with respect to the carrier housing. The lock nut slid by 0.18 mm on the planet carrier face, as shown in Fig. 4(c). Second, it was assumed that there was no initial gap between the spacer and the inner race. The structural analysis result shows that both bearings were not preloaded and only the inner races were loaded as the lock nut tightening force was applied (Fig. 5(a)). Next, as the dual tire load was applied, the upper part of the left bearing was the most severely compressed and the lower part of the right bearing was compressed, as shown in Fig. 5(b). Compared to the first case, the compressed portion of the roller was smaller in the second case" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000589_amr.591-593.92-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000589_amr.591-593.92-Figure3-1.png", "caption": "Fig. 3 Block diagram of optimization", "texts": [], "surrounding_texts": [ "Maximum Flight Speed. Compound helicopter maximum speed depends on the power balance, flow separation constraints and shock stall limit. Respectively, calculate the maximum speed corresponding to the power limit, flow separation limit and shock stall limit, take the minimum as the maximum speed. Calculate the maximum speed 1V by the power limit. 32 3 2 7 12 1 1 2 1 ( ) [ ( ) ] 150 ( ) 44 p r xp x kp N R R J C c V R V \u03c1\u03c0 \u03c3 \u03c1 \u03d5 = \u2126 + + \u2126 (1) a rN N N\u2206 = \u2212 (2) Calculate the maximum speed 2V by the flow separation constraints. 3 max 2 ( ) 24 4 T yR k C R V p \u03c1 \u03d5\u03c3\u2126 \u2126 = \u2212 (3) Calculate the maximum speed 3V by shock stall limit. ( )3 0.7aljV M R\u03b1= \u2212 \u2126 (4) Hovering Ceiling. Hovering ceiling out of ground effect OEGH . By the power balance relationship: a rN N= ( )3 7 1 75 2 ( 1 0.021 ) 64 250 e p xp p G H N A GJ R k C p \u03c2 \u03c3 \u03d5 = + \u2126 \u2206 \u2206 = \u2212 \u2206 (5) Iteration hovering height H, until a rN N N\u2206 = \u2212 is less than the required value. Hovering ceiling in ground effect IEGH . 3/2 2 0 1 7 1 7 1 1 ( ( ) ) 4 4 2 T k p k k p y y JC m k m m k C C\u03c3 \u03b4 \u03c3 \u03b4 \u03b4 \u03d5\u221e \u221e= +\u039b\u2206 \u2206 = + + (6) \u039b is a function of /h R . Using the power balance, and seek no ground effect hover ceiling method, in ground effect hover ceiling IEGH can be obtained. Endurance Performance. Maximum range: ( )max min / f e r G L C N V\u03c2 = (7) Maximum endurance: [ ]max min / f e r G T C N \u03c2 = (8) Maximum Rate of Climb and Severing Ceiling. Maximum Rate of Climb: ( )max min 75 y ps a r N V k N N N G \u2206 = \u2206 = \u2212 (9) Iteration flight altitude, until max 0.5 /yV m s\u2264 , corresponding height is the severing ceiling. While iteration, requirements of flow separation should be checked." ] }, { "image_filename": "designv11_101_0000037_s1064230711020109-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000037_s1064230711020109-Figure2-1.png", "caption": "Fig. 2. Swivel wheel.", "texts": [ " These program values are sent to the tracking systems. Visual feedback is performed during the robot motion. The operator observes the control process and determines, whether the platform has reached the desired position. After the control objective is reached, the handle is returned to the neutral position. DOI: 10.1134/S1064230711020109 326 JOURNAL OF COMPUTER AND SYSTEMS SCIENCES INTERNATIONAL Vol. 50 No. 2 2011 MARTYNENKO et al. 1. ROBOT DESIGN AND KINEMATICS Figure 1 shows the structural diagram of the considered vehicle, and Fig. 2 shows the wheel mounted on the fork. The vehicle consists of the platform which can move in the horizontal plane in four wheels. \u2013 are the wheel centers situated at the vertices of the regular rectangle whose plane is parallel to the platform. C denotes the center of the rectangle. The point \u0421 is motionless with respect to the platform. Solid lines in Fig. 1 show the planes in which the wheels are situated. The robot represents the system of nine solid bodies (platform, four forks, four wheels) connected by cylindrical joints" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000219_imece2011-64870-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000219_imece2011-64870-Figure1-1.png", "caption": "FIGURE 1. PROGRESSIVE CLEARANCE LABYRINTH SEAL CONCEPT", "texts": [ " When the tip-clearance reduces below the equilibrium clearance, fluidic feedback forces cause the packing ring to open. Conversely, when the tip-clearance increases above the equilibrium clearance, the fluidic feedback forces cause the packing ring to close. Due to this self-correcting behavior, the seal provides high differential pressure capability, low leakage and non-contact operation even in the presence of large rotor transients. Theoretical models for the feedback phenomenon have been developed and validated by experimental results. As shown in Fig. 1, we propose a labyrinth seal packing ring with the following features: \u2013 The rotor-labyrinth teeth clearances get progressively tighter from the upstream (phi) to the downstream (plo) side. \u2013 The packing ring is mounted on flexures. \u2013 The packing ring and the stator have a secondary noncontact sealing mechanism (details not shown). The progressively tighter clearances result in a selfcorrecting phenomenon that maintains a tight equilibrium clearance with the rotor. If the rotor moves too close to the seal, the seal moves away to protect the teeth" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003694_0307174x1304001010-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003694_0307174x1304001010-Figure2-1.png", "caption": "Figure 2. The geometry of a three-dimensional shell laid out on a reduced toroidal surface", "texts": [ " This makes it possible to calculate the geometry of the inflated tyre without allowing for its initial state, which was done in earlier studies [21, 22]. In contrast to the studies cited, in the new calculation method proposed here, the initial undeformed state is used. Of the many possible initial states, the state giving the simplest calculation formulae is selected. As that initial state, the position of a three-dimensional shell on a toroidal surface with a cross-section in the form of an arc of a circle of radius r0, passing through points of the tyre rim, was chosen (Figure 2). The radius r0 of the reduced circle is determined by iterations on condition that the length L0 is retained. After determination of r0, the angle of inclination of the cords bc on the equator of the torus is found, as well as other parameters. Figure 3 presents a diagram of displacements of a point of the cross-section when the shell is loaded. The material point of the cross-section of the round shell, having rectangular coordinates (ys, rs) and the curvilinear coordinate s, is displaced under axisymmetrical loads in the direction of the radius r0 of the circle by distance w, and in a direction perpendicular to the radius r0 by distance u" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003171_cp.2012.0276-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003171_cp.2012.0276-Figure1-1.png", "caption": "Fig. 1. Symbols and definition of subdomains.", "texts": [ " A 2-D relative airgap permeance was used in [7], but it is still too simple to accurately account for In this paper, an analytical model will be presented for accurate prediction of both radial and circumferential force distributions and UMF accounting for slotting effect, based on the subdomain field model accounting for tooth-tips. In addition, the CP model will also be employed in this paper to analyse the force distribution and UMF. The accuracy of the developed analytical model will be compared with that of the CP model on an 8-pole/9-slot machine. The analytical model is based on the following assumptions: (1) infinite permeable iron materials; (2) negligible end effect; (3) linear magnet properties; (4) simplified slot but with tooth-tips as shown in Fig. 1; (5) non-conductive stator/rotor laminations; (6) uniformly distributed current density in conductor area. The coils may be accommodated in the slot in two alternative ways as shown in Fig. 2, viz. two-circumferential-layer and two-radial-layer windings. The single layer winding, viz. one coil side per slot, can be considered as a special case of Fig. 2(b), having current densities Ji1=Ji2. Rs, Rm, Rr, Rsb, and Rt are the radii of the stator bore, magnet, rotor yoke, slot bottom, and slot top, respectively", " The 2D vector potential has only z-axis component and is governed by: 2 2 0 2 2 2 1 1z z z rA A A MM r r rr r (1) in the magnets (Az1), 2 2 02 2 2 1 1z z zA A A J r rr r (2) in the slot (Az3), and 2 2 2 2 2 1 1 0z z zA A A r rr r (3) in the airgap including the retaining sleeve (Az2) and slot opening (Az4), where r and are the radial and circumferential positions, and Mr and M are the radial and circumferential components of magnet magnetisation: rM MM r (4) 1,3,5, cos sinr rck rsk k M M k M k (5) 1,3,5, cos sinck sk k M M k M k (6) 0cosrck rk rM M k t k (7) 0sinrsk rk rM M k t k (8) 0sinck k rM M k t k (9) 0cossk k rM M k t k (10) For radial magnetisation: 0 4 sin / 2 / / 1, 3, 5,r rk p pBM k p k p k (11) 0 / 1, 3, 5,kM k p (12) For parallel magnetisation: 0 1 2/ / 1, 3, 5,rk r p k kM B A A k p (13) 0 1 2/ / 1, 3, 5,k r p k kM B A A k p (14) where 1 1 1 / 2 / sin 1 / 2 /k p pA k p k p (15) 1 2 1 / 2 / sin 1 / 2 /k p pA k p k p (16) where 0 is the relative permeability of air, Br is the residual flux density of magnet and p is the pole-arc to pole-pitch ratio, r is the rotor rotational speed, 0 is the rotor initial position, p is the number of pole pairs. The current density J in Fig. 2(a) can be expressed by: 0 cos / 2i in n sa i n J J J E b (17) for i-bsa/2 i+bsa/2, where i is the position of the ith slot, bsa is the slot width angle, Fig. 1, En=n /boa, Ji0=(Ji1+Ji2)d/bsa, and Jin=(2/n/ )(Ji1+Ji2cosn )sin(n d/bsa). Considering the boundary condition along the rotor yoke, the vector potential in the magnets can be expressed by: 1 1 1 2 3 1 1 2 3 cos sin z k k ck k rsk k k k sk k rck k A C A C M C M k C C C M C M k (18) where C1k=[(r/Rm)k+G1(r/Rr)-k], G1=(Rr/Rm)k, and 0 2 2 / 1 k k r rC R k r R r k (19) 0 3 2 / 1 k k r rC R r R kr k (20) The vector potential in the air-gap including the retaining sleeve can be expressed by: 2 7 2 8 2 7 2 8 2 cos sin z k k k k k k A C A C B k C C C D k (21) where C7k=(r/Rs)k and C8k=(r/Rm)-k", " Considering the boundary condition along the bottom and both sides of the slot, the vector potential distribution in the slot of the two-circumferential-layer winding machine can be derived as: 3 0 cos / 2z i n n sa i n A A A E b (22) where 2 2 0 0 0 32 ln / 4i sb iA J R r r Q (23) 3 3 2 2 0 2 / / 2 / / 4 n n n E E n i sb t Ein n sb sb n A D G r R r R J r E R r R E (24) where Rsb and Rt are the radii of the bottom and top of slot, respectively, and 3 / nE t sbG R R (25) For two-radial-layer winding machine, the vector potential distribution in the slot can be given by: 3 0 cos / 2z bi bin n sa i n A A A E b (26) for the bottom part of the slot and 3 0 cos / 2z ti tin n sa i n A A A E b (27) for the top part of the slot, where 2 2 0 0 2 32 ln / 4bi i sb biA J R r r Q (28) 2 2 0 0 1 2 2 0 2 3 ln / 2 / 2 ln / 2 ti i sm i sb sm ti A J R r r J R R r Q (29) 3 3 / /n nE E bin tin i sb tA A D G r R r R (30) Satisfying the boundary condition on both sides of the slot opening, the general solution of vector field in the ith slot opening can be given by: 4 4 4 4ln / / cos / 2 m mF F z i i i t i s m m oa i A D r Q C r R D r R F b (31) where boa is the slot opening width angle as shown in Fig. 1, D is constant, and Fm=m /boa. The radial and circumferential components of flux density in the airgap can be obtained from the vector potential: 2 sin cosr rsk rck k k B B k B k (32) 2 cos sinck sk k k B B k B k (33) where 2 2/ / /k k rsk s mB k A r R B r R r (34) 2 2/ / /k k rck s mB k C r R D r R r (35) 2 2/ / /k k ck s mB k A r R B r R r (36) 2 2/ / /k k sk s mB k C r R D r R r (37) The unknown coefficients A1, B1, C1, D1, A2, B2, C2, D2, C4i, D4i, D3i, Q3bi, Q3ti, Q3i, and Q4i in the expressions of vector potentials can be determined by using the infinitely permeable boundary conditions (zero circumferential component of flux density) and interface conditions (continuity of normal flux density and circumferential vector potential)" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002278_978-1-4471-4628-5_3-Figure3.90-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002278_978-1-4471-4628-5_3-Figure3.90-1.png", "caption": "Fig. 3.90 Schematic diagram of absolute encoder", "texts": [ " The light intensity signal Q passes through a switch which is modulated by the angular position of the disc (\u03b8 or \u03b8 + \u03c6, depending on the track). The output intensity from the switch modulates the MR element which models the photo resistor. The resistance of the photo resistor changes when light fall on it. The current passing through the photoresistor generates the pulse which is fed to the post-processing device such as a microprocessor. The voltage supplied to both LEDs and photoresistors is V and the rotary inertia and torque applied on the mechanical rotating system are Jd and \u03c4 , respectively. Absolute Encoder Figure 3.90 shows the schematic diagram of an absolute encoder for the measurement of angular displacement. Here we get the output in the form of a binary number of several digits. Here each number represents a particular angular position. Figure 3.90 shows a four-bit absolute encoder where the rotating disc has four concentric circles of slots. There are four light emitting diodes (LEDs) to emit the light and four photo resistors to detect the light. These slots are arranged in such a way (as shown in Fig. 3.90) that the sequential output from the sensors is a number in binary code. The number of tracks decides the number of bits in the binary number. For four tracks, there will be four bits and the number of positions that can be detected will be 24.Thus the resolution of encoder will be 360/24 i.e., 22.5\u25e6. More number of circles improves the resolution. The most common types of numerical encoding used in the absolute encoders are gray codes. The gray code is designed so that only one track (one bit) will change state for each count transition, unlike the binary code where multiple tracks (bits) change at certain count transitions" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001545_s1068366613040077-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001545_s1068366613040077-Figure1-1.png", "caption": "Fig. 1. Schematic for calculating temperature of fast mov ing linear source.", "texts": [ " CALCULATION The solution of the problem formulated includes the following two stages: (1) Deriving a formula of the temperature field for a plane (band) FMHS. Initially, original expres sion (1) is integrated over x1 using the substitution ( and the integration limits are u1 = \u2013\u221e at x1 = \u2013\u221e and u2 = \u221e at x1 = \u221e) to deter mine the following relation for the temperature of an instantaneous plane (IP) source with the density QIP as follows: (2) The subsequent calculation is related to the two fol lowing simplified assumptions for the passage of the ele mentary band dx on the surface of the half infinite body through a constant linear (CL) source [3] (Fig. 1): \u2014the time of the effect of the source on dx is taken as the instantaneous time dt; \u2014during the time dt, all of the heat flow propa gates only perpendicularly to the band dx. As a result, for the band dx, we have the IP source QIP (J/m2) and, for the time interval dt, we have the CL source qCL (W/m). From the heat balance equation QIPdx = qCLdt (J/m), we obtain QIP = qCL Then, at QIP = qCL/v and t = in Eq. (2), we obtain the fol lowing temperature of the linear FMHS: (3) The second integral transition leads to the follow ing dependence of the temperature field of a plane FMHS on the heat source shape or the function of the distribution of normal contact stresses : (4) 1 4 x x u at \u2212 = \u2212 1 4dx atdu= IP IP " ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000476_amm.419.795-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000476_amm.419.795-Figure1-1.png", "caption": "Fig. 1 WT wheelchair robot", "texts": [ " However, this kind of moving mechanism is so complicated that the interaction between robot and stairs environment should be taken fully account into during wheelchair robot's stair-climbing, and the pattern which the wheelchair is in during stair-climbing should be determined. The pattern of WT wheelchair robot during stair-climbing alternatively changing, which brings difficulty to the path planning of the robot. Path planning in the traditional sense is the ideas of inverse kinematics solution, which plans the path curve in the joint's space of robot [3]. This paper puts forward an idea that the seat plane of the wheelchair should be maintained horizontal to determine each joint's generalized coordinate variable. WT wheelchair robot, as shown in Fig. 1, is mainly composed of a body, two front flippers, two driving wheels, a back flipper and two tracks installed symmetrically. The body AB consists of a supporting frame and a seat connected together, and there are guide wheels A, carrier wheels B, and the belt pressing wheels F and G symmetrically arranged at both sides of the supporting frame; The All rights reserved. No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of Trans Tech Publications, www", " Experiments show that the process is so short that the influence of turning direction can be ignored, and the process can be simplified into the vertical plane. WT wheelchair robot's process of stair-climbing is shown in Fig. 2. First, the back flipper rotates counterclockwise so that the tension force of track increases, and then rotate the front flippers clockwise or counterclockwise to keep track tension constant. Finally, the seat plane angle relative to the horizontal plane can be controlled by swing the back flipper to different angles simultaneously driving the wheelchair robot backwards. The relative coordinate system, as shown in Fig. 1 (b), can be established. Its origin is located at the axis of wheel A, u axis is parallel to the plane of wheelchair's seat, and its v axis is vertical. q1 can be defined as the angle rotated counterclockwise from BA to BE, q2 and q3 can be defined as the angles of the left and the right front flippers rotated clockwise from CA to CD and to C'D' respectively, q4 and q5 can be defined as the angles of the left or the right driving wheels rotated clockwise around the back flipper BE respectively. The above can be seen as the WT wheelchair robot's generalized coordinate variables" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003368_b978-0-08-098332-5.00006-1-Figure6.25-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003368_b978-0-08-098332-5.00006-1-Figure6.25-1.png", "caption": "Figure 6.25 Shaded pole induction motor.", "texts": [ " The inherent difference in impedance is sufficient to give the required phase-shift between the two currents without needing any external elements in series. Starting torque is good at typically 1.5 times fullload torque, as also shown in Figure 6.24. As with the capacitor type, reversal is accomplished by changing the connections to one of the windings. There are several variants of this extremely simple, robust and reliable cage motor, which is used for low-power applications such as hair-dryers, oven fans, office equipment, and display drives. A 2-pole version from the cheap end of the market is shown in Figure 6.25. The rotor, typically between 1 and 4 cm diameter, has a die-cast aluminum cage, while the stator winding is a simple concentrated coil wound round the laminated core. The stator pole is slotted to receive the \u2018shading ring\u2019, which is a single short-circuited turn of thick copper or aluminum. Most of the pulsating flux produced by the stator winding by-passes the shading ring and crosses the air-gap to the rotor. But some of the flux passes through the shading ring, and because it is alternating it induces an e" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000282_icems.2011.6073966-Figure13-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000282_icems.2011.6073966-Figure13-1.png", "caption": "Fig. 13 The flux-lines of Motor-B", "texts": [ " It seems that the FEM results shown in Fig. 12 express the variation of the UMP center. As the average radius of Motor-A\u2019s airgap is 14.35 mm and the variation of the axial UMP center is less than 0.07 mm, the center variation shown in Fig. 12 can be considered as zero. It should be pointed out, the curves in the figure varies randomly; they are actually the numerical error which is ineluctable in using FEM. Using optimized FE mesh density and element shape can reduce such an error. The FEM result of Motor-B\u2019s flux lines are shown Fig. 13, and motor\u2019s UMP is shown in Fig. 14. As the motor has 3 stator slots and 2 pole-pairs, the minimum common multiple of Z and 2p is 12. Therefore, the cycle width of UMP\u2019s fundamental harmonic is 30\u00b0, and this is confirmed by the UMP curve shown in Fig. 14, which is obtained with FEM. The average radius of the Motor-B airgap is same as MotorA, i.e., 14.35 mm. As its Z=3 and p=2, from the analysis in the Section-V, the location of the UMP center varies in the motor operation, and it is confirmed by the FEM results shown in Fig" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000900_amr.562-564.654-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000900_amr.562-564.654-Figure2-1.png", "caption": "Fig. 2 The non-coincident links between everytwo adjacent loops", "texts": [ " For close loop mechanism, a link group is regarded as the combination of non-coincident links between any independent loop and its adjacent loop. The mobility of a link group can be equal to 0, or less than 0, or more than 0. The links of the link group can be driven or motive. The link group defined here is a generalized group, not the Assur group. For the planar 10-bar linkage shown in Fig. 1, it contains five independent loops, and has five link groups, ABCD, EFG, HK, RSP and JM, as shown in Fig. 2. All rights reserved. No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of Trans Tech Publications, www.ttp.net. (ID: 128.119.168.112, Univ of Massachusetts Library, Amherst, USA-11/07/15,07:43:21) To describe the motion transmission manner of two adjacent loops by a terminology, we define the concept of virtual kinematic pair. Assume that links M and J are responsible for the motion transmission between two adjacent loops" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001870_icmtma.2011.839-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001870_icmtma.2011.839-Figure1-1.png", "caption": "Fig. 1 The simulation experiment platform of torsional vibration of rolling mill\u2019s main driving system", "texts": [ " SIMULATION EXPERIMENT PLATFORM AND TESTING In this paper, the correlation dimension should be studied about the self-excited torsional vibration signal of the simulated crack rotating shaft of the rolling mill\u2019s main driving shaft which were simulated using the experiment platform. The results show that correlation dimension of fractal theory analysis could extract more sensitive malfunction characteristics of the running equipment. This kind of feature could reflect dynamic characteristics of the complex torsional vibration system. Fault identification means could also be reflected by the relation between lyapunov exponent and embedding dimension. Simulation experiment platform was show as Fig. 1. Torsional vibration platform was as the follows size: length was 5100 mm, bed base high was 810 mm, overall height was 1770 mm, overall width was 1600 mm. The signal acquisition system was show as Fig 2. Among it, test system was LMS Test. Lab. Experimental sensors parameters were as follows: acceleration sensors B&K4508, electromagnetic pressure sensor CZ-TD-305A, set-type torque sensor JN338, eddy current vibration displacement sensor FZD18, frequency meter XSM, amplitude meter XST, constant tension control device HD400, charge amplifier YE5850" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000294_pedes.2012.6484419-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000294_pedes.2012.6484419-Figure1-1.png", "caption": "Figure 1. Six-pole three-phase winding with parallel-connected PPGs. Only the phase A is shown connected. The numbers refer to all the n=6p=18 PPGs.", "texts": [ " However, the saturation must be taken in account through corrective factors, mainly for the on-load operation [8]. This paper shows the theoretical basis of the method and simulations for a 1950kVA machine. The method was implemented in a LabView virtual instrument (VI) [8]; experiments were performed on a 17kVA SG with an eccentric fully regulated flange appositely prepared [13], [14]. The method can be applied to armature windings with parallel-connected pole-phase-groups (PPGs, groups of seriesconnected coils of a phase under a pole), Fig. 1. This arrangement is often used in high-current low-voltage SGs, as on-board ship alternators [5]. Wave windings or unsymmetrical windings are not covered here. The SG stator model in (1) represents a symmetric three-phase winding with n=6p independent (unconnected) circuits (PPGs). vS=RSS iS d dt LSS i S d \u03c8SR dt . (1) In (1) vS, iS are stator voltage and current column vectors, RSS, LSS are stator resistive and inductive matrices, and d\u03c8SR/dt is a column vector containing the voltages induced by the rotor", " Bs=Bhcos t\u2212 pS Bs 2 cos t\u2212 p\u22121 S\u2212s Bs 2 cos t\u2212 p1 Ss . (5) They can be thought as produced by two 'virtual' rotors with 2(p\u00b11) poles and different speed \u03c9/(p\u00b11), Fig. 5. The correspondent stator-linked fluxes are readily obtained by integration: \u03c8k=hcos t\u2212 pk\u22121 s , p\u22121 cos t\u2212 p\u22121k\u22121 \u2212s s , p1 cos t\u2212 p1k\u22121 s (6) with amplitudes s , p\u00b11=\u03c1s h K p\u00b11/2 (Fig. 6). The additional flux linkage systems in (6) have 2(p\u00b11) poles and no current can be stimulated in the load. Their induced voltages only produce internal current circulation [7] as in Fig. 1. With rotor center dynamically displaced, Fig. 7, the flux density wave amplitude is modulated both in space and in time as in (7). The maximum flux variation is \u0394Bd=\u03c1dBh. Bd=Bh Bd cos t / p\u2212S\u2212d \u22c5cos t\u2212 pS . (7) Decomposition of (7) again yields two new 2(p\u00b11)-pole waves, as produced by two virtual 2(p\u00b11)-pole rotors running with the same speed \u03c9/p, Fig. 8: Bd=Bh cos t\u2212p S Bd 2 cos \u2212 p t\u2212 p\u22121S\u2212d Bd 2 cos p t\u2212 p1Sd .(8) The stator-linked fluxes are carried out as: \u03c8k=hcos t\u2212 pk\u22121 d , p\u22121 cos \u2212/ p t\u2212 p\u22121 k\u22121\u2212d d , p1 cos / p t\u2212 p1 k\u22121d (9) where d , p\u00b11=\u03c1d h K p\u00b11/2 , see Fig", " They are defined as follows: \u03c8 p = n 2 h e j t (13) \u03c8 p\u22121 = n 2 s , p\u22121e j t\u2212 S (14) \u03c8 p1 = n 2 s , p1 e j t S (15) \u03c8 \u2212/ p p\u22121 = n 2 d , p\u22121 e j \u2212/ p t\u2212 d (16) \u03c8 / p p1 = n 2 d , p1 e j / p td . (17) Time-derivative of fluxes (13)-(17) are applied in (10), so obtaining the five sequence circuits displayed in Fig. 10. They represent the five virtual centered-rotor machines actually superimposed in the model of the real eccentric-rotor machine. Since model (1) is unconnected, the parallel connection among PPGs as in Fig. 1 must be applied through current and voltage constraints, which must be transferred on the circuits in Fig. 10 by short-circuiting the terminals [7]. Finally, by solving the circuits, the fault-related (p\u00b11)-order currents are carried out as in (18), (19). They are theoretically independent from the load current i(p), and trace epitrochoidal loci on the complex plane. i p\u22121 =i p\u22121 i \u2212/ p p\u22121 = = \u2212nh L p\u22121/K p\u22121 [ \u03c1s e j t\u2212 S\u03c1d e j \u2212/ p t\u2212 d ] (18) i p1 =i p\u22121 i \u2212/ p p\u22121 = = \u2212nh L p1/K p1 [ \u03c1se j tS \u03c1d e j / p td ] " ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000285_s1068798x1306018x-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000285_s1068798x1306018x-Figure1-1.png", "caption": "Fig. 1. Composite gear.", "texts": [ " The rings have plane end surfaces for fric tional interaction with the gears and are attached to the rim (for example, by pins) so as to permit only axial motion. Practical adoption of this design is prevented by the following defects: (1) the difficulty of machining manipulating the pins at the external and internal sur faces; (2) the large axial force required and hence the considerable size of the spring device; (3) the increase in radius of the composite gear with internal engage ment at the pin depth. To eliminate these problems, we have designed the gear in Fig. 1, which consists of the following compo nents: a rim 1; individual gear crowns 2 with internal (type I) or external (type II) engagement, and inter \u2020 Deceased. mediate rings 3; and a spring device 4 to ensure axial compression of the system. The intermediate rings 3 with radial slots 5 are characterized by frictional inter action with the lateral surfaces of the individual gear crowns 2 and the rim 1. The lateral surfaces of the crowns 2 and the rings 3 are conical; the contact sur faces of the rim 1 and rings 3 are cylindrical" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003240_978-0-8176-8370-2-Figure8.10-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003240_978-0-8176-8370-2-Figure8.10-1.png", "caption": "Figure 8.10: As in Figure 8.3 (G-case) but with B = 0.02, E1 = 0.004, and E3 = 0. Notice that now the dynamical variables are rotated.", "texts": [], "surrounding_texts": [ "When electric and magnetic fields are orthogonal to one another, the Stark\u2013 Quadratic\u2013Zeeman system is said to be crossed. The two vectors \u2212\u2192 WS and\u2212\u2192 WD belong, as in the general case, to the plane of the two physical fields, but in this case they also have the same norm and form the same angle, with opposite sign, with respect to the magnetic field: \u03d1S = \u2212\u03d1D. The first order averaged Hamiltonian is reduced to a term proportional to the rotated G3, so that the rotated Delaunay action-angle variables are recommended. The corresponding frequencies will be denoted \u03c9L,\u03c9G,\u03c9G3 , respectively. Because \u03b13, s+, c\u2212 are null, the coefficient c also vanishes, and the Hamiltonian (8.1.3) is similar to that of the QZ problem: the intersection parabolas on the plane \u03be1\u03be3 are symmetric with respect to the \u03be3 axis, whereas the concavity value \u2212b/a is no longer constant. The SQZc problem was well studied in the 1990s, and even confining ourselves to the classical (i.e., non-quantum) problem the number of articles is huge. We quote Von Milczewski & Uzer (1997a), Von Milczewski & Uzer (1997b), and Cushman & Sadovski\u00ed (2000) as a small sample and refer to the bibliography in these articles. Here we follow closely Cordani (2008). 8.1 The Stark\u2013Quadratic\u2013Zeeman (SQZ) Problem 247 248 Some Perturbed Keplerian Systems In Figures 8.9 and 8.10 we consider the case B = 0.02, E1 = 0.004, and E3 = 0, with G3 = 0.2, where it is understood that \u2212\u2192 R and \u2212\u2192 G have been rotated, as explained previously. The close resemblance with the corresponding Figures 8.1 and 8.3 of the QZ case shows how the rotation of the two 2-spheres is able to \u201cabsorb\u201d the presence of an orthogonal electric field. The value of the electric and magnetic field has been taken somewhat small, in order to restrict the amplitude of the fast oscillations in the numerical output. In the following numerical example, regarding the FMI distribution, we choose B = 0.2 and E1 = 0.06. In Figure 8.11 (top) the whole phase space is displayed, with the value of the total energy fixed to \u22120.5. Taking into account that \u03c3FMI \u2248 \u22125 is approximately the threshold below which the corresponding motion is in practice regular, one clearly recognizes three zones. The right-top corner always corresponds to the relative equilibrium point (since we are considering rotated variables) and is filled by KAM tori; the left-top corner displays a very sharply delimited Chirikov zone containing some small islands of stability; lastly, the central area is occupied by the Arnold web, and some distorted images of Figure 3.2 appear. Figures 8.11 (middle and bottom) show some details with a better resolution. The Arnold web is clearly visible, along with the chaotic area at the crossing of the resonances. These pictures show how more and more resonances are highlighted by improving the resolution and increasing the integration time. Notice, in particular, the bottom-right picture which displays a detail inside a resonance, showing the auto-similarity of the structure with an emerging Arnold web made up of secondary resonances. Clearly, with an even better resolution and a longer integration time, one would be able to display the Arnold web made up of resonances of the third level, and so forth. The FMI is therefore a very sensitive indicator of the regularity of the orbit and very efficient in showing the position of the resonances. In order to investigate fine details and to get numerical information, it may be convenient to compute the fundamental frequencies along horizontal or vertical sections. See, for example, Figure 8.12 where\u03c9L only is reported; the other two frequencies carry basically the same information. The resonances are clearly recognized in the top picture and exactly confirm the position given by the computation of \u03c3FMI in the previous figures. The middle and bottom pictures display further details, and a secondary resonance is clearly enlightened. Applying the frequency analysis to an orbit starting inside a resonance, one is able to compute numerically the three fundamental frequencies, then the resonance vector satisfying the relation k \u00b7\u03c9 = 0. For example, k = (0,3,1) for the resonance including the point G = 0.5, G3 = \u22120.34, and k = (1,2,\u22124) for the resonance including G = 0.7, G3 = \u22120.30, with the argument of pericenter equal to \u03c0/2. In the third window of KEPLER define the two unimodular transformations M1 = \u239b\u239c\u239d1 0 0 0 3 1 0 2 1 \u239e\u239f\u23a0 , M2 = \u239b\u239c\u239d 1 \u22121 0 1 2 \u22124 \u22121 \u22121 3 \u239e\u239f\u23a0 for the two points, respectively; then, by clicking at the end of the computation on \u201cUser Functions\u201d and choosing AADelaunayprime_2.m, one can see that the resonance angle librates with a low frequency. Notice that the second row of M1 and M2 coincides with the resonance vectors k and, consequently, the second of the files AADelaunayprime_x.m has been chosen; the first and third rows complete the two unimodular matrices. Another interesting detail is found in Figure 8.13, which shows a very magnified section of the stochastic layer surrounding the (0,3,1) resonance. The dynamical evolution of the resonant angle of an orbit starting here is displayed in Figure 8.14, where circulating and librating motions alternate with each other at random. This is the source of chaos. In Figure 8.15 we show some typical examples of the different orbits 252 Some Perturbed Keplerian Systems occurring in quasi-integrable Hamiltonian systems. The values of the frequency \u03c9L are reported for long integration intervals; notice, however, the different scales. The top-left picture (2 \u00b7 106 revolutions) refers to a KAM orbit, starting in the very neighborhood of the relative equilibrium point, with rotated and normalized values G = 0.995 and G3 = 0.990. Notwithstanding the great magnification, the frequency appears constant. The top-middle picture (2 \u00b7 106 revolutions) refers to a motion starting in a \u201cdark blue zone\u201d of Figure 8.11 (middle-right). The frequency evolution shows a slight oscillation but is very regular, denoting a KAM orbit in the outer neighborhood of a resonance. The top-right picture (108 revolutions) refers to a regular resonant orbit, the rotated and normalized values being G = 0.5 and G3 = \u22120.34. Compare with Figure 8.11 (middle-right). In the bottom-left picture (5 \u00b7 107 revolutions) the double resonant case is reported. The starting point is G = 0.6 and G3 = \u22120.4, and comparing with the values given by the frequency map (not given here) one sees that the whole chaotic zone at the crossing of the resonance strips is in fact explored. In the bottom-middle picture (108 revolutions) an orbit in the stochastic layer is plotted, starting from G = 0.5 and G3 = \u22120.3318. Looking at figure 8.13, one clearly ascertains that the computed evolution stays permanently within the thin stochastic layer surrounding the resonance, but the most interesting information is the slow and random drift of the whole pattern that only this orbit exhibits, very likely a manifestation of the Arnold diffusion. Lastly, in the bottom-right picture (107 revolutions) an orbit of Chirikov type is plotted. Notice the remarkable and random variation of the frequency, indicating that the orbit visits a large zone of the phase space. Figure 8.16 exhibits the wavelet transform of the six orbits of the previous figure, in the same order. Since Figure 8.11 (top) covers the whole phase space and the equilibrium point we have considered takes up the top-right corner, one may ask: But what about the other three equilibrium points? In the rotated spheres the equilibrium point we have considered lies in the North-North poles, then the remaining three lie in the South-South, North-South, and South-North poles, respectively. For these two last points the signs of the two frequencies are different, thus they may be unstable. In fact, it seems impossible to find a periodic orbit numerically, so their instability is very probable; in Figure 8.11 they are both placed at the vertex G = G3 = 0, which is a singular point in the Delaunay parametrization, encompassing all collision orbits. The point in the South-South poles is instead surely stable and takes up the top-left corner in Figure 8.11 where, however, it appears at first sight immersed in the chaos of the Chirikov zone. But a much more detailed numerical analysis shows in effect a very small KAM zone around this point to which there corresponds a periodic orbit in the physical space; see Figure 8.17. This physical orbit exhibits however two (somewhat surprising) sharp corners, making it very different from an ellipse; the osculating parameters undergo a marked variation, so that the corresponding periodic orbit on S2 \u00d7 S2 will be very large. 254 Some Perturbed Keplerian Systems 8.2 The Non-Planar Circular Restricted Three-Body Problem 255" ] }, { "image_filename": "designv11_101_0001033_s10883-013-9180-9-Figure8-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001033_s10883-013-9180-9-Figure8-1.png", "caption": "Fig. 8. Synthesis of optimal trajectories", "texts": [ " Therefore, both of this and the initial trajectory of type IIb0 can be discarded. So, only the trajectory OMN of type IIc0 remains, which is then optimal. (Note that for the points lying on the ray AA\u2217 , trajectories OMN and OK \u2032N coincide one with another.) Thus, for all three sets Z2, Z3, Z4 , the optimal trajectories are unique and belong to type IIc0: first the point moves along the right circle, and then along its tangent line. The above consideration allows us to construct a synthesis of optimal trajectories (Fig. 8). Let be given a terminal point p = (xT , yT ) \u2208 R2 + , p 6= 0. Theorem 6. If p lies inside Z1 or on the arc (AD), i.e., (xT \u2212 1)2 + y2T < 1, yT > 0, (xT + 1)2 + y2T 6 5, then the optimal control is (u, v) = ((\u22121, 1), (1, 1)) with one switching point. If p lies on the semi-interval (OD], i.e., xT 6 \u221a 5\u2212 1, yT = 0, then there are exactly two optimal trajectories: (u, v) = ((1,\u22121), (\u22121,\u22121)) and (u, v) = ((\u22121, 1), (1, 1)), symmetric to each other w.r.t. the horizontal axis and giving the same value of the cost" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000137_beiac.2013.6560247-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000137_beiac.2013.6560247-Figure1-1.png", "caption": "Fig. 1. Geometry of journal bearing with partial slip configuration", "texts": [ " A modified form of Reynolds equation is derived for couple stress fluids considering partial slip configuration. The nondimensional pressure and shear stress expressions are obtained. Reynolds boundary conditions are used to solve the pressure distribution. Results of non-dimensional load capacity and coefficient of friction in one dimensional journal bearing with partial slip under steady state are analyzed for different values of nondimensional slip coefficient ( ) and couple stress parameter ( ). The schematic of journal bearing with partially slip configuration is shown in Fig. 1. The partially slip extent is . In the present analysis, the variation of pressure across the fluid film is assumed to be negligible and pressure in the journal bearing is a function of sliding direction. Reynolds boundary conditions are incorporated in the analysis. 978-1-4673-5968-9/13/$31.00 \u00a92013 IEEE 807 The equations of motion for Stokes\u2019 couple stress fluid film (0 ) are (1) The boundary conditions for velocity used in the analysis are: Navier slip boundary conditions are imposed on the part of bearing surface, no slip conditions are imposed on the other part of the plain bearing surface and on the journal surface, and the couple stresses vanish at the interface" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000719_j.arcontrol.2011.10.010-Figure7-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000719_j.arcontrol.2011.10.010-Figure7-1.png", "caption": "Fig. 7. A simple model of a hopping robot.", "texts": [ " 6.2. Example 3. Time varying feedback stabilizing control for a hopping robot in flight phase A simplified kinematic model of a hopping robot in the flight phase can be given in the form (see Murray & Sastry (1993)): _w \u00bc u1 _l \u00bc u2 _h \u00bc m\u00f0l\u00fe d\u00de2 I \u00fem\u00f0l\u00fe d\u00de2 u1 \u00f086\u00de The configuration variables w, l and h have the following description: w is the angle of the hip of the hopping robot in the flight phase, l the length of the leg extension, and h is the angle of the body of the robot, as shown in Fig. 7. The remaining symbols represent constants: I is the moment of inertia of the body, m is the mass of the leg concentrating at the foot, and d is the upper leg length. Assuming for simplicity that m = I = d = 1, and introducing a new set of state variables x = (x1,x2,x3) = (w, l + 1,h), the kinematic model can be written as: _x \u00bc f1\u00f0x\u00deu1 \u00fe f2\u00f0x\u00deu2; x 2 R3 \u00f087\u00de where; f 1\u00f0x\u00de \u00bc @ @x1 x2 2 1\u00fe x2 2 @ @x3 ; f 2\u00f0x\u00de \u00bc @ @x2 Calculating f3\u00f0x\u00de \u00bcdef\u00bdf1; f2 \u00f0x\u00de \u00bc 2x2 \u00f01\u00fe x2 2\u00de 2 @ @x3 f4\u00f0x\u00de \u00bcdef\u00bdf2; \u00bdf1; f2 \u00f0x\u00de \u00bc 2 6x2 2 \u00f01\u00fe x2 2\u00de 3 @ @x3 verifies the LARC spanff1\u00f0x\u00de; f2\u00f0x\u00de; f4\u00f0x\u00deg \u00bc R3 for all x 2 R3 \u00f088\u00de few first terms of the Lie bracket multiplication table for L(f1, f2) are \u00bdf1; f2 \u00bc f3 \u00bdf1; f3 \u00bc 0 \u00bdf2; f3 \u00bc f4 \u00bdf1; f4 \u00bc 0 \u00bdf2; f4 \u2013 0 \u00f089\u00de and it is easily checked the Lie algebra L(f1, f2) is neither nilpotent nor finite dimensional" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002317_978-94-007-4201-7_5-Figure5.3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002317_978-94-007-4201-7_5-Figure5.3-1.png", "caption": "Fig. 5.3 A spatial serial chain", "texts": [ " @2U @\u2019N@\u20191 @2U @\u2019N@\u20192 @2U @\u2019N@\u2019N 2 66666666666664 3 77777777777775 2 R3 N N (5.8) The second-order KIC matrix is also called Hessian matrix. Each element is a vector with three components as follows \u20acUk \u00bc _f T \u00bdHk _f\u00fe \u00bdG k: \u20acf k \u00bc 1; 2; 3 (5.9) The second-order KIC matrix can be considered a 3-D cubic matrix, as shown in Fig. 5.2. Each level is a scalar matrix. Equation (5.7) can also be rewritten as \u20acUk \u00bc _\u2019T \u00bdHk _\u2019\u00fe \u00bdG k: \u20ac\u2019 k \u00bc 1; 2; 3 The analysis of a serial chain [4] is the basic requirement for parallel mechanism. Figure 5.3 shows a typical serial chain consisting of n + 1 links, as well as n kinematic pairs. Figure 5.3 illustrates the respective kinematic pairs S1S1, S2S2 . . .. . .SjSj, and the corresponding common perpendicular vectors a12a12, a23a23. . .. . .aijaij. The corresponding offset is Si, the rotation angle is yi , and the twist angle is aij . The global coordinate system is O-XYZ and the local system is Oi-XiYiZi . The axis Zi is along Si, Xi along the common normal aij. Ti indicates the transformation matrix of the system i with respect to the global system. Vector a \u00f0i\u00de ij \u00bc fX\u00f0i\u00de ij Y \u00f0i\u00de ij Z \u00f0i\u00de ij gT indicates vector aij with respect to Oi-XiYiZi", "18) where _\u2019i is the ith component of generalized velocity vector _f . The ith column vector of [G] is expressed as [G]:i. Gh \u2019 h i :i \u00bc @vh @ _\u2019i \u00bc @vx h @ _\u2019i @vy h @ _\u2019i @vz h @ _\u2019i T (5.19) Where \u2018:i\u2019 indicates the ith column of matrix [G]. KIC can be obtained by taking the partial derivative of Eq. (5.16) with respect to _\u2019 [4] Gh \u2019 h i :i \u00bc Si in 8< : (5" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000444_amr.837.316-Figure7-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000444_amr.837.316-Figure7-1.png", "caption": "Fig. 7. a) Equivalent stress distribution for the cam and cam follower assembly and b) The equivalent stress distribution in the cam for the cam and cam follower assembly", "texts": [ " The maximum value of the stress is 11,69 MPa and is located on the ages of the cam in the coating. In Fig. 6. a) is presented the normal stress distribution. For the \u03c31 compression component the maximum obtained value is 2,35 MPa and for the \u03c32 traction component de maximum value is 3,38 MPa. The shear stresses distribution in the cam with coating for the cam and cam follower both with coating case is presented in Fig. 6. b). The \u03c41 component has the value of 1,65 MPa and the \u03c42 component has the maximum value of 1,42 MPa. The stress distribution for the model without coating. In Fig. 7. a) is presented the equivalent stress distribution in the assembly between cam and cam follower. Fig. 7. b) shows the Von Mises equivalent stress distribution in the cam. The maximum value of the equivalent stress is 3.38 MPa. The normal stress distribution in the cam is presented in Fig. 8. a). The normal stress in this case has only one component (only the \u03c31compression component, the \u03c32 component has a very small value). The maximum value of the \u03c31compression component is 1,23 MPa. In Fig. 8. b) is presented the shear stress distribution in the cam. The maximum value of the shear stress is 0,53 MPa" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001674_2011-01-0476-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001674_2011-01-0476-Figure2-1.png", "caption": "Figure 2. Schematic of 3D profilometer trace using the P-15 profiler", "texts": [ " These operating ranges were determined by the non-contact welding method [10]. It is important to note here that the range of minimummaximum powers was different for each CB level. CB levels control the depth over which the laser energy is absorbed. Defect size measurements The deformation of the black Class-A surface in the zdirection directly opposite the weld line was measured using a KLA-Tencor P-15 profilometer. This stylus-based tracing system made 9 traces on the black surface along the x-axis as shown in Figure 2. The parallel traces were equally spaced over the 35 mm long weld line and were numbered sequentially along the scan direction from 1 to 9. Each trace was 10 mm long in order to capture deformation located 5 mm on either side of the center of the weld line. The profilometer trace speed was 1000 \u00b5m/s at a sampling rate of 10 Hz. A total of approximately 100 points were therefore collected for each trace. No measurements were made on the transparent part. A fixture was used to ensure the assembly was placed in the same position on the profilometer stage for each measurement" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002030_ijhvs.2013.053008-Figure20-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002030_ijhvs.2013.053008-Figure20-1.png", "caption": "Figure 20 MB vehicle model", "texts": [], "surrounding_texts": [ "Once the driveline torsional characteristics have been tuned by means of the previously described lumped parameters model, a complete vehicle model, including the driveline, has been implemented using VI-Rail (see Figures 20 and 21). The model has more than 80 dofs and it is able to reproduce both the running and the comfort behaviour of the vehicle. Ad hoc templates for bogies and wheelsets have been implemented to take into account the torsional deformability of the driveline. In particular each shaft/axle has been divided into three rigid bodies connected by bushings (Figure 21). The stiffness and damping characteristics of the bushings have been tuned according to the previously performed lumped parameters analysis. This approach has been preferred with respect of introducing deformable bodies into the model in order to speed up the simulations while maintaining a good accuracy. Within VI-Rail, wheel-rail contact geometry is introduced by means of pre-calculated nonlinear tabular elements (Kik et al., 2000), while the normal force is calculated as a function of the compenetration between the wheel and the rail (Piotrowski and Kik, 2008). Once the contact geometry and the normal load acting on each wheel are determined, the creepage and the creep forces are calculated. Several models are supported by VI-Rail: SHE, Polach, etc. (see Section 2). In order to speed up the simulations, creep forces are generally determined by means of pre-calculated nonlinear tables that introduce the creep force-creep curve correspondent to each model. As already mentioned, for the purposes of this paper, the SHE model has been selected. The procedure implemented for correcting the creep forces calculated by the SHE model so that thermal effects in the contact area are taken into account as will be discussed in the next section (Section 6)." ] }, { "image_filename": "designv11_101_0001873_phm.2011.5939468-Figure5-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001873_phm.2011.5939468-Figure5-1.png", "caption": "Figure 5. Schematic diagraph of a hole in the gear (a) and its finite element model (b)", "texts": [ " At the same time, the slot must be kept small enough to avoid causing the outer race to fail. Cutting a slot will cause an increase in both stress and deflection, and deflection is the key limiting factor for its structural integrity. So a finite element-based simulation can be used to determine the optimal slot size (length, width and depth), as shown in Fig. 4 (b). width depth (a) 978-1-4244-7950-4/11/$26.00 \u00a9 2011 IEEE MU3007 2011 Prognostics & System Health Management Conference (PHM2011 Shenzhen) For gears, we can embed a miniaturized sensor module by drilling a hole as Fig.5 (a). Similarly, it needs to determine the location of the hole from the center of the gear and its diameter and depth, so that drilling such a hole just causes an acceptable impact on the structure integrity. Here natural frequencies and modal shapes of the gear are two important factors. Next we consider six bearings and four gears as ten candidate locations of embedded sensor modules. Then an optimal model is built based on quantitative signed directed graph (QSDG) of fault vibration propagation, which is used to determine how many and where sensors to be embedded optimally" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001177_detc2011-47574-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001177_detc2011-47574-Figure2-1.png", "caption": "Figure 2: Pad balance schematic diagram", "texts": [ " When 0\u03b5 \u2032 < \uff0c ( ) 24 6 sin 6cosRS q \u03c6 \u03c6= \u2212 \u2212 . Through the way above, scopes of the parameter \u03b5 \u2032 and \u03b8 \u2032 in Reynolds equation are changed to ( ) ( )1 2, 1,1q q q \u2208 \u2212 . The three parameters ( )0,0.95\u03b5 \u2208 , ( )0 ,360TH \u2208 o o and ( ) ( )1 2, 1,1q q q \u2208 \u2212 are used to build up the database of a single pad. When obtaining the oil-film force, \u03b5 ,\u03b8 , \u03b5 \u2032 and \u03b8\u2032 can be obtained by Equations (5). 2 2 sin cos arctan sin cos j j j j j j j j Y X X Y X Y X Y \u03b8 \u03b8 \u03b8 \u03b8 \u03b5 \u03b8 \u03b8 \u03b5 \u2032\u23a7 = \u2212 \u23aa =\u23aa \u23a8 \u2032 = \u2212 \u2212\u23aa \u23aa = +\u23a9 & & & & (5) Figure 2 is pad balance schematic diagram. Considering the moment of inertia, the pad angular acceleration is calculated by Equation (6). The swing angular velocity of a pad is attained by integrating the angular acceleration of a pad. The angular displacement of a pad is attained by integrating the swing angular velocity of a pad. Since the pad is regarded as one of the moving parts, the pad angular motion equation is coupled with the rotor dynamics equation. Obviously, the number of system dynamics equations is increased" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001902_icma.2011.5986278-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001902_icma.2011.5986278-Figure2-1.png", "caption": "Fig. 2 The coordinate system of 5-UPS/PRPU PMT with redundant actuation", "texts": [ " This parallel machine tool has the following prominent characteristics: The pose of the moving platform can be denoted by the direct analytical formula expression of joint variables of the PRPU limb; the pose of the cutter can be measured and calculated in real time if we install sensors on every joint of the PRPU limb; due to the milling moment can be endured by the middle PRPU limb, the driving force of the UPS limbs can be reduced, et al. 1945978-1-4244-8115-6/11/$26.00 \u00a92011 IEEE In order to improve the dynamic performance and the precision, redundant actuation is introduced into this parallel machine tool. Replacing the first prismatic pair of the passive PRPU limb with active one, a novel redundantly actuated PMT is constructed, as shown in Fig. 1. The mechanism diagram of the 5-UPS/PRPU PMT with redundant actuation is shown in Fig. 2. The fixed Cartesian frame {A}(OA-XAYAZA) is set up on the stationary platform. Point Ai (i=1,2, ,5) is the center point of the universal pair. Point A1 lies on the axis YA and is 0.78 meter from the point OA; the other four points Ai (i=2,3, ,5)are distributed evenly on a circle with the radius of 0.72 meter. The moving coordinate frame {B}(OB-XBYBZB) is set up on the center of the moving platform. Bi (i=1,2, ,5) is the center point of the spherical pair and they are distributed evenly on a circle with the radius of 0" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001172_ijmee.39.4.7-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001172_ijmee.39.4.7-Figure3-1.png", "caption": "Fig. 3 Impossible scenario of a force causing clockwise rotation of the cylinder.", "texts": [ " Of course this seems logical if the cylinder slipped on the surface. But would this really happen if the cylinder was to roll? The professor proclaimed that the situation could be simply understood according to the direction of application of the force (along with Newton\u2019s laws). The applied force acted to the right, so only a force to the right would act on the cylinder\u2019s center of mass, causing the cylinder to move to the right. This explanation did not seem convincing. If the cylinder were to roll to the right, it would be rotating clockwise, as in Fig. 3. The end of the rod would move to the left, since it would rotate clockwise further than the entire cylinder would translate along the surface. This suggests that if you pushed on the end of the rod with your fi ngers, and the cylinder moved to the right by rotating clockwise, the end of the rod would deadlock against your fi ngers. The forces would build up at the point of application until the cylinder was forced to slip. In other words, the cylinder could never rotate, since it would always have to slip" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002239_978-94-007-5006-7_3-Figure3.13-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002239_978-94-007-5006-7_3-Figure3.13-1.png", "caption": "Fig. 3.13 EAJs and singularity. (a) A constrained spherical joint. (b) Use of ZYZ EAJs (Singular configuration). (c) Use of YXZ EAJs (Nonsingular configuration)", "texts": [ " Discussion on the singularity of EAJs is beyond of the scope of this book, hence, no further discussion on how to avoid them is provided in this chapter. One may, however, be referred to Shuster and Oh (1981) and Singla et al. (2004) for the singularity avoidance algorithm. In reality, most of the physical joints have restricted motion, and hence, areas of gimbal lock stay outside the domain of the movement of joints. Typically, a spherical joint used in practice is constrained to move as shown in Fig. 3.13a. In such a situation, wise selection of EAJs helps in avoiding the singular configuration. For example, if one uses ZYZ EAJs as shown in Fig. 3.13b, singularity is encountered in the zero-configuration. This happens when axis of joint 1 coincides with that of joint 3. On the contrary, use of YXZ EAJs, as shown in Fig. 3.13c, has no singularity corresponding to the zero-configuration. Singularity occurs when joint 2 is rotated by 90\u0131 about X axis. However, such a configuration is never encountered due to the constrained movement of the spherical joint and hence singularity is avoided. The concept of EAJs can conveniently be used for the representation of a universal joint by using the composite rotations as discussed in Sect. 3.3.3. Table 3.8 shows EAJs for 2-DOF rotations by universal joints. Such EAJs representation will allow one to treat the velocity transformation relation in a similar way to that of a revolute joint as given by Eq" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002239_978-94-007-5006-7_3-Figure3.3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002239_978-94-007-5006-7_3-Figure3.3-1.png", "caption": "Fig. 3.3 A spherical joint represented by three intersecting revolute joints", "texts": [ " The concept is later adopted for a unified representation of revolute joint based rotations, i.e., revolute, universal and spherical, and used to develop an efficient fully O(n) recursive algorithm. A fully recursive algorithm is not possible if the representation of a spherical joint is done using the definition of Euler angles. This is highlighted in Sect. 6.1.2. As mentioned earlier, the EAJs are intersecting revolute joints that give Euler angle rotations. Architecture of an EAJ is shown in Fig. 3.3, where a spherical joint connects a moving link #M to a reference link #R. The spherical joint is described by using three intersecting revolute joints, where joint 1 connects real link #R to an imaginary link #1, whereas joint 2 connects two imaginary links #1 and #2, and joint 3 connects imaginary link #2 with a real link #M. Two coordinate frames OMXMYMZM and OR-XRYRZR are rigidly attached to links #M and #R, respectively. If these frames are denoted as FM and FR, the rotation matrix between these frames can be obtained by using any of the Euler angle sets defined in Sect", "7 provides an interesting conclusion. The symmetric EAJs ZYZ and ZXZ, and asymmetric EAJs YXZ are free from the requirement of multiplication of any constant rotation matrix. More specifically, only three sets of DH parameters are required to define these EAJs. Hence, one should use ZYZ and ZXZ EAJs if symmetric set is chosen for representing a three-dimensional rotation. On the other hand, YXZ EAJs is preferred if asymmetric set is chosen for the rotation representation of a 3-DOF joints. The spherical joint shown in Fig 3.3 connects the reference link #R with the moving link #M. In practice, a robotic system may have serial- or tree-type architecture with several multiple-DOF joints. This calls for a systematic numbering scheme for intersecting revolute joints and the associated imaginary or physical links. So, the use of Euler-Angle-Joints (EAJs) and the associated numbering scheme are presented here for the systematic representation of a spherical joint. For example, Fig. 3.12 shows a link, #(k 1), coupled to its neighboring link, #k, by a spherical joint, k, which has three rotational DOF" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001680_ccdc.2013.6561389-Figure6-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001680_ccdc.2013.6561389-Figure6-1.png", "caption": "Fig. 6: Collision interval calculation diagram Another is that the circular arc passed by the straightened mechanical arm and the line-shaped obstacle cross at a point, as shown in Fig.7. The cross point may be the endpoint P1(Fig.7(b))or P2 (Fig.7(a)), may be also the point P\u2019(Fig.7(c), Fig.7(d), Fig.7(e), Fig.7(f)) on the line-shaped obstacle. P1P2 in Fig.7(a) and Fig.7(d) don\u2019t meet the formula (1), which are processed as non-collision between the mechanical arm and the line-shaped obstacle. P1P2 (except for the endpoints) in Fig.7(b) and Fig.7(e) and P\u2019 P2 (except for the cross point) in Fig.7(c) and Fig.7(f) meet the formula", "texts": [ " + + \u2212++= z x zxl dzxl arctan 2 arccos 22 1 2222 1 1\u03b8 (2) \u03c0\u03b8 \u2212+\u2212\u2212+= d w dl zxdl ta 2 arcsin 2 arccos 2 1 2222 1 2 Assuming P1(z1, x1) and P2(z2, x2) are the two endpoints of the line-shaped obstacle, the straight-line equation is obtained: 2013 25th Chinese Control and Decision Conference (CCDC) 2653 ( )( ) ( )( ) ( ) ( )[ ] ( ) ( )[ ]( )21212121 121121 ,min,,max,,min,,max - xxxxXzzzzZ zzxXxxzZ \u2208\u2208 \u2212\u2212=\u2212 (5) The equation of circular arc passed by the straightened mechanical arm (the joint angle of small mechanical arm is zero) is: ( ) 4/2 2 2 21 22 wllXZ ++=+ (6) Combining and solving the formula (5) and the formula (6), P(z, x) is obtained, which include the following several cases: 1) P(z, x) has the unique solution The case includes two classes again, one is that the circular arc passed by the straightened mechanical arm and the line-shaped obstacle touch at a point, as shown in Fig.6. The tangent point may be the endpoint P1(Fig.6(a)) or P2(Fig.6(b)), may be also the point P\u2019(Fig.6(c)) on the line-shaped obstacle. But they don\u2019t meet the formula (1) and will be processed as non-collision between the mechanical arm and the line-shaped obstacle. (1), which can processed as collision between the mechanical arm and the line-shaped obstacle. (a) (b) (c) (d) (e) (f) Fig. 7: Collision interval calculation diagram 2) P(z, x) has the double solutions The case is that the circular arc passed by the straightened mechanical arm and the line-shaped obstacle have two cross points, as shown in Fig" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001788_pacc.2011.5979014-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001788_pacc.2011.5979014-Figure3-1.png", "caption": "Figure 3. Generalized Coordinates For The AltAz Telescope Mechanics", "texts": [], "surrounding_texts": [ "Altitude- Azimuth (Alt-Az) telescopes form a very important class of apparatus in space and aerial observation as mentioned in [5], [6] and [8]. From control perspective, the requirements of stable pointing of such telescopes towards objects present a nonlinear control problem. Many such telescopes are equipped with automated control mechanism for pointing and tracking as shown in [7], [9] and [10], based on given altitude and azimuth(bearing) angles of the target object. Further, the high accuracy requirement necessitates reliability and robustness of the control scheme. This paper presents a modeling and robust control strategy for the nonlinear control of such telescopes. The control problem is to reorient the telescope to the point of an object with any given alt-az combination, starting from any initial alt-az position of the telescope. The proposed strategy is based on feedback linearization of the nonlinear dynamics, followed by a robust control law for the resultant linearized system, obtained through an optimal control approach. The possible cancelation errors occurring in the feedback linearized stage are taken as uncertainties for the robust control of the linearized system. Overall robustness of proposed strategy for the nonlinear system is shown through numerical simulation." ] }, { "image_filename": "designv11_101_0001402_s13369-012-0232-3-Figure5-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001402_s13369-012-0232-3-Figure5-1.png", "caption": "Fig. 5 a The convex gear is created by the conical cutter. b The convex gear is created by the conical cutter. c The assembly model of the convex and the concave gears. d The pinion and the gear created by RP technology", "texts": [ " Substituting Equations (1)\u2013(6) into Equations (7) and (8), the equations of meshing can be written as: \u03c6i = (yi c yi c\u03b2 + zi czi c\u03b2)xi c j \u2212 (yi c yi c j + zi czi c j )xi c\u03b2 rpi (xi c j yi c\u03b2 \u2212 yi c j x i c\u03b2) (9) \u03b2 = 0 (10) The dimensional parameters of the gear mechanism are listed in Table 1. A computer program was employed to draw the complete profile of the gear with convex tooth and the pinion with concave tooth. Using Equations (1), (2), (6), (9) and (10) with i = 1, the first tooth of the gear was created. The other teeth were copied from the first tooth and rotated at its axis by 2cy\u03c0/N1, cy = 1, . . . , N1, where N1 is the number of teeth of the gear. The complete contour of the gear is shown in Fig. 5a. Using Equations (2), (3), (6), (9) and (10) with i = 2, the first tooth of the pinion was created. The other teeth were copied from the first tooth and rotated at its axis by 2cx\u03c0/N2, cx = 1, . . . , N2, where N2 is number of teeth of the pinion. The complete contour of the pinion is shown in Fig. 5b. The assembly model of the pinion and the gear is shown in Fig. 5c. Using the proposed mathematical model and rapid prototyping technology, the pinion and the gear created by a ring surface were obtained and shown in Fig. 5d. The stress analysis of the proposed gear drives was used to determine the contact stress between the pinion and the gear, which was created by the envelope of two-parameter family of the ring surfaces, as shown in Sect. 3. The finite element (FE) method [8] and general proposed computer program [9] were applied to analyze stress between gear pairs. The computer program ran on Windows XP operating system and was used to obtain the numerical solution for the contact stress analysis. The material of the pinion and the gear is steel C1020" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001167_amr.490-495.1191-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001167_amr.490-495.1191-Figure2-1.png", "caption": "Fig. 2 Strip and furnace roller finite element model in continuous annealing furnace", "texts": [ " Accordingly, the author established a dynamic simulation model, and conducted quantitative analysis of the effect of several main wave shapes on deviation in the continuous annealing furnace. Establishment of finite element model. Continuous annealing furnace has several zones of heating, soaking, slow cooling, rapid cooling and aging. In the furnace, strips move up and down driven by rollers, and because of the higher temperature of heating zone the deviation often happens in this zone in practical production. Consequently, via ANSYS this study established a strip-roller finite element model (Fig.2), which applies flat roller as the roller contour with diameter of 800mm to research the effect of wave shapes on deviation, in respect that the research focuses on the effect of strip wave shapes on deviation within the contact zone. In this model, it considers furnace roller element as rigid shell body in order to reduce both unit number and time for calculation [3,4]. Poisson ratio of strip and furnace roller is 0.22, density 7,800kg/m\u00b3, and the influence of temperature on Poisson is neglected" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000112_20120213-3-in-4034.00021-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000112_20120213-3-in-4034.00021-Figure1-1.png", "caption": "Figure 1: Altius Mk I", "texts": [ " Comparative study of control laws implemented in simulation is presented in this work and is aimed to be implemented in next upgrade of Aves autopilot. The performance of controllers is compared for full flight which consists of take-off, waypoint following and landing. The outer loop navigation and guidance used in the current version works on the assumption that the inner loop heading hold is of first order. The simulation set up provides the performance graph of inner loops as well as the overall effect on tracking a trajectory. The airframe used for evaluating the performance of different control system in this work is Altius Mk I (figure 1). It is propelled by gasoline engine in pusher configuration and can carry payload up to 6 kg. The payload bay is kept configurable for conducting missions with different payloads. The UAV has clocked over 100 flight hours and conducted missions up to range of 140 km and altitude 7,000 ft. Page 1 of 6 \u00a9 IFAC-EGNCA 2012 A state of the art UAV simulation platform has been developed along with autopilot. The structure of the simulation platform is shown in figure 2. The key features of the simulation platform are: sensor modeling for the IMU, GPS and pressure sensors used onboard, sensor fusion implemented on Aves, and control block" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001600_sisy.2011.6034297-Figure5-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001600_sisy.2011.6034297-Figure5-1.png", "caption": "Figure 5. Elastic line of link.", "texts": [ " (28) \u2013 87 \u2013 By superimposing the particular solutions (28), any transversal oscillation can be presented in the following form: )(\u02c6)\u02c6(\u02c6),\u02c6(\u02c6 1 ,1,1,1,11 tTxXtxy j jtojjjto \u2211 \u22c5= \u221e = . (29) Bernoulli wrote equation (29) based on \u2018vision\u2019. Euler and Bernoulli did not define the mathematical model of a link with an infinite number of modes, but Bernoulli defined the motion solution (shape of an elastic line) of such a link, which is presented in equation (29). Euler and Bernoulli left the task of a link modeling with an infinite number of modes to their successors. The equation of Bernoulli (29) (see Fig.5.) defines a geometrical position of any spot on the elastic body line 1\u02c6toy in direction 1y \u2013 axis, and in a direction of 1x - axis it would be a 1\u02c6tox coordinate which is also a geometrical size and it can be presented in an analogue way as well as the size 1\u02c6 toy . The position of a tip of a presented body with indefinite number of modes is defined by coordinates 1tox , 1toy in 1,1 yx level. It is supposed that all motions are made in 11 yx \u2212 level, and a coordinate is 01 =z in this case. Equation (29) is actually the solution of dynamics of the presented body\u2019s motion during the time. However, in order to calculate the coordinates 1\u02c6tox , 1\u02c6 toy in some specific moment of time (as is seen from Fig. 5) , it is necessary to know sizes of elastic deformations 1,1y , 2,1y , 3,1y ... jy ,1 and angles 1,1\u03c9 , 2,1\u03c9 , 3,1\u03c9 ... j,1\u03c9 of all modes defined in a space of local coordination system jijiji zyx ,,,,, . j,1\u03c9 is the rotation angle of the top of the same mode (see [9]). Generally, coordinates 1tox , 1toy are the total of elastic deformations, but precisely, in geometrical terms, it is the total of projected elastic deformations on axes 1x , 1y respectively. Equation (29) has a significance as elastic deformation for each mode for Meirovitch [8] and his followers and in this way defined is entered in the total dynamic model" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000382_s0278641913020088-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000382_s0278641913020088-Figure3-1.png", "caption": "Fig. 3. Scheme of the generalized method and area ERn.", "texts": [ " Conditions X < BBD[Z] thus prove to be false and the whole set of calculations is blocked. The algorithms then start to depend con siderably on the rounding error at ETA, which restricts their application. The situation is nevertheless not as bad as it might seem. METHOD GENERALIZATION The displacement method proposed above can be considered as representative of an entire class of methods that differ from each other in their parameters. Let us discuss one such parameterization method that generates the general displacement illustrated in Fig. 3a. Formally, the generalized method is described by a system of constraints on the values of parameters En and Rn.: Rln x; \u03b7( ) x Ai ki\u2013 i 2= \u03b7 \u220f 1\u2013 ki ALN i[ ], where ki 0 1 2, ,{ }.\u2208\u00d7 i 2= \u03b7 \u2211+= MOSCOW UNIVERSITY COMPUTATIONAL MATHEMATICS AND CYBERNETICS Vol. 37 No. 2 2013 This system is reduced to inequalities by equivalent transformations (5) Pairs (En, Rn) satisfying (5) form the triangular area ERn depicted in Fig. 3. Property 9 (a generalization of property 5). If and , then formula (6) and rela tionships (7) are true: (6) (7) After specifying the method for selecting En and Rn, the calculation for the logarithm using generalized displacement is given by the formula At some values of En and Rn, generalized displacement gains additional properties that can affect both the logarithm\u2019s area of application and our choice of directories. In particular, all ERn contain pairs ( , An + 1), where = is the middle of the half interval In" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001402_s13369-012-0232-3-Figure9-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001402_s13369-012-0232-3-Figure9-1.png", "caption": "Fig. 9 The geometric models of the rack cutter generated by the inverse envelope concept", "texts": [ " Thus, matrix Mc\u20322\u2032 can be written as: Mc\u20322\u2032(\u03c6\u2032 2) = \u23a1 \u23a2\u23a2\u23a3 1 0 0 0 0 cos \u03c6\u2032 2 sin \u03c6\u2032 2 rp2\u03c6 \u2032 2 0 \u2212 sin \u03c6\u2032 2 cos \u03c6\u2032 2 rp2 0 0 0 1 \u23a4 \u23a5\u23a5\u23a6 (13) To determine the envelope to the family of the pinion surfaces, the equation of meshing can be expressed as follows: f ( j ,\u03b1,\u03c6\u2032 2) = (y\u2032 2 j x \u2032 2\u03b1 \u2212 x \u2032 2 j y\u2032 2\u03b1) sin \u03c6\u2032 2 + (x \u2032 2 j z\u2032 2\u03b1 \u2212 z\u2032 2 j x \u2032 2\u03b1) cos \u03c6\u2032 2 \u2212 [ y\u2032 2(y\u2032 2 j x \u2032 2\u03b1 \u2212 x \u2032 2 j y\u2032 2\u03b1) + (z\u2032 2z\u2032 2 j x \u2032 2\u03b1 \u2212 z\u2032 2 j z\u2032 2\u03b1) rp2 ] = 0 (14) The values of x \u2032 2, y\u2032 2 and z\u2032 2 were obtained in Sect. 3. j is the design parameter as represented in Sect. 2 ( j = c, d, h). Using Equations (2), (3), (5), (6), (9)\u2013(11), (14) and Table 1, a mathematical model of the inverse envelope rack cutter was obtained, and together with a computer program, the cutting curves of the rack cutter were plotted, as shown in Fig. 9. A complete mathematical model of the pinion and the gear created by the ring surface and envelope theory of two-parameter family of surfaces was developed. A computer program was developed based on the proposed mathematical model. The developed computer program was applied to plot the geometric model of the pinion and the gear using a computer-aided design method. To demonstrate the contours of a pinion and a gear, rapid prototyping and manufacturing technology is used to make the gear pairs. Based on the inverse envelope concept and the engagement relationships between the pinion and the rack cutter, a mathematical model of the rack cutter for gear machining was established" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001260_s11465-011-0225-z-Figure5-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001260_s11465-011-0225-z-Figure5-1.png", "caption": "Fig. 5 The redundantly actuated PRPRP mechanism in the improved 6-DOF erector. The sections marked as joints A1 and B1 denote the two driving hydraulic cylinders, joints A2 and B2 are two revolute joints, load M denotes the segment, and joint C is the redundant actuator. The length of link A2 is constant, whereas that of link B2 is changeable", "texts": [ " In the practical control system, no matter what hydraulic elements are adopted (e.g., the synchronal valves or advanced controlling strategies), the precise synchronization of the two cylinders is difficult to achieve. 2.3 A redundantly actuated PRPRP mechanism in improved 6-DOF erectors This paper proposes a redundantly actuated PRPRP mechanism to implement the radial precise synchronization of the two hydraulic cylinders in the 6-DOF erector; the circumferential mechanism and the pose adjustment mechanism are not drawn (Fig. 5). Two degrees of freedom exist in this mechanism, satisfying the synchronization and pose adjustment in theory. When the redundant actuator (joint C in Fig. 5) is locked or produces enough pretightening tensile force, the PRPRP mechanism becomes equivalent to the PRRP mechanism (Fig. 3), thus ensuring the synchronization of the two driving hydraulic cylinders based on the mechanical structure. Moreover, the redundant actuator can apply two equal flexural torques at the hydraulic cylinders, thus preventing the overload of a single cylinder. Furthermore, the redundant actuation can alleviate the undesirable effects of clearances appearing in the mechanism joints. When the redundant driving joint C in Fig. 5 is locked or produces enough pre-tightening tensile force, the distance between joints A2 and B2 is kept constant. If some errors or disturbances occur in the control system (e.g., if the velocity of joint A1 exceeds the velocity of joint B1), then joint A1 immediately drags joint B1 through the rigid links A2 and B2; therefore, the synchronization of two hydraulic cylinders is guaranteed. After the pose adjustment of the segment, the PRPRP mechanism can easily restore its synchronous status. As link B2 restores its initial length, the radial mechanism also restores the synchronization", " The redundant actuator not only ensures the synchronization, but also applies two equal flexural torques to the driving hydraulic cylinders, thus preventing the overload of a single cylinder. Here, let us suppose that mA1, mA2, mB1, mB2, and md denote the masses of links A1, A2, B1, B2, and the segment, respectively; and LA1, LA2, LB1, LB2, and Ld denote the lengths of links A1, A2, B1, B2, and the distance between joint C and the segment, respectively. The segment is combined with link A2, and the angle between axes XP and XO in Fig. 5 is \u03b8. In addition, g denotes the gravity acceleration, and f denotes the redundant driving force of joint C. Suppose the two hydraulic cylinders perform the synchronous uniform motion, then the flexural torque of joint C is given by MC \u00bc LB2g\u00bdLA2\u00f02md \u00fe mA2 \u00fe mB2\u00decos \u00fe 2Ldmdsin 2\u00f0LA2 \u00fe LB2\u00de : (1) The flexural torque of joint A1 is MA1 \u00bc LA1 2 \u00bd\u00f02md \u00fe mA1 \u00fe 2mA2\u00deg sin \u2013 2f : (2) The driving force of joint A1 is FA1d \u00bc gf\u00bdLA2\u00f02mA1 \u00fe mA2\u00de \u00fe LB2\u00f02md \u00fe 2mA1 \u00fe 2mA2 \u00fe mB2\u00de cos \u2013 2Ldmdsin g 2\u00f0LA2 \u00fe LB2\u00de : (3) The flexural torque of joint B1 is MB1 \u00bc LB1 2 \u00bd\u00f0mB1 \u00fe 2mB2\u00deg sin \u00fe 2f : (4) The driving force of joint B1 is FB1d \u00bc gf\u00bdLB2\u00f02mB1 \u00fe mB2\u00de \u00fe LA2\u00f02md \u00fe 2mB1 \u00fe 2mB2 \u00fe mA2\u00de cos \u00fe 2Ldmdsin g 2\u00f0LA2 \u00fe LB2\u00de : (5) If MA1 = MB1 is required, then we express f as f \u00bc LA1\u00f02md \u00fe mA1 \u00fe 2mA2\u00de \u2013 LB1\u00f0mB1 \u00fe 2mB2\u00de 2\u00f0LA1 \u00fe LB1\u00de g sin : (6) According to the circumferential rotating angle \u03b8, as well as the lengths of links LA1 and LB1, the redundant driving force f can be controlled in real time to apply the same flexural torques to the driving hydraulic cylinders" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000205_vppc.2012.6422625-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000205_vppc.2012.6422625-Figure3-1.png", "caption": "Fig. 3. Normal forces acting on each tire of four wheels.", "texts": [ " A feature of this method is that the lateral force required for cornering is guaranteed by the following procedure that is based on each friction circle of the tires of the four wheels. First, after the lateral and longitudinal accelerations ax and ay, respectively, are detected, the lateral and longitudinal load movements Zx and zY' respectively, are calculated using Mearay\u00b7 Hear Zy = b . ([) Through the load movements obtained between the left and right wheels, the front and rear normal forces Fzji and FzJ;, respectively, (i = I: left; i = r: right) (Fig. 3) acting on the left and right tires of the front and rear wheels are compensated during cornering. For example, for a left turn, these forces are compensated as follows: [n this case, the lateral forces FYJi and FYJ; (i = I: left, i = r: right) that act on the left and right tires of the front and rear wheels, respectively, are expressed as follows: Here, FYJ and FYJ are the front and rear lateral forces, respectively, which are defined by the two-wheel vehicle model shown in Fig. 4. After each lateral force is obtained by using ([ )-(3), the maximum longitudinal force corresponding to each lateral force is obtained for the following procedure using the principle of a friction circle (Fig" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003263_b978-0-08-097016-5.00007-3-Figure7.4-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003263_b978-0-08-097016-5.00007-3-Figure7.4-1.png", "caption": "FIGURE 7.4 Characteristics of tire side force, \u2018intersection length\u2019 s a, and relaxation length sa.", "texts": [ " Fully Nonlinear Model In Eqn (7.7), the relaxation length sa is replaced by s a and we get dva dt \u00fe 1 s a jVxjva \u00bc jVxjtan a \u00bc Vsy (7.29) with apparently s a \u00bc 1 CFy Fy tan a0 \u00bc sao CFa Fy tan a0 z sao CFa jF0 yj \u00fe CFa3F jtan a0 f j \u00fe 3F (7.30) with the initial relaxation length (at a0 \u00bc 0): sao \u00bc CFa CFy (7.31) to which s a approaches when a0/0. To avoid singularity, one may use the last expression of (7.30) with small 3F and add to a0 the shift Da to arrive at a0f as indicated in Figure 4.22. Figure 7.4 presents the characteristic of the sliding range and decreasing \u2018intersection\u2019 length s*. intersection length together with the side force characteristic from which it is derived. Also in the equation for the camber deflection response (7.11), the relaxation length sa is replaced by s a. A more direct way to write Eqn (7.29) is yielded by eliminating s a with the use of Eqn (7.28): dva dt \u00fe jVxjtan a0 \u00bc jVxjtan a \u00bc Vsy (7.32) The transient slip angle a0 is obtained from the deflection va by using the inverse possibly adapted F0 y\u00f0a0\u00de characteristic, cf", " If the vertical load remains constant, the last term vanishes and we have the often used equation of the restricted fully nonlinear model: sa d tan a0 dt \u00fe jVxjtan a0 \u00bc Vsy (7.35) with sa \u00bc 1 CFy vFy v tan a0 (7.36) If we consider an average slip angle ao and a small variation ~a of tan a and the corresponding lateral slip velocities, Eqn (7.35) becomes, after having subtracted the average part, sa d~a0 dt \u00fe jVxj~a0 \u00bc ~Vsy (7.37) which indicates that the structure of (7.35) is retained and that sa (7.36) represents the actual relaxation length of the linearized system at a given load and slip angle. Its characteristic has been depicted in Figure 7.4 as well. Obviously, the relaxation length is associated with the slope of the side force characteristic. It also shows that the relaxation length becomes negative beyond the peak of the side force characteristic which makes the solution of (7.35) but also of the original Eqn (7.29) or (7.32) unstable if the point of operation lies in that range of side slip. We may, however, limit sa downward to avoid both instability and excessive computation time: sa \u00bc max(sa, smin). The transient response of the variation of the force proceeds in proportion with the variation of ~a0 as ~Fy \u00bc \u00f0vFy=vtan a\u00de~a0" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000643_1.4736880-Figure5-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000643_1.4736880-Figure5-1.png", "caption": "FIGURE 5. (a) Displacement measurement set-up, (b) Displacement characteristics of APA", "texts": [ " A diamond shaped Amplified Piezo Actuator (APA) fabricated using six multilayered piezo stacks produces maximum displacement of 173\u03bcm at 175V and the amplification factor of 4.3. Photographs of fabricated ML stacks and amplified actuators are presented in Fig.4a4b. The displacement of fabricated simple ML stack and the amplified actuator was measured. The actuator is placed on a plane rigid support on top of which the tip of the strain gauge is placed with an initial reading set to zero. A test set up measurement for displacement of the PZT stack actuators is presented in Fig.5a. The terminals of the actuator are connected to appropriate terminals of a dc source and the voltage is gradually increased. It is observed that the displacement increases with increase in voltage and a maximum displacement of 10\u03bcm and 173\u03bcm is measured for simple ML stack and amplified actuator respectively. The typical plot of the displacement vs. voltage of amplified actuator is presented in Fig. 5b. The simple ML stack is characterized for block force using a block force measuring unit. The actuator is placed on top of a force sensor (load cell) inside the sample holder and its positive and negative terminals are properly connected to the respective terminals of the voltage source. For measurement of block force, a constant pre-stress is applied from top of the actuator through 3-4 springs of different stiffness. The values of displacement and force generated by the actuator for all the springs are plotted by block force measurement software" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000981_isam.2013.6643451-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000981_isam.2013.6643451-Figure1-1.png", "caption": "Fig. 1 Unified coordinate system of fix-pad journal bearing", "texts": [ " Traditional model needs various geometric formulas which features heavy workload and is not conductive to expansion and promote of the program. A universal computing model unifies geometric describe of fix-pad journal bearing will greatly reduces programming workload and benefits expansion and promotion of performance calculation program. III. PERFORMANCE CALCULATION MODEL BASED ON A unified coordinate system is established, which takes the bearing center as its origin and takes the positive direction of y-axis as the angle starting line. As shown in Fig.1, O is the bearing center, iO is the center of pad i , jO is the shaft center, e is the eccentric distance and \u03b8 is the attitude angle, i\u03b4 is the eccentric distance of pad i , i\u03b2 is the angle between eccentric distance and angle starting line, name i\u03b2 the eccentric distance angle or preload angle in dimensionless. In the unified coordinate system coordinate of iO is: sin( ) cos( ) i i i i i i x y \u03b4 \u03b2 \u03b4 \u03b2 =\u23a7 \u23a8 =\u23a9 (1) Coordinate of jO is: sin( ) cos( ) j j x e y e \u03b8 \u03c0 \u03b8 \u03c0 = +\u23a7\u23aa \u23a8 = +\u23aa\u23a9 (2) In the unified coordinate system, eccentricity ratio and attitude angle of shaft center relative to pad center can be calculated with following equation: 2 2( ) ( ) a rc tan i i j i j i j i i j e x x y y x x y y \u03b8 \u23a7 = \u2212 + \u2212 \u23aa\u23aa \u23a8 \u239b \u239e\u2212 =\u23aa \u239c \u239f\u239c \u239f\u2212\u23aa \u239d \u23a0\u23a9 (3) Equation (3) calculates the eccentricity ratio and attitude angle of the shaft center relative to any pad, which is suitable for all forms of fix-pad journal bearing" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000246_amm.52-54.834-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000246_amm.52-54.834-Figure2-1.png", "caption": "Figure 2 The 5-DOF parallel manipulator and structural decomposition", "texts": [ " The kinematic compatibility equation of SOC{-S1-S2-} is shown as: 22 21 2 21 2 21 )()()( Lzzyyxx ssssss =\u2212+\u2212+\u2212 (7) The kinematic compatibility equation of SOC{-R1\u2225R2- R3\u2225R4-} can be obtained as follows: 0)()()( 2 2112 2 2112 2 2112 =\u2212\u2212\u2212+\u2212 ABABBCBCACAC (8) Where 3011 zx \u22c5= aA ; 30111 )( zxz \u22c5\u00d7= aB ; 31 zD \u22c5\u2212=C ; Dx \u22c5= 012 2aA ; Dxz \u22c5\u00d7= )(2 0112 aB ; DD\u22c5+\u2212= 2 3 2 32 aaC ; )( 1444332222 OOzzzxD \u2212++++\u2212= llla . The modular approach for kinematic analysis of complicated parallel kinematic manipulators (which coupled degree equal to 2) is introduced as follows, and demonstrated in a 5-DOF parallel manipulator. As shown in Fig.2, such mechanism consists of five limbs\uff081, 2, 3, 4, 5\uff09, and li (i=1, \u2026,5) denote length of the five corresponding limbs respectively. The first is composed of a revolute joint connecting the limb to base platform, then an actuated prismatic joint and a spherical joint connecting the limb to mobile platform. The rest limbs are identical and are composed of a spherical joint, then an actuated prismatic joint and a spherical joint. The joint centers Ai on the base and Bi (i=1, \u2026,5) on the mobile platform are located in circle shape, and R, r denote circumradiuses of the circles", "11, l4=1.27, l5=1.40, \u03b81=0\u00b0, where angular unit is degree (\u00b0) and other units are millimeters (m). The coordinates of joint centers Ai in the fixed coordinate system Ob-xbybzb and Bi in moving coordinate system Op-xpypzp are shown in Tab.3 respectively. Table 3 The coordinates of joint centers The coordinates of Ai in Ob-xbybzb The coordinates of Bi in Op-xpypzp A1 A2 A3 A4 A5 B1 B2 B3 B4 B5 x -R 0 R 0.6R -0.6R -r 0 r 0.6r -0.6r y 0 -R 0 0.8R 0.8R 0 -r 0 0.8r 0.8r z 0 0 0 0 0 0 0 0 0 0 As shown Fig.2, the 5-DOF parallel manipulator can be decomposed into five actuated prismatic joints and four SOCs. By the way, an additional revolute movement around axis B1B3 is generated because both the joint B1 and B3 are spherical joints, and we acquire an additional revolute joint Op in SOC2, as shown Fig.2. The detailed steps for such parallel manipulator are shown as below. Step 1 For the first component SOC1{-A1-B1-B3-A3-}, directly applying the model of SOC{-R-S1-S2-S3-} shown in Tab.2, the rotary angle of joint A1, denoted as \u03b2, can be determined by the following equation: 0sincos * 1 * 1 * 1 =++ cba \u03b2\u03b2 (9) The value of \u03b2 can be obtained by Eq.(9). Further, the positions of joint centers B1, B3 and Op can be easily computed respectively. For the SOC2{-A2-B2-Op-}, directly applying the model of SOC{-R-S1-S2-} shown in Fig" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003921_iros.2011.6048332-Figure6-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003921_iros.2011.6048332-Figure6-1.png", "caption": "Fig. 6 User posture.", "texts": [ " From the measurement results, in the case of ascending-stairs, the averages of the height and the width of the step are 186mm and 292mm, respectively. On the other hand, in the case of descending-stairs, the averages of the height and the width of the step are 169mm and 288mm, respectively. After all, the robot recognizes one step of stairs by using the laser range finder. B. Calculation of ZMP In this paper, the robot judges whether the user falls down by calculating ZMP. The ZMP is calculated based on human model as shown in Fig. 6. The supporting leg is found by tactile switches located in both soles. Then, the posture of the user\u2019s body region is calculated by these angles based on the supporting leg as follows. aspksphspb ,,, TTTT (9) where Tb is the angle of upper body region, Tsp,h, Tsp,k and Tsp,a are the angles of hip, knee and ankle joint of the support leg, respectively. Tsw,h, Tsw,k and Tsw,a are the angles of hip, knee and ankle joint of the swing leg. The positions and accelerations of the center of gravity (COG) of upper body, thigh and shin are calculated based on each joint angle" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003001_detc2013-12231-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003001_detc2013-12231-Figure1-1.png", "caption": "Fig. 1 Marine drive transmission with small down angle Since the beveloid gear was first proposed by H.E.Merrit[1] and A.S.Beam[2], the meshing theory, manufacturing and tooth contact characteristics had been investigated a lot. Hiersig[3]\u3001Sz\u00e9kely[4] and Roth[5] did a lot of work and laid foundation for the geometry design of beveloid gears. Mitome[6-10] conducted both analytical and experimental investigations on the concave conical gears. Brauer[11-13] theoretically derived the mathematical model of conical involute gears and", "texts": [ "or gear transmissions with small shaft angle (<45\u00b0), beveloid gears have an obvious mesh performance than the conventional worm gears and hypoid gears. A typical application for beveloid gears is the marine transmission with small shaft down angle as shown in Fig.1. However, the main weakness in this type of marine transmission is that the mesh for the intersected beveloid gears is theoretically in point contact. This unfavorable tooth contact behavior leads to a lower tooth surface durability and causes harmful vibrations. studied the transmission errors in anti-backlash conical * Corresponding author. E-mail: cczhu@cqu.edu.cn 1 Copyright \u00a9 2013 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/pdfaccess.ashx?url=/data/conferences/asmep/77583/ on 02/16/2017 Terms of Use: http://www" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001006_s1068799813040028-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001006_s1068799813040028-Figure1-1.png", "caption": "Fig. 1.", "texts": [ " (32) By differentiating relations (31) with respect to the global coordinates s and \u03b8 , we determine by analogy the components of the first and second derivative of the step-by-step displacement vector. By transforming Lagrange\u2019s functional (20) taking into account (12), (16), (32) and minimizing it with respect to the step-by-step nodal unknowns { }yW \u0393 , we can obtain the following matrix relation: [ ]{ } { } { } 1 y yK W f R k \u0393 = \u2212 , (33) where [ ]K and { }yf is the stiffness matrix and the column of nodal forces at the step of loading; { }R is the Newton\u2013Raphson correction. EXAMPLE OF CALCULATION As an example (Fig. 1) we consider the stress-strain state of a shell of revolution, the radius of which is specified by the functional dependence ( )( ) cosr x A B x C= + . We accept the following initial data: KLOCHKOV et al. RUSSIAN AERONAUTICS Vol. 56 No. 4 2013 332 q = 0.2 MPa, the shell thickness is t = 0.01 m; the modulus of elasticity is 52.06 10E = \u00d7 MPa; Poisson\u2019s ratio is 0.3\u03bd = ; A = 1.3 m; 0.4 mB = . The left edge of the shell was hinged and the right one was simply supported. The calculations were carried out in two variants" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003689_s1474-6670(17)70081-x-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003689_s1474-6670(17)70081-x-Figure2-1.png", "caption": "Figure 2. Wave~rorms and phase plane trajectory in transient state: (a) waveforms; (b) phase plane trajectory", "texts": [], "surrounding_texts": [ "On the Subharmonic Oscillation of the Relay Servomechanism J. YAMAGUCHI, M. NISHIMURA, K. FUJII and T. MARUHASHI\nIntroduction\nIn the frequency response of a non-linear feedback control system, the output frequency is not always equal to the input frequency, but may become a fraction of the latterl - 3.\nAs an example of such a subharmonic oscillation, a servo system, controlling a velocity lag linear element by an ideal relay, was analysed in detail by solving linear differential equations4 successively expressing the system behaviour in each linear portion of the relay characteristics. Some interesting features of the subharmonic oscillation for this system were obtained by such an analysis. For example, when the input versus output characteristics of the relay are symmetrical, the output frequency is equal to I ln of the input frequency, where n is an odd integer, and the order n of the possible subharmonic oscillation is determined analytically, if the frequency and magnitude of the input sine wave and relative values of the system parameters are given. It cannot be denied, however, that this orthodox method is very troublesome and impractical.\nConsidering the subharmonic components existing in these closed loop signals, there is a time delay between those input and output components of the relay element owing to the effect of the input sine wave. It is possible to represent the delay caused by input sine wave by a time delay element and consider that the system is under free oscillation instead of the corresponding forced subharmonic oscillation. Thus this problem can be solved as that of a free oscillation by the well known Describing Function Method5\u2022 This approximate method is much simpler than the orthodox one described above and the results obtained coincide well quantitatively with those obtained by the orthodox method.\nExact Analysis Based on Differential Equations\nFundamental equations\nWith respect to the error signal ee, the equations describing the dynamics of the relay servo system shown in Figure 1 are\nTO. + e. = TO i + ei - KA, (ee > 0, ec = +A) (l a) TO. + O. = TO i + ei + KA, (ee < 0, ec = -A) (I b)\nThe system input ei is assumed to be\ne i = e im sin ((I) t + <1\u00bb (2) Put\ntiT = T, e ,IKTA = rP\" e ;/KTA = rPi\nd2rP./dT2 = ~e , drPeldT2 = ~e d2rPddT2 = ~ i' drP ddT = ~ i\nct) T = U\nthen equations I a, 1 band 2 become the dimensionless represen tations as\nequation 3a expressing the behaviour of this system in the interval 9 <: T. <: Tl, under the condi~ions .that at T = 0, rP. = 0, CP. = CP' 1 and at T = T1, rP. = 0, CP. = CP.2, the following equations are obtained\n(1 + ~el - IIrP im cos <1\u00bb(1 - e- T ,) - rP im sin 1/\n- Tl + rP im sin (IIT1 + <1\u00bb = 0\n. r + UrP im cos (UTI + <1\u00bb = CPe2 J\n(5)\nSimilarly, in the interval Tl <: T <: T2, by solvi.ng equation 3b under the conditi(;ms th.at at T = Tl, rPe = 0, CP. = rPe2 and at T = T2, rPe = 0, CPe = CP e3' we obtain [~e 2 - I - IIrPim cos (IITl + <1\u00bb][1 - e - h - T,l] + (T2 I . - T1) + rP im[sin (UT 2 + <1\u00bb - sin (IIT1 + <1\u00bb] = 0 ~\n[CPe2 - I - IIrP im cos (IIT1 + <1\u00bb] e - h - T ,) + I . 1\n+ IIrP im cos (IIT 2 + <1\u00bb = CPe 3 J Conditions under which subharmonic oscillations occur\n(6)\nAfter the system has reached the steady-state, initial condi tions must be equal to final values in a complete cycle of the subharmonic oscillation, then\nCPel = CPe3 (7) and\nrP im sin = rP im sin (UT2 + <1\u00bb\n398", "ON THE SUBHARMONIC OSCILLATION OF THE RELAY SERVOMECHANISM\nthus T~ = 21/1T/1I, (1/ = 1,2,3 ... ,)\nwhere T' is the time when the error 4>e(T) becomes minimum in (8) the interval 0 < T < T1 in Figllre 3. Clearly, at that instant,\nT1 obtained from equations 5-8, is given by\nT1 = 1/1T/1I = T2/2\nFrom equations 5-8, <1> can be derived as follows\nsin <1> -sin (1/1T + <1\u00bb = 2 (tanh ;: - ;:) /4>'\"1\nLet 1/ be an odd integer (I, 3, 5, 7 ... ,), then\n<1> = -sin-1 [(;: - tanh ;:) /4>illlJ\n(9)\n(10)\n(11)\nSince 1/1T/2u > tanh 1/1T/2u in equation 11, <1> is always negative. Thus the input phase angle <1> to the fundamental (n = I) or subharmonic response (1/ 3) can be determined.\nIf 1/ is an even integer (2, 4, 6, 8 ... ,), we obtain from equation 10\n1/1T 1/1T tanh- - - = 0\n21t 2u ' or 1/1T = 0 21t\nThis result is irrational, and the subharmonic oscillations of the even order cannot occur in this system.\nApplying equations 9 and 11 to equation 5 gives for the initial condition\n. 1/1T ~e1 = tanh? + U4>illl cos <1> _u\n(12)\nThe solution of equation 3a under this initial condition is as follows\n1/1T ( 4>iT) = 1 + 2u - 1\n+ 4>im sin (UT + <1\u00bb (13)\nIf the subharmonic oscillation occurs in this system, the input phase angle <1>, expressed by equation 11, must have a real value. Thus the following inequality must be satisfied\n1/1T n1T 4> :> - - tanhIm ~ 211 2u (14)\nHowever, this is not a sufficient condition for subharmonic oscillations to occur, since we may also derive the subharmonic oscillation having an irrational wave-form, as shown in Figllre 3, from the above condition only. The restraint\nexcluding such an irrational subharmonic wave is given by\n4>e( T') ;;;, 0\nthen from equation 13,\n1 + n1T 211\n~e\"\" sin (liT' + <1\u00bb :> 0 ~\n( n1T)' I + tanh?, e- 7 - 1 _11 + 114>/,,, cos (liT' + <1\u00bb = 0 I J , ,\nwhere\n<1> = -sin-1 [(;: - tanh ;:) /4>imJ\n(15)\nThe relation between 4>i'\" and 11 obtained by eliminating T' from equation 15 gives the condition for excluding the irrational subharmonic wave. Then, the necessary and sufficient condi tions of occurrence of the subharmonic oscillation are given by equations 14 and 15.\nFigure 4 shows the domains where equations 14 and 15 are satisfied simultaneously, the coordinates of which are the\nI3\n103 'r--r\n~ 1 ~ ~~~ p\n\" ~~ ~x~ fJ~';--' ~5 ~ i\"~~'l~ ' ... r------t-3 0 t\"\"1 ~C\"\" ~ ~ t--... l~~~ ....... ~ \"~ ~:1 ....... f~'t-G .......\n~ 10\n....... ~Q\nI ----i\"'\" I 1\n1 It--- ....... 10-4 10-3 10-2 10-1\nQ>im (=6im /KTA )\nF(f{ure 4. Domains where subharmonic oscillations occllr shown by full, dotted and chain lines; theoretical shown by dots, crosses, etc. analogue complIterca/cu/ations\nfrequency 11 and the magnitude 4>iln of the input sine wave. The domain of the possible fundamental oscillation is given by the area above the solid line labelled n = I, and the domains of the subharmonic oscillations are presented by the inner parts of the boundaries, each of which is labelled in the order n of the corresponding subharmonics, and these domains shift to upper left according to the order n. In Figllre 4, the domains of higher order than the If7th subharmonic are omitted.\nThe lower boundaries of each domain of the subharmonic oscillation other than the fundamental are obtained from equation 14, and it is verified from equation 11 that the phase angle <1> of the system input is -1T/2 for each subharmonic oscillation on these boundaries. On the other hand, the upper boundaries are derived from equation 15. Only the fundamental oscillation is free from the condition of equation 15, and there is no upper boundary.\nThe facts derived from Figure 4, are as follows. (I) Let 4>im be kept constant; the higher the input frequency, the greater will the order n of the possible subharmonic oscillation become.\n399" ] }, { "image_filename": "designv11_101_0000382_s0278641913020088-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000382_s0278641913020088-Figure2-1.png", "caption": "Fig. 2. Properties 1, 2, and 3 as the bases of property 4 and formula (1).", "texts": [ " The properties apply to the numbers from the half interval [0.5, 1), presented without proofs for the sake of simplicity. Let us first introduce special denotations for three numerical sequences: Sequence An generates a family of enumerated nonintersecting half intervals I2, I3, \u2026 in the form In = . As a group, these half intervals comprise the decomposition (Fig. 1) of the set [0.5, 1); i.e., [0.5, 1) = . For the half interval [An, 1), we use the denotation In+ = . Property 1. An \u2013 1 < Bn < Cn < An for all n \u2265 2 (Fig. 2a). Any number x from [0.5, 1) belongs to one half interval In, which means that there exists the single valued function z: [0.5, 1) {2, 3, \u2026}, assuming the value n if x \u2208 In. For example, z(0.625) = 2, z(0.828125) = 3. Let x \u2208 In = [ , Bz) \u222a [Bz, Az). We consider the number x/Az. Property 2. If , then (Fig. 2b). Property 3. If , then (Fig. 2b). As follows from properties 2 and 3, transforming the half interval Iz by dividing its numbers into Az leads to a dual result: \u2014Numbers from [Bz, Az) are displaced into Iz+. \u2014Numbers from [Az \u2013 1, Bz) remain in Iz, but property 1 ensures that the repeated division into Az will still displace them into Iz+. An 2 n 1\u2013 2 n , Bn 2 n 1\u2013 2 n \u239d \u23a0 \u239b \u239e 2 An 2 , Cn 2 n 2\u2013 2 n 1\u2013 , where n 1 2 \u2026., ,= = = = = An 1\u2013 An ),[ I2 I3 \u2026\u222a \u222a In 1+ In 2+ \u2026\u222a \u222a Az 1\u2013 x Az 1\u2013 Bz ),[\u2208 x/Az Cz Az ) Iz\u2282,[\u2208 x Bz Az ),[\u2208 x/Az Az 1 ),[\u2208 Iz+= Keywords: Logarithm, method, reduction, error, algorithm" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000444_amr.837.316-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000444_amr.837.316-Figure2-1.png", "caption": "Fig. 2. a) The mesh for the cam with coating and cam follower assembly b) The mesh for the cam with coating and cam follower with coating assembly c) The mesh for the cam and cam follower assembly without any coating", "texts": [ " The rotational velocity of the camshaft is 375 rot/min. The Static Structural module calculated the equivalent stress distribution using the von Mises theory, the normal tension distribution, total deformation and the shear tension distribution. In this module the mesh of the model was defined and also the load and boundary conditions. The mesh consists of tetrahedral elements with refinement in the contact zone. The mesh of the model with the coating only on the cam is made of 71656 nodes and 12279 elements (Fig. 2 - a.). The mesh of the model with coating on the cam and on the cam follower is made of 81693 nodes and 13689 elements (Fig. 2 - b.). The mesh for the model without coating is made of 6182 nodes and 3204 elements (Fig. 2 - c). The cam is supported by a cylindrical clamping with the tangential direction free [1, 2]. The cam follower is also supported by a cylindrical clamping with an axial directional free. The follower applies the force to the cams sliding surface. The cam has a rotational velocity around its axis of 375 turn / min. The material properties of the components were defined in the Data Engineering module from ANSYS 13. a) b) Finite element analyses results The stress distribution for the cam with coating and cam follower assembly In Fig" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001277_amm.391.72-Figure4-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001277_amm.391.72-Figure4-1.png", "caption": "Fig. 4 S1-coordinate system", "texts": [ " S1-coordinate is fixed in the follower surface and S2-coordinate is fixed in the cam surface [2].By applying the homogeneous coordinate transformation matrix [3], the relationship among S, S1 and S2 can be expressed as S1\u2014S: 1 1 1 1 0 0 0 cos sin 0 sin cos xx y y z z (3) S\u2014S2: 2 2 2 cos sin 0 sin cos 0 0 0 1 x x y y zz (4) The rotary velocities of the coordinate systems are as follows: 1 2 i k (5) In the S1-coordinate system as show in Fig. 4, the position vector of the contact point is p while its projection in the coordinate z1-axis is p\u2019, and op\u2019=r. 1 1 1 1R (r ) x y z rtan cos , rtan sin , r , ( , , ) =( ) (6) From Eq. 3, the contact point p in S-coordinate system can be expressed as 1 0 0 rtan cos ( , ) 0 cos sin rtan sin 0 sin cos r x R r y z (7) According to theoretical mechanics [4], the velocity vector of p in S1-coordinate system can be expressed as: 1 1 ( )v R yk z j (8) Similarly, the velocity vector of p in S2-coordinate system can be expressed as: 2 2 ( )v R x j yi (9) The relative velocity of p is expressed as: 12 1 2 ( )v v v yi z x j yk (10) In S1-coordinate system, based on differential geometry [5], the normal vector at the contact point p in the moving coordinate system S1 is expressed as: 1 1 1 1 1 cos cos / / sin cos / / sin R r R n R r R (11) In the fixed coordinate system S, the normal vector at the contact point p is expressed as: 1 1 0 0 cos cos 0 cos sin sin cos cos sin sin 0 sin cos sin cos sin cos sin n n (12) In differential geometry, the meshing condition can be expressed as 12 0v n (13) From Eq" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002063_ciima.2013.6682792-Figure7-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002063_ciima.2013.6682792-Figure7-1.png", "caption": "Fig. 7. Marco de referencia para definir la orientacio\u0301n de los objetos", "texts": [ " Es decir, la orientacio\u0301n de la tuerca debe igualarse en la llave para permitir que la herramienta pueda realizar una accio\u0301n sobre el elemento. Los a\u0301ngulos se calculan utilizando la Ecuacio\u0301n 3, donde uno de los vectores atraviesa los dos puntos de intere\u0301s tanto en la llave como en la tuerca. El otro vector se define de acuerdo al marco de referencia: en el caso de la llave, se utiliza el punto de intere\u0301s ma\u0301s a la izquierda y a partir de este se genera un vector vertical en direccio\u0301n norte; en el caso de la tuerca, se toma el punto de intere\u0301s ma\u0301s bajo y a partir de este se genera un vector horizontal en direccio\u0301n oriente (Fig. 7). Para el ca\u0301lculo de la orientacio\u0301n de la tuerca existe una consideracio\u0301n especial. Dado que los puntos de intere\u0301s se 5 definen como las aristas de la tuerca que se encuentran en una posicio\u0301n ma\u0301s baja, existe la posibilidad de que el a\u0301ngulo que se calcula sea \u03b1, como lo describe la Fig. 7c. Para calcular el a\u0301ngulo correcto \u03b8 (Fig. 7b), se realiza la correccio\u0301n que se muestra en la Ecuacio\u0301n 4 y que se obtiene de la medida del a\u0301ngulo interior de la tuerca. \u03b8 = 50o \u2212 \u03b1 (4) Los experimentos fueron realizados sobre ima\u0301genes obtenidas de la ca\u0301mara del robot NAO, con las librer\u0131\u0301as provistas por la compan\u0303\u0131\u0301a desarrolladora. Tanto la tuerca como la llave deben estar en el a\u0301rea de visio\u0301n de la ca\u0301mara del NAO. La llave se aseguro\u0301 a la mano del auto\u0301mata para evitar rotaciones indeseadas de la herramienta (Fig. 8). Algunos ejemplos de los ca\u0301lculos de la orientacio\u0301n de la llave y la tuerca se presentan en la Fig" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000474_1.4704227-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000474_1.4704227-Figure2-1.png", "caption": "FIGURE 2. Geometry of partially textured journal bearing", "texts": [ " (17), (18), (19) results in as ' : 8 ;V;>4;>' < ':R 8 W;V;>4;>' 4;>' < :R 8 W;V;>4;>' < XV: (21) The net load support of the bearing is obtained by integration of nondimensional pressure as A 8 S , T,$ ,$ 8 E S,< T ,(+F,$ 5 ,$ 8 E , ,( F,$ ,$ R R 8 S , T,$ ,$ (+ (22) Integrating the nondimensional shear stress in Eq. (4) over the bearing surface yields !H\"#$% 8 S , T,$ ,$ R 8 S , T,$ ,$ YV Z8 E ,( F,$ ,$ 8 E ,(+F,$ 5 ,$ [ R Z8 E ,(+F,$ ,$ 8 E ,( F,$ ,$ [ (+ (23) PARTIALLY TEXTURED JOURNAL BEARING ANALYSIS The schematic of partially textured journal bearing representing the inception of partial texture with land and groove are shown in Fig. 2. The partially textured surface is composed of a number of successive regions of land and groove. The angular extent of successive regions of land and groove are \\ ( \\ ( \\ ( \\ ( \\ and \\ (+ \\ ( \\ (+ \\ ( \\ respectively. The textured length is \\ . Convergent Journal Bearing The nondimensional film thickness for the plain journal bearing is expressed in Eq. (24) and the nondimensional film thickness in the textured journal bearing is expressed as . S ]^_`\\T (24) The non-dimensional pressure gradient for the classical plain journal bearing is a (25) The nondimensional shear stress is expressed as " ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001402_s13369-012-0232-3-Figure6-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001402_s13369-012-0232-3-Figure6-1.png", "caption": "Fig. 6 Finite element model of two meshing gears", "texts": [ " 3 were developed by a CAD computer program and then transferred to the FE package Working Model 4D for stress analysis. The pinion comprised 20 teeth and the gear comprised 30 teeth. The pinion with concave teeth was fixed at its axis hole. The gear with convex teeth was allowed to rotate at its center, with a counterclockwise torque of 2.5 Nm applied to its center. The stress analysis was performed with a pressure angle of 20\u25e6. The contour of the mesh of the proposed gear mechanism is shown in Fig. 6. The total number of elements of the pinion is 34,482 with 51,302 nodes. The total number of elements of the gear is 28,014 with 42,890 nodes. Figure 7a shows the three-dimensional stress distributions of the gear and the pinion plotted by von-Mises stress contours. Figure 7b displays the distribution of the contact stress between the gear and the pinion. The maximum stress of von-Mises was 19.7 MPa. Although the mathematical model of the gear and pinion was obtained in Sect. 3, the rack cutter for manufacturing the pinion or the gear was unknown" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002063_ciima.2013.6682792-Figure6-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002063_ciima.2013.6682792-Figure6-1.png", "caption": "Fig. 6. Restricciones implementadas para la seleccio\u0301n de los puntos de intere\u0301s", "texts": [ " Con la informacio\u0301n del centro y el radio del c\u0131\u0301rculo, se limitan el nu\u0301mero de a\u0301ngulos encontrados en el punto anterior, restringiendo a\u0301reas que no son de intere\u0301s. Debido a que el robot NAO sostiene la llave con la mano derecha, se generan restricciones dadas por la movilidad del robot. Espec\u0131\u0301ficamente, la llave nunca estara\u0301 apuntando hacia la derecha o hacia abajo. La seleccio\u0301n de a\u0301ngulos se limita a puntos del contorno ubicados en el a\u0301rea 20% a la izquierda y 20% en el borde superior (Fig. 6a y 6b). En el caso de la tuerca, so\u0301lo puntos ubicados en el 50% inferior del contorno son seleccionados. De las mu\u0301ltiples aristas encontradas en este sector, se seleccionan las dos ubicadas ma\u0301s abajo en la imagen, como se observa en la Fig. 6c. Para determinar la orientacio\u0301n, primero se define un marco de referencia sobre cada objeto. Los a\u0301ngulos quedan definidos de tal forma que tanto la llave como la tuerca se alineen cuando el a\u0301ngulo es el mismo en ambos objetos. Es decir, la orientacio\u0301n de la tuerca debe igualarse en la llave para permitir que la herramienta pueda realizar una accio\u0301n sobre el elemento. Los a\u0301ngulos se calculan utilizando la Ecuacio\u0301n 3, donde uno de los vectores atraviesa los dos puntos de intere\u0301s tanto en la llave como en la tuerca" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003339_b978-0-08-098332-5.00002-4-Figure2.10-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003339_b978-0-08-098332-5.00002-4-Figure2.10-1.png", "caption": "Figure 2.10 Output voltage waveforms of single-phase fully controlled rectifier supplying an inductive (motor) load, for various firing angles.", "texts": [ " Fortunately, however, it turns out that the output voltage waveform for a given a does become independent of the load inductance once there is sufficient inductance to prevent the load current from ever falling to zero. This condition is known as \u2018continuous current\u2019, and, happily, many motor circuits do have sufficient self-inductance to ensure that we achieve continuous current. Under continuous current conditions, the output voltage waveform only depends on the firing angle, and not on the actual inductance present. This makes things much more straightforward, and typical output voltage waveforms for this continuous current condition are shown in Figure 2.10. The waveforms in Figure 2.10 show that, as with the resistive load, the larger the delay angle the lower the mean output voltage. However, with the resistive load the output voltage was never negative, whereas we see that, although the mean voltage is positive for values of a below 90 , there are brief periods when the output voltage becomes negative. This is because the inductance smoothes out the current (see Figure 4.2, for example) so that at no time does it fall to zero. As a result, one or other pair of thyristors is always conducting, so at every instant the load is connected directly to the supply, and therefore the load voltage always consists of chunks of the supply voltage", " In a single-phase diode bridge, for example, the commutation occurs at the point where the supply voltage passes through zero: at this instant the anode voltage on one pair goes from positive to negative, while on the other pair the anode voltage goes from negative to positive. The situation in controlled thyristor bridges is very similar, except that before a new device can take over conduction, it must not only have a higher anode potential, but it must also receive a firing pulse. This allows the changeover to be delayed beyond the point of natural (diode) commutation by the angle a, as shown in Figure 2.10. Note that the maximum mean voltage (Vdo) is again obtained when a is zero, and is the same as for the resistive load (equation (2.3)). It is easy to show that the mean d.c. voltage is now related to a by Vdc \u00bc Vdocosa (2.5) This equation indicates that we can control the mean output voltage by controlling a, though equation (2.5) shows that the variation of mean voltage with a is different from that for a resistive load (equation (2.4)), not least because when a is greater than 90 the mean output voltage is negative" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001430_s12206-012-1266-x-Figure10-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001430_s12206-012-1266-x-Figure10-1.png", "caption": "Fig. 10. Failed lock nut in experimental simulation.", "texts": [ " The rear end of the wheel shaft was connected to an electric motor simulating the driving carrier. The flange housing was clamped to a vertical thick plate clamped on the table. The motor speed was set to 150 rpm and the load was set to be 34 kN. Since the vertical displacement at the end of the wheel shaft began to increase, indicating that the lock nut was being loosened, the experimental simulation was stopped. The wheel shaft assembly was disassembled and the contact surface of the lock nut was observed. The lowered annular region was developed but not completely. As shown in Fig. 10, the worn out regions are currently developed as twenty separate regions. Circular scratches of radius 0.3 ~ 0.4 mm were seen in Fig. 11. The radius is smaller than that of the lock nut in the field operation. As the contact surface is increasingly worn out, the radius is expected to reach 0.5 mm. The contact surface of the lock nut first appears to have collapsed by excessive contact pressure. However, structural analysis results show that the stress was not large enough to initiate yielding, but the slippage on the contact surface was considerable" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002096_aimsec.2011.6011034-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002096_aimsec.2011.6011034-Figure1-1.png", "caption": "Fig. 1. The three areas with a indiviidual in the center", "texts": [ " In the Year 2002, Couzin et.al simulated this model with the help of computer5.Consider a group with N Individuals(i=1...N),every individual has a position denoted by ci and a direction vector vi.For each step of time,the max angle a individual rotate is \u03b8.Individuals can only see their neighbours within distance denoted by a .The max angle of view is \u03b1.There are three areas,say the zone of repulsion(zor),the zone of orientation(zoo) and the zone of attraction(zoa).Obviously a = zor + zoo+ zoa(see Fig.1). For each step,a individual will scan it\u2019s zor first,if there are any other companion,it\u2019s next direction is as below(Eq.1). di(t+ 1) = \u2212\u2211nr j =i rij(t) |rij(t)| (1) Where rij = cj \u2212 ci is the direction vector the individual i towards j. Eq.1 shows that,if individual space is too crowded,one will move away. If there is nobody in zor,the individual will scan the zoo and zoa,and if there is anyone in it,the next direction is provided by Eq.2 di(t+ 1) = 1 2 ( \u2211no j=1 vj(t) |vj(t)| + \u2211na j =i rij(t) |rij(t)| ) (2) Eq" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001164_2013-01-1911-Figure8-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001164_2013-01-1911-Figure8-1.png", "caption": "Figure 8. Tire vibration model", "texts": [ " For comparison of the radial vibrations between output positions, Figures (a) and (b), show phase reversal of both amplitudes in the circumference direction. Therefore, the lateral bending mode is estimated for the outof-plane vibration of the tread area. In addition, it is found that comparison of circumference and radial vibration supposes that the tread deformation follows the inextensional deformation assumption. Table 1. Natural frequencies of translational and lateral bending mode Figure 5. Frequency response function and mode shape at peak frequencies Tire Vibration Model Based on Cylindrical Shell Theory Figure 8 shows the vibration model based on a thin cylindrical shell theory. Moreover, this model consists of a thin circular ring that represents the tread and circumferentially distributed radial, tangential and lateral springs that represent the sidewall, which connects the circular ring to the wheel center. The wheel is a rigid body because of its high stiffness in comparison with tire stiffness. Kirchhoff-Love's hypothesis was supposed in this model [3]. The model parameters are: radius R, thickness of tread b, mass density \u03c1, are radial, tangential and lateral sidewall stiffness K\u03b8, Kr and Ku" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001115_s12555-012-0087-0-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001115_s12555-012-0087-0-Figure2-1.png", "caption": "Fig. 2. Membership functions of fuzzy controller.", "texts": [ "6366 1]=C , d di [0 0.1]= =C C , i 1= =D D , i i 0.1=A A , di di 0.2 ,=A A i i 0.3 .=B B The membership functions of x1(k) of the T-S fuzzy model are shown in Fig. 1. Based on the IPM technique, the Theorem 1 and Theorem 2 are employed to design state feedback fuzzy controller and output feedback fuzzy controller, respectively. In this example, it is assumed that the premise variable x1(k) of the fuzzy controller is constrained within [ 2, 2]\u03c0 \u03c0\u2212 and the corresponding membership functions are given in Fig. 2. Imperfect Premise Matching based Fuzzy Control with Passive Constraints for Discrete Time-Delay Multiplicative\u2026 621 4.1. State feedback case Before starting the calculation process, let us set the parameters as 1 ,S I 2 0.8,S 3 0.8S and \u03c4 9= for designing state feedback fuzzy controller. Using the LMI Toolbox of MATLAB, the following feasible solutions and state feedback gains can be obtained via Theorem 1 and Algorithm 1. [ ]1 0.7838 0.2006 ,= \u2212F [ ]2 0.5659 0.2123 ,= \u2212F [ ]3 0.4048 0.2161= \u2212F " ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002030_ijhvs.2013.053008-Figure21-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002030_ijhvs.2013.053008-Figure21-1.png", "caption": "Figure 21 MB bogie and driveline models", "texts": [ " Once the driveline torsional characteristics have been tuned by means of the previously described lumped parameters model, a complete vehicle model, including the driveline, has been implemented using VI-Rail (see Figures 20 and 21). The model has more than 80 dofs and it is able to reproduce both the running and the comfort behaviour of the vehicle. Ad hoc templates for bogies and wheelsets have been implemented to take into account the torsional deformability of the driveline. In particular each shaft/axle has been divided into three rigid bodies connected by bushings (Figure 21). The stiffness and damping characteristics of the bushings have been tuned according to the previously performed lumped parameters analysis. This approach has been preferred with respect of introducing deformable bodies into the model in order to speed up the simulations while maintaining a good accuracy. Within VI-Rail, wheel-rail contact geometry is introduced by means of pre-calculated nonlinear tabular elements (Kik et al., 2000), while the normal force is calculated as a function of the compenetration between the wheel and the rail (Piotrowski and Kik, 2008)" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000524_amr.301-303.573-Figure4-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000524_amr.301-303.573-Figure4-1.png", "caption": "Fig 4 The geometrical structure of 4rd flow field", "texts": [ "(10) and taking into initialization of the entrance into consideration, the flow owing to pressure drop can be calculated as follows [5]: ( ) ( ) 2 2 4 43 3 3 3 3 3 3 3 3 3 3 355 13 3 3 3 3 3 3 8 1 2 1 2 1 24 2 22 1n q p p n n q s h s h s th h h th s c c ul c l s hn \u03c0 \u03c0 \u03c0 \u00b5 \u221e = \u2032 \u2206 \u2206 \u2212 \u2212 = = + \u2212 + \u2212 \u2211 (11) Where 3 3 3 3 (1 ) 1 64 H Re d c k l = + + \uff0cand 1 4 3 3 3 2.62 H l k Re d = \uff0c ( ) 3 3 3 3 2 q Re s h \u03c1 \u00b5 = + According to Eq.(11) , hydraulic resistance R3 can be calculated from the equation ( ) ( ) 3 3 3 3 3 2 2 4 4 3 3 3 3 3 3 3 355 1 3 3 1 8 1 2 1 2 1 24 2 22 1n p c l R q n n s h s h s th h h th s s hn \u00b5 \u03c0 \u03c0 \u03c0 \u221e = \u2206 = = \u2212 \u2212 + \u2212 + \u2212 \u2211 (12) Section 4rd is orifice outflow. Fig.4 shows the geometrical model. a\u2014a end is the entrance end with pressure p4 ; c\u2014c end is the exit end with pressure p4o. According to Bernoulli equation and continuity equation, the leakage can be calculated as follows: 4 4 3 2 3 4 ( ) 2 / d q C r r h p \u03c1= \u2212 \u2206 (13) Cd4\u2014\u2014flux coefficient of 4rd flow field Reynolds number of flow field 4rd is estimated using [ ] 4 4 3 2 3 2 ( ) q Re r r h \u03c1 \u00b5 = \u2212 + (14) The flux coefficient of 4rd flow field relates to its structure and Reynolds number of moving fluid", " Together with the formulas of hydraulic resistances, non-linear equations containing the three unknown parameters Cd4, q, c3 can be drown as a mathematical model of static sealing flow field. ( ) ( ) ( ) ( ) ( ) ( ) 2 0 1 2 3 02 2 2 4 3 3 2 1 4 3 3 3 0.5 1.5 2 2.54 6 8 4 4 4 4 4 4 0 2 ( ) 1 1 1 2.62 0 64 0.01128+0.13034 0.00803 +2.56124 10 4.06155 10 +2.52316 10 d d q q R R R R p C h r r qw c qw w q w q w q w q w q C \u03c1 \u2212 \u2212 \u2212 + + + + \u2212 = \u2212 \u2212 \u2212 + = \u2212 \u2212 \u00d7 \u2212 \u00d7 \u00d7 = (16) Where ( ) ( ) [ ] 3 3 3 42 3 2 32 1 3 3 4 2 , ( ) s h w w r r hr r s h \u03c1 \u03c1 \u00b5\u00b5 = = \u2212 +\u2212 + The principle of the test bench is shown in Fig.4, composed by mechanical system, hydraulic system and test system. Test device includes: 1.tank 2.temperature sensor 3.loading motor 4.gear pump 5.filter 6.pressure sensor (spillover valve) 7. frequent conversion motor 8.speed and torque sensor 9.input shaft 10.gear transmission 11.pressure sensor (entrance of sealing system) 12.entrance where pressed oil enters test box 13.inlet bushing 14.distributing bushing 15.the sealing test segment 16.rotary sealing ring 17.main shaft 18.exit where leaking oil outflows 19" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001600_sisy.2011.6034297-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001600_sisy.2011.6034297-Figure3-1.png", "caption": "Figure 3. Position of tip after involved simplification.", "texts": [ " Detailed explanation of all components of (1)-(6) can be finding in [3], [4]. A. Example Let us analyze the behavior of the robotic pair consisting of elastic gear and flexible link with one mode as depicted in Fig. 2, 3 and 4. \u03be\u03b8\u03b3\u03d1\u03be\u03b8 +=++= ,rq . (7) The dynamic model (both the model of flexible line and model of the motion of mode tip) is defined according to classical principles but with the previously introduced new DH parameters, using Lagrange\u2019s equations. The following quantities are adopted: q , \u03b3 , q\u03d1 and \u03b8 as generalized coordinates (see Fig. 3 and 4). We accept that the top of the mode is moving continuously on the surface of ball, which radius is l without shortening for mode. \u03bd \u03d1qq lf sin\u22c5= . (8) \u03bd\u03bd \u03d1sin\u22c5= lf . (9) If we consider small bending angles and adopt that l fq q =\u03d1sin , q r r f f =\u03d1sin , l fr v =\u03d1sin , and vv \u03d1\u03d1 sin\u2248 , qq \u03d1\u03d1 sin\u2248 , rr \u03d1\u03d1 sin\u2248 , we obtain (see Fig. 3 and 4): rqv \u03d1\u03d1\u03d1 \u22c5= . (10) \u2013 86 \u2013 \u03bd\u03bd \u03d1\u03d1 sin,sin \u22c5=\u22c5= lflf qq . (11) Potential, dissipative energy as a result of elasticity link is: 22 2 1 2 1 rsqspels fCfCE \u22c5\u22c5+\u22c5\u22c5= , 22 2 1 2 1 rsqsels fBfB \u22c5\u22c5+\u22c5\u22c5=\u03a6 on the top of link. To bring previous expressions in the form dependent of generalized coordinates, it should be expressed via generalized coordinates using (7)-(11). Follows that deflections are: )( \u03b3\u03d1 \u2212\u22c5\u22c5= qlf qr . (12) )()( \u03b3\u03d1\u03b3\u03d1 \u2212\u22c5\u22c5+\u2212\u22c5\u22c5= qlqlf qqr . (13) By applying expression (7)-(13) in previous equations, potential energy of elastic link is needed: 22222 )( 2 1 2 1 \u03b3\u03d1\u03d1 \u2212\u22c5\u22c5\u22c5+\u22c5\u22c5\u22c5= \u22c5 qlClCE qsqspels " ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002063_ciima.2013.6682792-Figure5-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002063_ciima.2013.6682792-Figure5-1.png", "caption": "Fig. 5. Ejemplo de los datos retornados por la funcio\u0301n de c\u0131\u0301rculo de a\u0301rea m\u0131\u0301nima que encierra el objeto", "texts": [ " \u03b8 = arc cos ( ~v1 \u00b7 ~v2 \u2016~v1\u2016\u2016~v2\u2016 ) (3) En el caso de la llave, los a\u0301ngulos deben permitir hallar los puntos correspondientes a las salientes de la herramienta (Fig. 3a), mientras que en la tuerca se desea encontrar dos aristas de la parte inferior (Fig. 3b). La librer\u0131\u0301a utilizada para el procesamiento de ima\u0301genes se denomina OpenCV [18]. Entre las funciones de descripcio\u0301n de forma y ana\u0301lisis estructural se encuentra la funcio\u0301n minEnclosingCircle, la cual encuentra el c\u0131\u0301rculo de a\u0301rea m\u0131\u0301nima que encierra un contorno (Fig. 5). Con la informacio\u0301n del centro y el radio del c\u0131\u0301rculo, se limitan el nu\u0301mero de a\u0301ngulos encontrados en el punto anterior, restringiendo a\u0301reas que no son de intere\u0301s. Debido a que el robot NAO sostiene la llave con la mano derecha, se generan restricciones dadas por la movilidad del robot. Espec\u0131\u0301ficamente, la llave nunca estara\u0301 apuntando hacia la derecha o hacia abajo. La seleccio\u0301n de a\u0301ngulos se limita a puntos del contorno ubicados en el a\u0301rea 20% a la izquierda y 20% en el borde superior (Fig" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001101_1369-4332.16.11.1871-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001101_1369-4332.16.11.1871-Figure1-1.png", "caption": "Figure 1. Spinning process for preparing samples of stepped frusta", "texts": [ " Common modes of deformation for circular tubes include axial crushing, lateral indentation, lateral flattening, inversion and splitting. Numerical studies were done on axial crush on corrugated surface tubes by Chen and Ozaki The experiment mainly requires test samples, compression machine to compress the samples and fixture to hold the samples in position while testing. The samples were made by spinning process with commercially available aluminum sheet. A typical mandrel and setup of spinning process is shown in Figure 1. One end of the mandrel is attached to rotating shaft of spinning machine and aluminum round sheet to be spin is fixed at other end with a pressing lug which also rotates. The aluminum sheet is then pressed to take shape of mandrel as shown in Figure 1. The samples were divided in four batches as they were made from 4 types of mandrel. The 1st batch samples were made from the mandrel having 10 degree semi apical angle (side inclination with vertical) for bottom and mid cone. They were designated as A10_B1_S1\u2026. etc. This designation represent the sample have bottom and mid semi apical angle ~10 degree and it is from batch number 1 i.e. B1. \u201cS\u201d stands for sample followed by its number. The second batch was made from the radial scaled down mandrel of 1st batch" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003689_s1474-6670(17)70081-x-Figure7-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003689_s1474-6670(17)70081-x-Figure7-1.png", "caption": "Figure 7. Waveforms ofBi , B\" Bo' Be' and B,", "texts": [ " Approximate Analysis by the Describing Function6,7 In Figure I, () e(t), which is the input signal of the relay element, is considered to be a sum of the system input ()i(t) 400 ON THE SUBHARMONIC OSCILLATION OF THE RELAY SERVOMECHANISM and the subharmonic component O/(t); thus Beet) = 0im sin \u00ab(I)( + \u2265 = (7) Assuming that \u03b1 denotes the angle between x0-axis and the line that connect the point O0 and any point on the parabolic curve, Eq. (7) can be rewritten as ( ) 0 1 2 0 1 2 cot 0 22 cot x p y p \u03b1 \u03c0 \u03b1 \u03b1 = < < = (8) If 0\u03b1 = , the following relationship exist: 0 0 0 0 x y = = (9) Supposing that the parameter 'r is a non-dimensional parameter, which can be described as 1 1 2' 2 p r k r= = , the following relationship exist: 1 1 22p k r= (10) where 1k is the ratio of 'r to 2r , 2r is the radius of the pitch circle of the driven gear. According to the geometry of the triangle OMF in Fig. 2, we have sin sin 2 OM MF \u03c0\u03b8 \u03b1 = \u2212 (11) where OM represents the distance between any point M on the parabolic curve and the center point O, MF denotes the distance between any point M on the parabolic curve and the focus of the parabola F. According to the characteristic of the parabola, OM and MF can be expressed as follows, respectively. 2 2 0 0OM x y= + 1 0 2 p MF y= + (12) Where ( 0 x , 0 y ) denotes the coordinate values of point M in coordinate system \u03a3 (X, O, Y). Substituting Eqs. (8) and (12) into Eq" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000886_amr.317-319.764-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000886_amr.317-319.764-Figure1-1.png", "caption": "Fig. 1 Deformation model of the work piece", "texts": [ " All rights reserved. No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of Trans Tech Publications, www.ttp.net. (ID: 137.222.24.34, University of Bristol, Bristol, United Kingdom-14/04/15,02:46:32) The constant load of the lifting system mainly includes the mass of the clamp, work piece and the clamp support with motors and gearboxes. The time-varying load is the resistance force from the deformation of the work piece. Fig. 1 shows the schematic deformation model of the work piece [4]. In this model, the work piece is simplified to be a cantilever beam between point O and point P. Point O is held in the clamp and point P is the center of the vertical cross section of the pressed area [5,6]. The relationship between the deformation of the work piece and the vertical resistance force on the clamp can be expressed as y 3 f F 3EI l \u2206y = . (1) Where \u2206y is the relative displacement of point P to O in vertical direction, Fy is the vertical resistance force on the clamp, lf is the distance between point O and P in horizontal direction, E is the elastic modulus of the work piece, and I is the moment of inertia of the cross-sectional area with respect to the neutral axis" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000366_j.proeng.2013.09.235-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000366_j.proeng.2013.09.235-Figure1-1.png", "caption": "Fig. 1. Quadruped machines with different types of legs. (a) type RRPRRRR; (b) type PRRRRR; (c) type RRRRRR", "texts": [ " It was obtained the first-order derivative pseudo-inverse Jacobian matrix using two different numerical methods for multi-joint legs: the right pseudo-inverse, and by singularity properties using the singular value decomposition approach. \u00a9 2013 The Authors. Published by Elsevier Ltd. Selection and peer-review under responsibility of the organizing and review committee of IConDM 2013. Keywords: multi-leg; mobile robotics; kinematics control; dynamic-control. 1. Introduction Hyper-static balanced multi-legged walking robots are mechatronic vehicles capable to walk on multi-joint legs (see fig.1). Multi-legged robots with three or more extremities are statically stable when walking, [1]. However, depending on its gait configuration in use, legs must correctly be synchronized while developing free-walking over * Corresponding author. Tel.: +52-656-688-4800 E-mail address: edmartin@uacj.mx \u00a9 2013 The Authors. Published by Elsevier Ltd. Selection and peer-review under responsibility of the organizing and review committee of IConDM 2013 all-terrain. If some legs become disabled, the robot may still be able to walk, since not all legs might be needed to accomplish stability", " Looking into the biology literature [8], [9], [10], one can find an amazingly rich variety of insect's combination of joint legs. Arachnids, crickets, ants and so forth, which are invertebrate animals with eight or more degrees-of-freedom (DOF) in each leg. Much attention has been paid to develop algorithms for gaits control strategies, research on insects biology discloses interesting information that may enrich kinematic and dynamic schemes for gaits control for insect-like artificial walking machines; [8].Figure 1 depicts kinematic combinations of walking robot's legs. This manuscript's main significance and contribution is purposed to provide a generalised velocity-based navigation model of redundantly kinematic control law for walking machines of n-leg and k-joint, other similar approaches did not consider redundant Jacobians [4-11]. The proposed model is stated in terms of robot's global acceleration, and formulated as an averaged leg's Cartesian speeds. The state vector is defined as a function of robot's tangential acceleration, and leg's Cartesian velocities are described by their first order Jacobian, which results in redundant kinematics systems [12]" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001129_amr.230-232.554-Figure3-3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001129_amr.230-232.554-Figure3-3-1.png", "caption": "Figure 3-3, it is divided into 4720 tetrahedral unit grid nodes to 1706. Figure 3-4 shows the mesh diagram of contact area .", "texts": [], "surrounding_texts": [ "In order to simplify analytical process, only a planetary gear is permitted in the transmission system.analyseing the stress distribution in a position. As follows: 3.1 The establishment of the geometric modeling process described above. We reduce the calculation time without affecting precision of the premise.we are trying to simplify the model that do not affect stress distribution characteristics in the contact region. such as chamfer, fillet and keyway and so on . 3.2 Setting material properties of the model. the three parts are of steel, setting the Young's modulus as 199GPa and Poisson's ratio 0.27\u00b5 = 3.3 Setting the solid model, taking the constraints into account .only the X, Y, Z three directions of moving freedom, in order to apply Freedom of Motion to the model, establishing the cylindrical coordinate system, taking the theta direction instead of the DOF rotational degrees of freedom. Worm constraints as shown in Figure 3-2 (b), the worm Z-axis, limiting the freedom of R and Z , after the model of restraint shown in Figure 3-2 (a) below; planetary bound plane diagram in the Figure 3-2 (d), binding surface for the planetary gear center hole surface,the constraint model to planetary gear. Figure 3-2 (c) 3.4 Setting the load of the model .the worm is active body, so it has an input torque T1.In the Pro / MECHANICA, in the real model, it can only impose TLAP (Total Load At Point) torque, before the torque applied, first of all you should create surface point by the force then addingTLAP at this point, 3.5 Creating a contact region which is peculiar step of contact analysis. between the planetary gear and the worm or stator tooth contact part, establishing contact region, running the overall intervention at the same time, testing whether there is intervention between the parts. If intervention is exsist, you should adjust the profile to remove interference 3.6 The finite element meshing models. in an integrated Pro / MECHANICA the mesh is AutoGEM (automatic meshing device), using self-applied meshing methods. Setting the grid unit line angle and surface angle from the range of 50 to 1750. After completing the model mesh and shown in 3.7 definiting the analysis tasks, operating analysis, checking the \"include the contact region\"(include contact regions), selecting the single-channel to adaptation (Single-Pass Adaptive), selecting the \"Localized Mesh Refinement\", improving the contact region of the grid density. Operating analysis, after analysis the operation status from the maximum 7, the stress error is 0.4%. 3.8 The picture shows results. In the \"Display Settings\", select \"part / layer\", respectively, each part shows the stress distribution of the contact region. Figure 3-5 for the worm where the stress contours, Figure 3-6 for the planetary gear stress contours. Fig.3-5 Stress distribution in the worm Fig.3-6 Stress distribution in the planet 4.Result We can find theresult from Figure 2-5: 4.1 The worm stress on each contact region is uneven. Close to the torque input, the contact stress is maximum (1309MPa), as shown in Figure in the left. At the other side of the worm, the minimum contact stress (390MPa), as shown in Figure in the right. the ratio was 3.3. 4.2 Each contact region shape is long strips on the worm. At each side of the contact region, the stress is great, but the middle is less, showing a \"dumbbell-shaped\" distribution. because it is caused by elastic deformation when the planet gear meshing with the worm. Figure 2-6 labeled 1,2 and 3 of the planetary gear meshing with the worm. 4.3 In the planetary gear teeth 1,2 and 3 because closing the worm gear 1 input, the contact stress is greater. 5.Conclusion: This paper is based on Pro / ENGINEER to create a toroidal geometry model of the transmission system, analysing its finite element structure.getting contours of the stress. the results show that: the worm stress on each contact area is uneven; each contact area shape is long strips on the worm. At each side of the contact area, the stress is great, but the middle is less, showing a \"dumbbell-shaped\" distribution.; as the planetary change. the stress distribution show a cyclical varition." ] }, { "image_filename": "designv11_101_0002022_amr.430-432.1597-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002022_amr.430-432.1597-Figure3-1.png", "caption": "Fig. 3 Mechanics model of a sliding bearing Fig. 4 Layout of bearings", "texts": [ " Suppose the two bearings of one shaft have the same initial positions and their displacement are represented by y1, y2. The oil film stiffness and damping coefficient are k1, k2 and c1, c2. The dynamic equation of the system can be expressed as: 1 1 1 1 1 1 1 1 1 1 2 2 2 2 2 2 2 2 2 2 m x k x c x k y c y m x k x c x k y c y + + = + + + = + (1) Where y1, y2 are the displacement response of bearings when vertical excitation loaded on the gearbox base, which could be obtained by FEA simulation. The mechanics model of bearing is shown in Fig. 3, numbered in Fig. 4. kxx, cxx, kyy, cyy are stiffness and damping along with coordinate direction. kxy, kyx, cxy, cyx are the cross-coupled stiffness and damping, where the first subscript stands for the force direction, the second one denotes the direction of displacement and velocity. The parameters of bearing oil film are shown in Table 1. In actual calculation, the total stiffness and damping of one shaft are equaled by the parallel connection of two bearings. Because the contact normal is along the meshing line during the meshing process, suppose the displacement difference of gear shaft along the direction equal to the contact deformation: 1 2( ) [ ( ) ( )]cosz t x t x t\u03b4 \u03b1= \u2212 (2) In view of the Hertz space contact theory, the contact stress response is: 2 2 1 2 1 2 | |2 1 1 ( ) z H R E E \u03b4 \u03c3 \u03bd \u03bd \u03c0 = \u22c5 \u2212 \u2212 + (3) Where R is the synthetic curvature radius at meshing point could be expressed as: 1 2 1 2 = R R R R R+ (4) 1 1= sin ycR r g\u03b1 \u00b1 (5) 2 2= sin ycR r g\u03b1 \u2213 (6) between meshing point and meshing pitch point, could be expressed as: 2 2 2 1 1 1sin ( sin )yc cg r r r r\u03b1 \u03b1= \u00b1 \u2212 +\u2213 (7) Where rc(showed in Fig" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001826_s2079096111040159-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001826_s2079096111040159-Figure1-1.png", "caption": "Fig. 1. Construction and technological scheme of the module (machine) aggregate to cut trees and the tops of their trunks: 1\u2014 the basic and mounted unit; 2\u2014cutting working unit; 3\u2014rotating mechanism; 4\u2014longerons; 5, 6, 8\u2014control hydro cylinders; 7\u2014hydro power station.", "texts": [ " Tree harvesting machines such as LP 19 are large and not factored into the planting width of forest belts; in addition, they cannot work with continuous operation of the aggregate, which is necessary for cut ting the tops of trunks. Domestic and foreign garden cutters can only cut tops of trees with diameters of branches up to 60 mm and are not intended to cut sep arate trees and tops of trunks larger than this. A machine serving as a module of a multifunction forestry aggregate for silvicultural use in protective for est plantations (Fig. 1) was developed in the VNIALMI to perform the two indicated operations. The machine is installed on basic raising and mounted equipment 1 (Fig. 1) attached to the front part of a 1.4 class tractor frame (MT3 80/82), in front of its cabin; it has cutting function in the form of a slit ting milling cutter 1000 mm in diameter powered by a hydraulic drive, four link rotation gear, base frame with longerons, and basic mounted equipment managing the hydraulic cylinders and hydraulic power station. Keywords: new method, rejuvenation of forest plantations, new technical equipment, hydraulic drive. DOI: 10.1134/S2079096111040159 238 ARID ECOSYSTEMS Vol. 1 No. 4 2011 ZHDANOV et al. Two settings of the machine\u2019s cutting functions are shown on the construction and technological scheme of the aggregate: (A) cutting trees and (B) cutting off tops of trunks (Fig. 1) The machine is equipped with two independent riparian hydraulic systems: the first one manages the movement of the cutting device, and the second sup plies the hydraulic drive of the cutting device. The first system includes the elements of a tractor hydraulic system, such as a hydraulic tank, NSh 32 gear pulser, R 75 33P distributor, oil filter, hydraulic locks, and hydraulic cylinders; the second system con tains an extra hydraulic power station that includes two parallel and combined NSh 50 pumps with a gear from the rear power take off, an oil tank and filter, blow off valve, and NPA 64 hydraulic motor. Two hydraulic cylinders 5 symmetrically located on each side of the tractor (Fig. 1) are managed by the raising and mounted device; two hydraulic cylinders are for putting the saw in a horizontal position; man aging hydraulic cylinders 8 are for lateral motion of the cutting device, and two hydraulic locks 7 are to main tain the saw in the necessary position. The P\u201375\u201333P distributor placed on the tractor allows one to control the motion speed of the stem of hydraulic cylinder 8 in the interval between the \u201cneu tral\u201d and \u201cforced drop\u201d positions of the distributor handle (Ksenevich, 1984) and, consequently, to con trol automatically the speed of forward travel of the saw when the pressure on it changes during sawing" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000881_apec.2013.6520300-Figure7-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000881_apec.2013.6520300-Figure7-1.png", "caption": "Figure 7. Experimental test machien : SMIIR", "texts": [ " If the rotor-side inverter makes injq+ -axis peak current and the stator-side inverter senses injq+ -axis peak current, then the estimated rotor position is in the +d-axis. Otherwise, if the stator-side inverter senses injq\u2212 -axis peak current, then the estimated rotor position is in the -d-axis as shown in Fig. 6(b). IV. EXPERIMENTAL RESULT To figure out the feasibility of the proposed sensorless method for the SMIIR, experiments has been carried out. The parameters of the test machine are shown in Table 1. Test machine is modified from an off-the-shelf wound rotor induction machine (WRIM) and the diameter of the rotor-side inverter is about 20cm as shown in Fig.7. Fig. 8 shows experimental results verifying (21). The angle of injected voltage, i\u0302nj\u03b8 , slowly changes from \u2013\u03c0 to \u03c0 to confirm the electrical saliency. As shown in Fig. 8(a), the SMIIR reveals no saliency for a sensorless control. While, as shown in Fig. 8(b) the waveforms of i\u0302nj\u03b8 , \u02c6 inj r d shi , \u02c6 inj r q shi ,and f\u03b5 clearly reveal the salient responses. In this experiment the intended injection angle, inj\u03b8 , is set as 4/\u03c0 . And, as mentioned in III B, \u02c6 \u02c6 / 4r inj\u03b8 \u03b8 \u03c0= \u2212 . Hence, (21) can be confirmed from Fig" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001974_978-1-4419-9792-0_60-Figure7-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001974_978-1-4419-9792-0_60-Figure7-1.png", "caption": "Figure 7. Determination of wing section\u2019s weight and center position", "texts": [ " The remaining wing weight and length are measured after each cutting step. Weight is measured by Fisher Scientific accuSeries\u00ae accu-124 analytical balance which has 0.1 milligram readability and \u00b10.2 mg linearity. The length is measured by a caliper. The weight of each wing section can be determined by the weight difference between two adjacent cutting steps, and the position of wing section along the wingspan direction can be calculated by the remaining wing length between adjacent cutting steps. In Figure 7, W1, W2 present the remaining wing weight and L1, L2 are the remaining wing length after corresponding cutting step, so the weight and center position of wing section in Fig.6 can be calculated by the following equations: 1 2-W W W (1) 1 2( - ) 2 L LL (2) IV. Result The spanwise wing section weight distribution of both wings is shown in Fig.8. The data of the artificial wing is scaled down to the same length dimension (divided by the cubed length ratio between the artificial and the natural wing)" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000121_1.3662792-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000121_1.3662792-Figure1-1.png", "caption": "Fig. 1 Average oil velocity in a journal bearing", "texts": [ " T h i s pos i t ion 0 ' is chosen b y t h e journal as a result of equi l ibr ium b e t w e e n t h e fixed load IF ( represented b y the f o r ce v e c t o r R JF) app l ied to the bearing, and t h e integrated h y d r o d y n a m i c oil film pressure react ion r e p r e - sented b y the equa l and o p p o s i t e f o r ce v e c t o r WR. If w e assume that oil flow in the s u p p o r t i n g oil film reg ion is laminar a n d tha t its v e l o c i t y var ies l inearly b e t w e e n t h e b o u n d - aries, the a v e r a g e angular v e l o c i t y of the oil m i l b e A r , / 2 ( see F i g . 1). Cond i t i ons s h o w n in Fig. 1 are for the s teady - s ta te running of a l oaded bear ing w i t h o u t oil whip . If w e r e m o v e the load IF and then strike t h e sha f t in such a w a y tha t the journal c e n t e r will r e - turn to its f o r m e r pos i t i on 0 ' , oil w h i p m a y b e induced . T h e oil pressure react ion WR, n o longer c o u n t e r b a l a n c e d b y RIF, acts on 0' pushing the journal a round 0 w i th t h e t o r q u e WR X OB. W h e n 0' m o v e s to a n e w pos i t ion , WR will c h a n g e direct ion , cont inu ing to ro tate 0 ' a b o u t O in the d irect ion of sha f t rotat ion N j " ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002338_978-94-007-2184-5_8-Figure8.5-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002338_978-94-007-2184-5_8-Figure8.5-1.png", "caption": "Fig. 8.5 The computational meshes for two three-dimensional fl ow chambers. The membranes are modeled as planes in the middle of the chambers, lying in the symmetry-planes normal to the z- axis. Both chambers have a symmetry-plane in the x\u2013z plane, which is why only half the chambers are presented. The chambers are presented in counter-current cross-fl ow confi guration with the feed and draw solutions being pumped through the chamber compartments in opposite directions. The zoom-in on chamber A demonstrates how the computational domains are fi nely graded towards the membrane so as to accurately describe concentration polarization effects", "texts": [ " The solution variables (e.g. velocity vector and pressure) in each computation point are found by numerical solution of matrix equations. The quality of the results from a CFD model is strongly depending on the quality of the computational grid. Often, it is possible to take advantage of symmetrical conditions and hereby reduce the computational domain. The grid has to be designed in such a manner that the resolution (size of each grid cell) is suffi ciently dense to capture the important physical phenomena, see Fig. 8.5 . On the other hand, the computational power demand increases with the number of grid cells. Thus, to keep down the computation costs, the design of the mesh if often a tradeoff between these two considerations. The boundary conditions represent the infl uence of the surroundings that have been cut off when defi ning the computational domain. The solution inside the computational domain may strongly depend on the boundary conditions. Proper choice of these is therefore very important. Alternatively the boundaries of the computational domain should be far away from the region of interest in order not to \u201ccontaminate\u201d the solution with the inaccurate boundary conditions" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001428_icra.2012.6225046-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001428_icra.2012.6225046-Figure2-1.png", "caption": "Fig. 2. Considered target objects: (a) The selected set of objects whose models were used in the evaluation; (b) & (d) Reconstructed meshes based on unfiltered point clouds acquired by range sensing; (c) & (e) Corresponding ground-truth CAD-model meshes; Note the surface distortions and double layers on the sensor-acquisition based meshes;", "texts": [ " \u2022 It was verified whether GF passed the preselection routine defined in Section III-B.3. \u2022 If both grasps GR and GF fulfilled the given task requirement, CC n , CR n and CF n were computed. The task wrench space (see Section II-B) associated with each grasp was formulated as the largest origin-centered insphere of the GWS associated with GC , multiplied with a factor \u03b1 = 0.8. \u2022 The respective intersections of CR n and CF n with CC n were generated using Algorithm 2. Furthermore, the quality measures Qn for CR n and CF n were computed. On the CAD meshes of every object in Fig. 2-(a) two test sets, containing 200 randomly generated four- and five-finger force closure grasps respectively, were produced in order to ensure an unbiased and statistically significant evaluation. Already during the generation of the test grasp sets we found that the proposed smoothing/preselection method increases the number of grasps which preserve the task requirement after mapping them onto the range-data based meshes. More specifically, 98.3% of the grasps on the filtered mesh were eligible, while only 88% of them fulfilled the task after projection onto the raw mesh" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001228_icit.2012.6210047-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001228_icit.2012.6210047-Figure2-1.png", "caption": "Fig. 2. A simplified schematic of the hand-over-hand movement of kinesin nanomotor through four (a) to (d) steps.", "texts": [ " Conventional kinesin (hereafter kinesin) nanomotor is a protein dimer with two globular heads connected together via a short and flexible neck linker at the long central coiled coil stalk region that ends in a tail domain opposite the heads [2] (Fig. 1). Kinesin nanomotors are able to recognize the direction of movement along MTs and transport cargos towards the plus-end of MTs in a hand-over-hand fashion [6, 7]. Intracellular cargo transport occurs along MTs when the appropriate nanomotor simultaneously binds to a cargo and to MT through its tail and heads respectively. According to the Fig. 2, the hand-over-hand movement process is as follows: 1) the movement begins with both heads of the nanomotor in the ADP form, while one head is bound to MT, 2) the exchanging of bound ADP for ATP in the binding head leads to lock of this head to MT and throw the second head towards the plus-end of MT, 3) ATP-hydrolysis occurs in the first head and primes the movement of the nanomotor, and 4) after releasing of Ph through exchanging of A TP for bound ADP in the second head and a rotation of the back hand into the front the movement of the nanomotor is accomplished" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000194_amr.712-715.2192-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000194_amr.712-715.2192-Figure1-1.png", "caption": "Fig. 1. Configuration of the pilot stage for jet-pipe servovalve driven by GMA", "texts": [ " Giant magnetostrictive actuator (GMA) based on giant magnetostrictive materials (GMM) has large blocked force, high energy density, and large bandwidth capabilities. In addition, GMA has no moving parts and are therefore, mechanically less complex than conventional torque motor applied in servovalve [3,4]. Hence the reliability of jet-pipe servovalve driven by GMA is higher than conventional servovalve. So, the research on the characteristics of GMA is important to develop a higher performance servovalve. As shown in Fig.1, the pilot stage of jet-pipe servovalve consists mainly of GMA and jet-pipe hydraulic amplifier. GMA includes a bias solenoid coil used for providing a bias magnetic field, a driving solenoid coil generating a control magnetic field, a GMM rod, an output rod and a spring using for applying a pre-stress on GMM rod to obtain a larger magnetostrictive strain with same magnetic field. When the sign of magnetic field produced by the driving solenoid coil is the same as the sign of the bias magnetic field, the length of GMM rod elongates" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001430_s12206-012-1266-x-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001430_s12206-012-1266-x-Figure3-1.png", "caption": "Fig. 3. Boundary conditions.", "texts": [ " In industry, the lock nut is tightened until the torque, needed to rotate the wheel shaft under no wheel load, increases to a certain level. The lock nut is tightened to the predetermined torque T. The tightening force Q of the lock nut is calculated from Eq. (1) and Table 2. The friction on the carrier face is considered as well as the friction on the threads. T = Q dm tan(\u03bb+\u03c1) /2+fQRm (1) The wheel load was applied vertically on the wheel shaft. Either single tire or dual tires can be mounted on the wheel. If dual tires are mounted, the center of the load shifts 74 mm left off the wheel shaft as shown in Fig. 3. Compared to the single tire type, the dual tire type induces more bending moments in the wheel shaft since both the magnitude of the wheel load and the distance of the load point are larger. Since the wheel housing was deleted, the center of the load was applied at the vacant point. The virtual part, similar to the rigid link, in CATIA structural analysis was used to connect the center of the load to the wheel shaft. As described in the role of the spacer, the preload on the bearing depends on the gap between the spacer and the inner race" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003692_204124791200300101-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003692_204124791200300101-Figure1-1.png", "caption": "Figure 1. Schematic illustration of the formation of \u03b3-Fe2O3/PANI\u2013levodopa", "texts": [ " \u03b3-Fe2O3/PANI was obtained by centrifugation (Model 80-2, 4000 r min-1), washed successively with double-distilled water and ethanol, and finally vacuum dried at 70\u00b0C for 12 h. In order to obtain the \u03b3-Fe2O3/PANI base, the synthesised \u03b3-Fe2O3/ PANI was immersed in 0.2 mol dm-3 aqueous ammonium during stirring for 24 h, filtrated and washed with double-distilled water to a pH of 6\u20137, and then vacuum dried at 70\u00b0C for 12 h. A doping method has been developed for the immobilisation of levodopa, with the magnetic nanomaterial \u03b3-Fe2O3/PANI as the carrier (Figure 1). The most striking advantage of this procedure is that the structure of the immobilised levodopa is hardly affected. The preparation method of \u03b3-Fe2O3/PANI\u2013levodopa is as follows. A quantity of 0.22 g of \u03b3-Fe2O3/PANI base was suspended in 40 L of ethanol, and 0.034 g of levodopa was dissolved in 50 mL of 0.2 mol dm-3 HCl, and the two solutions were then mixed overnight at room temperature with constant stirring. \u03b3-Fe2O3/PANI\u2013levodopa was obtained by centrifugation of the suspension, washed successively with 0" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001802_emeit.2011.6023251-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001802_emeit.2011.6023251-Figure2-1.png", "caption": "Figure 2. Radial static air bearing working principle Diagram", "texts": [ " On this basis, further analysis of the influence of structural parameters on Loadcarrying capacity is carried out. II. STRUCTURE AND WORKING PRINCIPLE OF RADIAL STATIC AIR BEARING The structure of radial static air bearing is shown in Figure. 1.After the compressed air is filtered and dried, it is guided into the clearance of the bearing, and the lubricating film formed around the spindle, so as to sustain the radial load. In addition, the exhaust passage is machined in the spindle, which is open to the atmosphere, and it can make the bearing exhaust conveniently. Figure 2 is the working principle diagram of radial static air bearing. The working principle of radial static air bearing is shown as follows, the flow resistance of the restrictor unchanged, while the flow resistance of the air film changes with the change of the bearing clearance. When a load is added on the bearing, the axis moves an eccentric magnitude 978-l-61284-088-8/ll/$26.00 \u00a92011 IEEE 948 12-14 August, 2011 along the direction of the load, and the clearance decreases in the near side of the spindle and bearing, so the flow resistance of the air film increases, thus the pressure in this clearance rises" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002030_ijhvs.2013.053008-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002030_ijhvs.2013.053008-Figure3-1.png", "caption": "Figure 3 Scheme of the driveline", "texts": [ " The effect of a decreasing section of the creep force-creep curve on torsional vibrations in the driveline will be investigated in Section 4, while the implementation of the proposed contact model with a temperature dependent friction coefficient into a MB vehicle model including torsional deformability of the driveline will be discussed in Sections 6 and 7. A heavy-duty diesel locomotive (model D146) has been considered in this paper (see Figure 2). The main characteristics of the vehicle are reported in Table 1, while the main elements constituting the vehicle driveline are shown in Figure 3. Two bogies are connected to the locomotive, having two powered axles each (axle arrangement B\u2019B\u2019 according to the UIC classification, Table 1). As it can be seen from Figure 3, the driveline layout is symmetric with respect to the central hydraulic fluid clutch. Front/rear primary cardan shafts are connected at one side with the hydraulic fluid clutch, receiving the traction power from the diesel engine, and at the other side with the front/rear secondary cardan shafts. Finally, secondary cardan shafts and wheelsets are connected by means of gearboxes. In order to study torsional vibrations in the locomotive driveline, experimental tests have been carries out by Trenitalia s" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002278_978-1-4471-4628-5_3-Figure3.167-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002278_978-1-4471-4628-5_3-Figure3.167-1.png", "caption": "Fig. 3.167 Bond graph model of laterally diffused NPN BJT with linking current revealed in the model", "texts": [ " The correct model showing the energetics of the problem must be able to explain the linking current. Thus, instead of modeling the individual current injections, one can model only the linking current. From Fig. 3.166, we find that the flow variables in the two ports of the R field are IDBC \u2212 IDB E and IDB E \u2212 IDBC . The magnitude of these two flow variables being equal indicates the existence of a hidden 1-junction. The opposite signs can be adjusted through power directions. Thus, we can replace the R field by a 1-junction and one-port R element as shown in Fig. 3.167. The effort variable in the bond connected to the one-port R element becomes VB\u2032 E \u2032 \u2212VB\u2032C \u2032 = VC \u2032 E \u2032 . The R element cannot use this effort information to compute the flow, which is the linking current IL = IDB E \u2212 IDBC because the constitutive relation for the linking current given in Eq. 3.192 requires values of VB\u2032 E \u2032 and VB\u2032C \u2032 . Therefore, the one-port R element is modulated by the required effort signals. The bond graph in Fig. 3.167 shows the path of linking current from the emitter to the collector and gives a better physical understanding of the power transfer in the system compared to a flow injection-based model. In [27], the same model given in Fig. 3.167 was obtained differently. From Fig. 3.165, one can write the weak junction laws for the two 0-junctions (flow sum relations) connected to flow injection sources as follows: IC + IDB E \u2212 IC BC \u2212 IDBC = 0 \u21d2 IC + (IDB E \u2212 IDBC ) = IC BC (3.193) and IE + IDBC \u2212 IC B E \u2212 IDB E = 0 \u21d2 IE = (IDB E \u2212 IDBC ) + IC B E (3.194) Moreover, VC \u2032 E \u2032 = VB\u2032 E \u2032 \u2212 VB\u2032C \u2032 (3.195) Using the above three relations, the two-port R field in Fig. 3.166 can be decomposed into the one-port modulated R element. In realistic transistor models, it is important to account for the operating temperature [8]" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000937_1.4025818-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000937_1.4025818-Figure3-1.png", "caption": "Fig. 3 The direction of hole i in the mine frame", "texts": [ " Instead, the procedure goes back to the start point to re-produce another set of h1 and h5 using two random functions. In order to solve the inverse kinematics, the position and orientation of the end-effector needed to drill each hole on the mine face as shown in Fig. 2 should be determined first. The position of hole i can be given as hm i \u00bc \u00bdxm i ; y m i ; z m i T in the mine frame Omx- mymzm with Omxmzm lying on the mine face. The direction ei of hole i can be defined by two slant angles bi and ci as shown in Fig. 3, where bi is the angle rotated about the xm axis from the ym axis to the projection of ei in the Omymzm plane, and ci is the angle rotated about the zm axis from the ym axis to the projection of ei in the Omxmym plane. Hence, ei can be given in the mine frame as ei \u00bc \u00bd srisci; sricci; cri T (17) where ri \u00bc tan 1\u00f01=tanbicosci\u00de for bi 0, or ri \u00bc tan 1\u00f01= tanbicosci\u00de \u00fe p for bi < 0. Also, ei can be produced in the following process: the unit vector that is initially coincident with the ym axis firstly rotates by an angle /i around the xm axis" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000683_peds.2013.6527083-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000683_peds.2013.6527083-Figure2-1.png", "caption": "Figure 2. Magnetic flux lines and density according to mover position", "texts": [ " Thus, Equation (7) and (8) represents the equation to apply (1). 2 2 1 2 2 1 ( , ) ( , ) ( , ) ( , ) i x i i x i i i x i x i i \u03c6 \u03c6 \u03c6 \u03c6 \u03c6 \u2202 \u2212 \u2212 \u0394= \u2202 \u0394 \u2212 = \u2212 (7) 2 2 2 1 2 1 ( , ) ( , ) ( , ) ( , ) i x i x x x x i x i x x x \u03c6 \u03c6 \u03c6 \u03c6 \u03c6 \u2202 \u2212 \u2212 \u0394= \u2202 \u0394 \u2212 = \u2212 (8) Table I. shows specifications of the linear motor. The input current according to the mover positions is supplied in the linear motor using the finite element analysis method. Fig. 1 show respectively magnetic flux lines and density according to the input current and mover position. Fig. 2 and Fig. 3 show respectively the flux linkage and thrust force according to the input current and mover positions. The input current range is -3A ~ 3A and moving distance range is -8.1mm ~ 8.1mm. Fig. 5 shows nonlinear electrical equivalent modeling of linear motor from (1). Inductance and motor constant of the conventional modeling is fixed constant but the improved modeling is considered the nonlinearity. Fig. 6 shows delta values of the input current and stroke using time delay 0.1msec. From these values, the nonlinear inductance and motor constant can be derived" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002278_978-1-4471-4628-5_3-Figure3.7-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002278_978-1-4471-4628-5_3-Figure3.7-1.png", "caption": "Fig. 3.7 Schematic diagram of cam drive", "texts": [ " A cam is a reciprocating, oscillating, or a rotating body which is in contact with another body and imparts reciprocating or oscillating motion to it. As the cam rotates, the follower is made to rise, dwell, and fall. The length of time spent at each of these positions depends on the shape of the cam. The rise section drives the follower upwards, the fall section lowers the follower, whereas the dwell section keeps the follower at same level. \u2022 Model with flow input: Let us consider a general cam and follower arrangement shown in Fig. 3.7. This cam follower arrangement is used to operate a valve. Let the cam is driven by a velocity V . Let at a position \u03b8 cam radius be r(\u03b8). The contact between cam and follower has contact stiffness Kc and damping Rc. Let the follower has mass m. A spring of stiffness Ks is attached to a follower as shown in figure. Let the force acting at the other end of the follower be represented by Fext which in this case will be P2 A2 \u2212 P1 A1 where P1, P2, A1, and A2 are the pressures and areas on valve faces" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001154_0954410013493055-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001154_0954410013493055-Figure2-1.png", "caption": "Figure 2. Variables used for MAV course correction.", "texts": [ " Here, the vector field, as shown in Figure 1, is defined in terms of the desired direction d as:6 d \u00bc 1 2 tan 1\u00f0ky\u00de \u00f01\u00de where y is the lateral distance of the MAV, depicted as a point on the vector field, from the desired path. It is to be noted that when y is large, the MAV is expected to approach the path at a direction 1, while the angle d decreases to zero, as y approaches the desired path. The rate of transition of course from 1 to zero is influenced by k, a constant that takes only positive values. Now, consider the MAV, shown as a point object in Figure 2, flying at a constant airspeed, Va, which is achieved by the control of its longitudinal dynamics. The desired direction to be followed by the MAV is given by d, while the actual direction followed is the course angle, , which is the deviation of the airspeed velocity vector Va from the vertical. The presence of wind with velocity W makes the MAV take a detour and hence, it travels with a changed velocity, termed as ground speed, Vg. The deviation of the wind velocity W from the horizontal is given by the angle w, while g is the angle between W and the ground speed Vg and is the yaw (heading) angle, which is defined as the angle between the vertical and Vg. These various velocity components and the associated angles are shown in Figure 2. Due to the small size of MAVs, wind disturbances become one of the major challenges faced in the navigation of MAVs. It is to be noted that the groundreferenced measurements of the ground speed Vg and the course angle , as used in Ref. 6, have been used for the controller. However, for practical purposes, the known inputs for the MAVs are the airspeed Va and the yaw angle . In order to obtain the controller inputs from the measured inputs, it is observed from Figure 2 that the angle between Vg and W, defined as g, is given by, \u00bc 2 w , 4 0 2 w \u00fe , 5 0 where is defined to be positive when measured clockwise, and w is the angle made by the wind velocity with the horizontal. Hence, the value of the at NATIONAL CHUNG HSING UNIV on April 12, 2014pig.sagepub.comDownloaded from course angle can be determined using the triangular law of vector addition as follows. Vg \u00bc ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi V2 a \u00feW2 2VaW cos q \u00f02\u00de \u00bc cos 1 V2 a \u00fe V2 g W2 2VaVg \" # \u00f03\u00de The fixed wing MAV used in the present work is assumed to have a constant altitude (h) and a constant airspeed (Va) by the control of its longitudinal dynamics" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001174_icees.2011.5725320-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001174_icees.2011.5725320-Figure1-1.png", "caption": "Figure 1 Structure of the permanent magnet synchronous machine", "texts": [ " Samarium cobalt magnet has the advantages such as high saturation flux density and low specific core loss which are the most important factors that have influence on the generator performance, hence it is used for the simulation [4]. In this paper, the basic operation of PMSG is discussed in section II and section III presents simulation of SMPMSG. Sections IV and V presents the effect of varying length and diameter and the simulation results respectively. Section VI presents the conclusions II. PRINCIPLE OF THE PMSG The cross-sectional view of a four pole surface mounted permanent magnet generator is shown in Fig.1. The stator carries a three-phase winding, in which three phase emf is induced [5]. The magnets are mounted on the surface of the rotor core. They have the same role as the field winding in a synchronous machine except that their magnetic field is constant and there is no control on it [3]. When the rotor is rotated by external means, the magnetic field of the permanent magnet cuts the three phase winding and emfs are induced in the three phase windings of the stator. III. SIMULATION OF SMPMSG The PMSG dimensions are calculated as follows" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002682_msec2011-50018-Figure5-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002682_msec2011-50018-Figure5-1.png", "caption": "Fig. 5: FEM model of the cover and its boundary conditions.", "texts": [], "surrounding_texts": [ "Damping elements are placed on the rear part of the cover segment. In covers that are not equipped with a guiding scissor mechanism the damping elements damp force shocks occurring when the segments hit each other during cover spreading or folding. The damping properties of a group of damping elements are influenced by element material, shape and total element number in the group. The damping elements are usually made of rubber, polyurethane or polymeric foams. The damper submodel describes a damping element made of CELLASTO polymeric foam. The submodel is based on the Ogden hyperelastic constitutive model for foam rubbers. The material parameters have been identified by means of experiments, as will be shown in the next chapter \u201cShock absorber submodel\u201d." ] }, { "image_filename": "designv11_101_0000198_indin.2013.6622905-Figure12-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000198_indin.2013.6622905-Figure12-1.png", "caption": "Fig. 12. 3D virtual assembly of the ControlledBioBall.", "texts": [], "surrounding_texts": [ "adapter. Figure 8 shows the system architecture. In this project, the PCB solution will be used only in the future for controlling the ControlledBioBall due to its low cost, since electronic devices are cheap and highly flexible because is easy to add additional features.\nThe selected solution used for controlling the motor A28L was the commercial R/C brushless controller because is the cheapest and easiest solution that can be applied in this phase of conceptual design of the ControlledBioBall. The controller used was the commercial TowerPro Brushless Speed Controller [10] (fig. 9). It can send PWM pulses up to 8 kHz with a power supply of 10 VDC.\nHowever, for future implementation, PCB electronic circuit was already developed to control the motor without Hall sensors. Figure 10 shows the developed electronic circuit that will be used in the future to control the sensor less brushless DC Motor A28L. The signals from Hall sensors are replaced by the back EMF voltages from coils of the motor that are read by ADC ports. Resistances are used to ensure the back EMF in the range from 0 V to 5 V for ADC reading.\nThe electronic circuit presented in figure 10 was simulated with the Proteus software. The Proteus software [11] provides models of components including the brushless DC motor and the microprocessor. After the control program is compiled into hex format, it was loaded into the model of the microprocessor in Proteus software for simulating the system operation. The simulation showed that the program can control the brushless DC motor, display the speed and read ADC to change the PWM duty cycle. It is noted that the value of the speed displayed in the LCD module is not correct because the simulation is not running in real time and there are differences about the clock pulses. Practical experiments must be done to do necessary modifications.\nIII. MECHANICAL DESIGN\nBy using a motor as a spinning mass, the rotation movement of the mass is kept under control. Therefore, the ring and the groove in the Powerball\u00ae whose function is to maintain and speed up the mass become unnecessary and are eliminated in the mechanical design of the ControlledBioBall.\nThe ControlledBioball was developed with the reuse of the typical Powerball\u00ae shell (diameter of 70 mm) as in figures 11 and 12, respectively, in 2D and 3D. In this version, the out runner brushless motor A28L (fig. 5) will be used. The design allows the use of an additional mass (9) to change the total weight of the spinning mass. The left shell (1) and right shell (2) are from the normal Powerball\u00ae and the left shield (12) is added to cover the motor inside. The plate (6) will be used for mounting the motor and will be at the position of the ring and the groove in the original Powerball\u00ae. There are two screws in the shell of the normal Powerball\u00ae that can be used to fix the plate (6).", "IV. SOFTWARE DESIGN\nThe implementation of the test architecture began with the development of software using the graphical programming language LabVIEW\u00ae [12-13]. This software was developed based on a computing program to allow the generation of the PWM duty cycle that is sent to the motor controller, and thus, to control the speed motor rotation.\nThe developed software allows the following main motor\ncontrol functions:\n- To create a ramp of acceleration, in that the increment of\nthe motor speed is controlled;\n- When reaching the required maximum speed, this is\nmaintained until opposite command;\n- To create a slowing down ramp, in that the decrement of\nthe motor speed is controlled.\nIn the developed LabVIEW\u00ae software constant and adjustable parameters exist, that they are presented, respectively, in the tables 3 and 4. The constant variables are initialized by the software and maintained fixed. The adjustable variables can be modified in the software by the user according the mode and time of training intended.\nThe maximum motor speed can take any value until 10000 rpm, what corresponds (theoretically) to a duty cycle of 10%. The increment of the motor speed is accomplished indirectly through the duty cycle increment of 0.01% each 10 ms.\nThere are also the following options that the user can use: - To define the deceleration motor rate by introducing a value or used the automatic deceleration rate of the motor controller;\n- To select the possibility of saving the main data of the\nmotor test;\n- To stop the motor, without being reached the pre-defined\ntime of operation.\nThe software front panel or user interface contains the\nparameters presented in the table 5 and is shown in figure 13.", "This software was tested linking the data acquisition board NI DAQPad-6015 for USB (from National Instruments) to the PC. Following the motor and the controller referred previously were connected to the acquisition board in order to receive the data sent by the developed software.\nThe first experimental tests that were performed suggested that the overall system of the ControlledBioBall developed, that it includes the mechanical part, motor DC brushless out runner, commercial R/C brushless controller and USB data acquisition board with LabVIEW software is adequate to overcome the limitations of the Powerball\u00ae existent in the market, and this way, improving the effectivity of the wrist rehabilitation processes. However they will be necessary to perform more experimental tests, with the insertion of an incremental encoder to allow to measure and to control in closed-loop the rotation speed of the motor DC. Also, it will be necessary afterwards to be checked and validated in real conditions, following systematic tests established by specialized medical specialists and physiotherapists\nFigures 14 and 15 show the developed ControlledBioBall,\nrespectively, with separate and mounted main components.\nV. CONCLUSIONS AND FUTURE WORK\nAn investigation was carried out in order to design and develop a new type of Powerball\u00ae named ControlledBioBall\nto improve the effectivity of the wrist rehabilitation processes. This will have adjustable speed given by an electric motor inside the ball, and by this manner the manual start and the requirement of minimum torque will be eliminated.\nAccording to the performed conceptual design a mechanical versions for the system was proposed for the new ControlledBioBall device due to the easy-to-implement characteristics of its mechanical design. Regarding the electronic design, the actuation system will be performed based on a sensor less motor, more complex to control than if an hall sensor motor was used, but the performed simulation of the developed electronic circuits, proved that they were well appropriate to the performance of the adjustable speed motor required for the device.\nThe first experimental tests performed suggested that the overall system of the ControlledBioBall developed, is well suitable to reduce the limitations of the existing type of Powerball\u00ae in the market, and this manner, increases significantly the wrist rehabilitation performance.\nThe future work will be focus initially on the insertion of an incremental encoder to allow to measure and to control in closed-loop the rotation speed of the motor DC. Afterwards, it will be necessary to be checked and validated in real conditions, following systematic tests established by specialized medical specialists and physiotherapists. During this validate phase will be carried out all the necessary corrections and modifications to ensure the proper operation of the ControlledBioBall.\nREFERENCES\n[1] http://www.powerballs.com/tour.php?m=Tour [accessed June 5th, 2012]\n[2] R. Deimel. Mechanics of the gyroscope - The dynamics of rotation, Dover publications, 1950, USA.\n[3] http://www.aviationgroundschool.com/sample_pages/gyro.html\n[accessed May 5th, 2012].\n[4] http://www.mydfx.com/ [accessed May 5th, 2012].\n[5] http://www.kernpower.de/ [accessed May 5th, 2012]. [6] http://www.hobbyking.com/hobbyking/store/uh_viewitem.asp?idproduct\n=8474 [accessed July 25th, 2012].\n[7] G.Yarrish. Getting Started in Radio Control Airplanes. Model Airplanes News, 2000, USA.\n[8] M. Benfield. Radio-Control Car Manual: The complete guide to buying,\nbuilding and maintaining. Haynes Publications, Inc. Newbury Park, California, 2008, USA.\n[9] http://www.youtube.com/watch?v=CWItbhW7b6c&NR=1 [accessed\nMay 5th, 2012].\n[10] http://www.hobbyking.com/hobbyking/store/uh_viewItem.asp?idProduc t=659 [accessed January 5th, 2013].\n[11] http://www.proteussoftware.com/ [accessed December 29th, 2012].\n[12] G. Johnson, R. Jennings. LabVIEW Graphical Programming. McGrawHill Professional Publishing, 2001, USA.\n[13] B. Mihura. LabVIEW for Data Acquisition. Prentice Hall PTR, 2001,\nUSA.\n336\nPowered by TCPDF (www.tcpdf.org)" ] }, { "image_filename": "designv11_101_0000476_amm.419.795-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000476_amm.419.795-Figure3-1.png", "caption": "Fig. 3 Sketch map to the stairs\u2019 location", "texts": [ " q1 can be defined as the angle rotated counterclockwise from BA to BE, q2 and q3 can be defined as the angles of the left and the right front flippers rotated clockwise from CA to CD and to C'D' respectively, q4 and q5 can be defined as the angles of the left or the right driving wheels rotated clockwise around the back flipper BE respectively. The above can be seen as the WT wheelchair robot's generalized coordinate variables. q6 and q7 can also be defined as the angles measured by the two-dimensional tilt sensor installed on WT wheelchair robot. The rest are structural parameters that can be measured. A single stair's height is h=0.15(m), width is b=0.32(m), and stairs' slope angle can be defined as arctan h b \u039b = . If connecting the stair vertex, we can get the stair slope line, as shown in Fig. 3. We can draw straight lines through stair vertex perpendicular to stair slope line, and then we can label the stair from bottom to up in turn. When the driving wheel's center is located in between two certain straight lines, we can use the smaller number of stair to calibrate the robot. A complete process of stair-climbing can be decomposed into a number of segments, from the E wheel's center running after a calibration line to the E wheel's center running after the next calibration line. Meanwhile we introduce a parameter l EH= to calculate the numerical of calibration, where EH is the length of the line EH where H is the point crossed by the line EH which is parallel to the stair slope line and the line AH which is parallel to the level ground" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001230_gtindia2012-9586-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001230_gtindia2012-9586-Figure3-1.png", "caption": "Figure 3. DETAILS OF BEVEL GEAR WITH (A) CHIPPED TOOTH, (B) MISSING TOOTH, (C) WORN GEAR AND (D) MOUNTING OF ACCELEROMETER ON THE GEAR BOX", "texts": [ " Present experiments were conducted to study of the fault detection and diagnosis in gears. In MFS experimental setup, 3\u2212phase induction motor is mounted to rotate the rotor, which is connected to gear box through a pulley and belt mechanism. The motor speed can be manually control by a controller. The gear box and its assembly are illustrated in Figure 2. In the study of faults in gears, three different types of faulty pinion gears namely the chipped tooth, the missing tooth and worn gear along with a healthy gear were used (illustrated in Figure 3 (a\u2212c)). The real-time data in time domain were recorded using a tri-axial accelerometer (illustrated in Figure 3 (d) with sensitivity: x\u2212axis 100.3 mV/g, y\u2212axis 100.7 mV/g, z\u2212axis 101.4 mV/g) and data-acquisition hardware. The measurements were taken for running speed of 10 Hz to 30 Hz in the intervals of 5 Hz for each of four gears with faults and no-fault conditions. For each measurement set, 300 cycles of data with 2000 samples were taken and data were collected at the rate of 20,000 samples per second. All together 300 2000\u00d7 data points (number of data set \u00d7 number of samples) were collected for each of three directions (x, y and z directions)" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000080_cdc.2012.6425948-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000080_cdc.2012.6425948-Figure2-1.png", "caption": "Fig. 2. Schematic representation of an open-water channel with overshoot gates. The upstream water level in each pool is denoted by yi, pi represents the position of each gate, and di models disturbances caused by water offtakes onto farms or by spillage.", "texts": [ " In this paper, we assume that the interconnection among plants is acyclic or cascaded, that is, the plants can be ordered such that the dynamics of a given plant are only affected by the state or inputs of plants that come \u201cafter\u201d it. One possible application where the previous assumption holds is water distribution management of irrigation networks [15], [16]. An irrigation network is designed to take water from reservoirs and deliver it to farms by means of a network of open-water channels or canals. An open-water channel 978-1-4673-2064-1/12/$31.00 \u00a92012 IEEE 7553978-1-4673-2066-5/12/$31.00 2012 I consists of a series of pools divided by gates that regulate the flow of water from one pool to the next (see Fig. 2). For a particular kind of gate called an overshoot gate, this type of system can be modeled as a chain or string of plants where each plant represents a pool as illustrated in Fig. 3. Since a chain is a particular type of acyclic interconnection, this problem has the structure required for the application of the results established in this paper. However, it should be noted that an open-water channel is not a linear time-invariant (LTI) plant since its dynamics are described by nonlinear partial differential equations" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000444_amr.837.316-Figure8-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000444_amr.837.316-Figure8-1.png", "caption": "Fig. 8. a) The normal stress distribution in the cam for the cam and cam follower assembly and b) The shear stress distribution in the cam for the cam and cam follower assembly", "texts": [ " The shear stresses distribution in the cam with coating for the cam and cam follower both with coating case is presented in Fig. 6. b). The \u03c41 component has the value of 1,65 MPa and the \u03c42 component has the maximum value of 1,42 MPa. The stress distribution for the model without coating. In Fig. 7. a) is presented the equivalent stress distribution in the assembly between cam and cam follower. Fig. 7. b) shows the Von Mises equivalent stress distribution in the cam. The maximum value of the equivalent stress is 3.38 MPa. The normal stress distribution in the cam is presented in Fig. 8. a). The normal stress in this case has only one component (only the \u03c31compression component, the \u03c32 component has a very small value). The maximum value of the \u03c31compression component is 1,23 MPa. In Fig. 8. b) is presented the shear stress distribution in the cam. The maximum value of the shear stress is 0,53 MPa. In the case of the cam and cam follower assembly, both without coating, the stresses are lower than in the other two situations. The increase of stress when a coating is deposited happens due to the fact that the shear stresses increase at the interface between the cam and coating. Also the traction component (\u03c32) increases as value. For the cam without coating the value of the \u03c32 component is very small (not noticeably in the analyses) and for the case of cam with coating and cam follower with coating the value of the \u03c32 component is 3,38 MPa (for the cam follower without coating the \u03c32 component in the cam has the value of 2,43 MPa which is intermediate)" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000027_ijmr.2012.050101-Figure4-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000027_ijmr.2012.050101-Figure4-1.png", "caption": "Figure 4 Four-axis roughing (a) rotary pattern (b) z-offset (see online version for colours)", "texts": [ " Moreover, tooth space profile is symmetrical reducing by half the computation through mirroring. An important issue is to move both involute and trochoidal points to the same coordinate system X0OY0, whose vertical axis OY0 coincides with the symmetry axis of the gear tooth space, see Figure 1(c). Involute curve which is generated in \u03a7\u039f\u03a5 coordinate system should be rotated by an angle \u03c6inv\u2013 = \u03c6t/2\u2013 \u03b1g, whereas trochoidal curve, which is generated in XTO\u03a5T coordinate system, should be rotated by an angle v = \u03c6t/2\u2013 w, see Figure 4, through the use of rotation matrices. The angle between two sequential teeth is given as: \u03c6t = 2\u03c0/n and the angle between the tooth symmetry-axis and trochoidal O\u03a5Taxis is given as: 1 02 ,tw \u03c6 d r= \u2212 where ( ) ( )1 0 0 02 tan cos .od l B r= \u2212 \u2212\u03b1 \u03b1 \u0391ngle \u03b1g between the involute curve starting point on base circle and the tooth symmetry axis is calculated as: ( )0 0 0 02 tan .g S r= + \u2212\u03b1 \u03b1 \u03b1 Having both involute and trochoidal curves in a common coordinate system, the two points \u2013 one from each curve \u2013 which are closest to each other are identified", " In addition, the tool height restricts tooth height. Peripheral milling is the main cutting mode in this case, see Figure 3(b). In peripheral milling plunging is one way to reach the prescribed cutting depth, but this is strictly only necessary in closed pocket milling. Due to its lower permissible feed plunging is avoided in the case of gear tooth milling, since both sides of the tooth in the axial direction do allow tool entry. Rotary milling using a meander-like path around the teeth with simultaneous increments of A-axis, see Figure 4(b), may be accomplished within one rotation of the gear blank. In this case the restriction of tool height due to the restriction in tool diameter which is necessary in order to reach the narrowest gap at the trochoidal region is expected to pose problems in terms of chip load, vibrations and interference in the allowed space. So, multiple rotations are necessary to reduce axial depth of cut which leads to increased cutting time. Another possibility is to complete each tooth gap before advancing to the next. The tool enters axially, parallel to the gear axis, and after cutting the full tooth width it is moved laterally normal to the gear axis at the same Z level to cut the next pass with opposite feed direction. This is repeated until the whole gap at each Z level is cut, in which case the tool moves to the next lower Z-level and repeats the same, see Figure 4(b). When cuts at all Z levels are completed the A axis is rotated by one tooth pitch. This is the roughing strategy adopted in this work. This strategy, however, cannot be adopted in the finishing stage, because cusp crests will have the shape of a continuous linear segment extending over the whole of the tooth width, leading to micro-indentations in the collaborating gear teeth and possible material failure. A finishing path involving constant inclination of the ball ended tool relative to the surface to be machined could be materialised through continuous, yet small, rotary moves of A-axis, which seeks to keep cutting force constant, see Figure 5(a)" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000747_amm.437.152-Figure7-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000747_amm.437.152-Figure7-1.png", "caption": "Fig. 7 Corresponding finite element model", "texts": [ "1%, slip occurs, then repeat process under condition that X_2 is assigned to X_1; else if X_2\uff1eX_1, and \uff1c0.1%, slip occurs but is regarded as convergent. Finally, when each displacement of contact nodes is convergent, dynamic response of system is computed, and natural frequency can also been calculated. The test system [13] for measuring the dynamic response of blade with assembled dovetail attachment is shown in Fig.6. Centrifugal force is simulated by loading arm which is clamped at the base. Accordingly, Dynamic response and nature frequecy are calculated based on finite element model in Fig. 7 that has the same dimension and loading with the test device. In order to check the accuracy of the method presented, a relatively low centrifugal force 0.54KN is chosen. Two kinds of computations (without joint damping and with joint damping) are considered so as to evaluate the damping effects. Maximum displacement versus exciting frequency is plotted with experimental comparison in Fig. 8. Meanwhile, first order natural frequency versus centrifugal force is also plotted with experimental comparison in Fig" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001431_amr.694-697.512-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001431_amr.694-697.512-Figure1-1.png", "caption": "Fig. 1 Face gear drive coordinate system Fig. 2 Formation of flank\u2160involute helicoid", "texts": [ " When the shaft angle \u03b3=90\u00b0, the bevel gear tooth will be distributed in a circular plane, then the bevel gear is the face gear, thus known as the face gear drive. According to the relationship between the axes of the drive and driven gear, the face gear drive can be separated into four cases: orthogonal face gear, non-orthogonal face gear, offset orthogonal face gear and offset non-orthogonal face gear. The curved tooth face gear drive, which is a new kind of gear transmission pair, is that the curved tooth face gear meshes with the cylindrical worm. In this paper, the offset orthogonal face gear will be taken as the research object, as shown in Fig. 1, the pinion is the ZI-type cylinder worm, the gear is the face gear. During meshing, the face gear axis is perpendicular to the axis of the worm, and they also has a certain offset distance. That design, which is the under bias of pinion\u2019s axis, is conductive to reducing the body\u2019s center of gravity height, obtaining a larger contact ratio and greatly improving the rigidity of system by the use of straddle support [3]. In the gear meshing analysis process, using the following four coordinate systems(fig. 1): two fixed coordinate systems of the initial position of the worm 1 and the face gear 2: kjiOS ,,,00 \u2212 and , , ,p p p p pS O i j k\u2212 , two moving coordinate systems of the worm 1 and the face gear 2: 1 1 1 1 1, , ,S O i j k\u2212 and 2 2 2 2 2, , ,S O i j k\u2212 .The origins of coordinates pO and 2O coincide, as well as 0O and 1O . The axes pk and 2k coincide with the rotational axis of the face gear, and axes k and 1k coincide with the rotational axis of the worm. 1\u03c6 and 2\u03c6 represent the angle of the pinion and gear that their moving coordinates rotate to respective fixed coordinates" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000198_indin.2013.6622905-Figure4-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000198_indin.2013.6622905-Figure4-1.png", "caption": "Fig. 4. Brushless DC motor.", "texts": [ " This will have adjustable speed given by an electric motor inside the ball, and consequently the manual start and the requirement of minimum torque will be eliminated. As well, a mechatronic system composed by sensors, data acquisition board and human-machine interface will be considered to allow the process of rehabilitation to be fully controlled. II. CONCEPTUAL DESIGN To have the desired properties described previously the new Powerball\u00ae needs to have interconnected the elements presented in the table 2. Brushless DC motor (fig. 4) was selected as this motor rotates at high speed; also because commutator and brushes in DC motors wear very fast. The brushless motor can be a Hall sensor type or sensor less type. A brushless motor is constructed with a permanent magnet rotor and wire wound stator poles. Electrical energy is converted to mechanical by the magnetic attractive forces between the permanent magnet rotor and a rotating magnetic field induced in the wound stator poles. The brushless DC motor has three electromagnetic circuits connected at a common point" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001653_imece2013-63679-Figure10-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001653_imece2013-63679-Figure10-1.png", "caption": "Figure 10: Stress analysis, Team B.", "texts": [], "surrounding_texts": [ "The last step of the project was the final presentation where the students had to sell their designs and show that they met all the constraints and that the designs can compete in the market. Figure 11 and 12 show the final designs for the two teams. Comparing these final designs to the original selected designs in the options reviews, Figures 5 through 8, it is clear that each team improved the details of their design while maintaining the same concept. Team A continued with the \u201cStanding Rack\u201d design with minimal changes in the folding arms. Team B improved the \u201cBi-Fold Style\u201d significantly by having a double swing arms instead of one to decrease the needed clearance area behind the car. One further development by team B was the unique design of the hitch connection. They developed a full detailed design of a connection that can be used for 2 and 1.25 inch receivers. Figure 13 shows the connection design for team B. Figures 14 through 17 are the summaries that the teams presented with each design to show how they met the design weight and price target. Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 01/30/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use 7 Copyright \u00a9 2013 by ASME Downloaded From: http://proceedings.asmedigitalcollection.asme.org/ on 01/30/2016 Terms of Use: http://www.asme.org/about-asme/terms-of-use 8 Copyright \u00a9 2013 by ASME Figure 16: Cost details, Target Budget is $391.16, Team A. Figure 17: Cost details, Target Budget is $391.16, Team B." ] }, { "image_filename": "designv11_101_0000533_s0021894412040086-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000533_s0021894412040086-Figure1-1.png", "caption": "Fig. 1. Formulation of the problem and coordinate systems.", "texts": [ " The calculated results are fitted to available experimental data in terms of three parameters: the ratio of the shear moduli of the drop and matrix materials, the yield stress of the matrix material, and the distance covered by the drop until complete destruction of molecular bonds in the matrix. Let r, z, \u03d5 be a cylindrical coordinate system with the origin at the drop center, where the z axis is directed toward the other drop (as the drops approach each other, the velocity vector of the drop is directed along the z axis) (Fig. 1). We apply the coordinate transformation r = r0 e y sinx, z = r0 e y cosx, where r0 is the drop radius. Then the flow domain is conformally mapped onto a rectangle 0 \u2264 x \u2264 \u03c0, 0 \u2264 y \u2264 Y [Y = ln (r2/r0), where r2 is the radius of a sphere inside which the fluid moves]. The coordinate system moves together with the drop. The drop boundary is y = 0, the external boundary of the flow domain is y = Y , and x = 0 and x = \u03c0 are segments on the axis of symmetry. The dimensional parameters of the problem are the drop radius r0 = 0" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003245_b978-0-08-096979-4.00004-9-Figure4.18-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003245_b978-0-08-096979-4.00004-9-Figure4.18-1.png", "caption": "FIGURE 4.18", "texts": [ " The engine is joined to the back of this unit usually by four studs or bolts and the structure is completed by attachment of the gearbox casing to the rear face of the engine. The chassis, engine, and gearbox, therefore, form a \u2018box-beam\u2019 structure which carries the inertial loads to their reaction points at the four corners of the car as depicted in Figure 4.17. Arranged around, and attached to, these are the remaining components e wing structures, underbodies, cooler ducting, and bodywork, as illustrated in Figure 4.18. This general arrangement is exactly that as has been used by most single-seater racing cars since the 1960s. It has been mentioned that the chassis component is of major importance to the working of the structure. During the course of \u2018setting up\u2019 a racing car at a circuit, changes are made to the suspension elements (springs, dampers, antiroll bars) with the intention of modifying its handling. Ideally, any small change in a component stiffness should be felt in the balance of the car. This will not occur if the structure transmitting the loads is of insufficient stiffness" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003245_b978-0-08-096979-4.00004-9-Figure4.2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003245_b978-0-08-096979-4.00004-9-Figure4.2-1.png", "caption": "FIGURE 4.2", "texts": [ " A paper on the car written in 19851 commented that the concept stood up to examination after three years and that is largely true today; it should be remembered that the radically different technology introduced then has been adopted by a number of production vehicles referred to in the preceding chapter. The general specification, with views of the vehicle and body structure, is shown in Figure 4.1. As for current new designs (see Chapter 8), the vehicle was planned with due regard for the total energy consumed in the vehicle\u2019s life cycle as well as total vehicle ownership costs, which related to the factors shown in Figure 4.2. Apart from fuel consumption and servicing costs, two other cost-related factors were identified as being important to potential owners, corrosion and lowspeed accident damage, the latter affecting insurance costs. The body materials adopted were key to managing these factors and because of its relevance to current Contributions to improved performance1 Total ownership costs1 developments a summary of the rationale for the choices made is presented in Table 4.1. The concept of a clad substructure was not new and that of ECV 3 was similar to the base-unit used on the Rover 2000 between 1963 and 1975" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000198_indin.2013.6622905-Figure11-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000198_indin.2013.6622905-Figure11-1.png", "caption": "Fig. 11. 2D assembly drawing of the ControlledBioBall.", "texts": [], "surrounding_texts": [ "adapter. Figure 8 shows the system architecture. In this project, the PCB solution will be used only in the future for controlling the ControlledBioBall due to its low cost, since electronic devices are cheap and highly flexible because is easy to add additional features.\nThe selected solution used for controlling the motor A28L was the commercial R/C brushless controller because is the cheapest and easiest solution that can be applied in this phase of conceptual design of the ControlledBioBall. The controller used was the commercial TowerPro Brushless Speed Controller [10] (fig. 9). It can send PWM pulses up to 8 kHz with a power supply of 10 VDC.\nHowever, for future implementation, PCB electronic circuit was already developed to control the motor without Hall sensors. Figure 10 shows the developed electronic circuit that will be used in the future to control the sensor less brushless DC Motor A28L. The signals from Hall sensors are replaced by the back EMF voltages from coils of the motor that are read by ADC ports. Resistances are used to ensure the back EMF in the range from 0 V to 5 V for ADC reading.\nThe electronic circuit presented in figure 10 was simulated with the Proteus software. The Proteus software [11] provides models of components including the brushless DC motor and the microprocessor. After the control program is compiled into hex format, it was loaded into the model of the microprocessor in Proteus software for simulating the system operation. The simulation showed that the program can control the brushless DC motor, display the speed and read ADC to change the PWM duty cycle. It is noted that the value of the speed displayed in the LCD module is not correct because the simulation is not running in real time and there are differences about the clock pulses. Practical experiments must be done to do necessary modifications.\nIII. MECHANICAL DESIGN\nBy using a motor as a spinning mass, the rotation movement of the mass is kept under control. Therefore, the ring and the groove in the Powerball\u00ae whose function is to maintain and speed up the mass become unnecessary and are eliminated in the mechanical design of the ControlledBioBall.\nThe ControlledBioball was developed with the reuse of the typical Powerball\u00ae shell (diameter of 70 mm) as in figures 11 and 12, respectively, in 2D and 3D. In this version, the out runner brushless motor A28L (fig. 5) will be used. The design allows the use of an additional mass (9) to change the total weight of the spinning mass. The left shell (1) and right shell (2) are from the normal Powerball\u00ae and the left shield (12) is added to cover the motor inside. The plate (6) will be used for mounting the motor and will be at the position of the ring and the groove in the original Powerball\u00ae. There are two screws in the shell of the normal Powerball\u00ae that can be used to fix the plate (6).", "IV. SOFTWARE DESIGN\nThe implementation of the test architecture began with the development of software using the graphical programming language LabVIEW\u00ae [12-13]. This software was developed based on a computing program to allow the generation of the PWM duty cycle that is sent to the motor controller, and thus, to control the speed motor rotation.\nThe developed software allows the following main motor\ncontrol functions:\n- To create a ramp of acceleration, in that the increment of\nthe motor speed is controlled;\n- When reaching the required maximum speed, this is\nmaintained until opposite command;\n- To create a slowing down ramp, in that the decrement of\nthe motor speed is controlled.\nIn the developed LabVIEW\u00ae software constant and adjustable parameters exist, that they are presented, respectively, in the tables 3 and 4. The constant variables are initialized by the software and maintained fixed. The adjustable variables can be modified in the software by the user according the mode and time of training intended.\nThe maximum motor speed can take any value until 10000 rpm, what corresponds (theoretically) to a duty cycle of 10%. The increment of the motor speed is accomplished indirectly through the duty cycle increment of 0.01% each 10 ms.\nThere are also the following options that the user can use: - To define the deceleration motor rate by introducing a value or used the automatic deceleration rate of the motor controller;\n- To select the possibility of saving the main data of the\nmotor test;\n- To stop the motor, without being reached the pre-defined\ntime of operation.\nThe software front panel or user interface contains the\nparameters presented in the table 5 and is shown in figure 13.", "This software was tested linking the data acquisition board NI DAQPad-6015 for USB (from National Instruments) to the PC. Following the motor and the controller referred previously were connected to the acquisition board in order to receive the data sent by the developed software.\nThe first experimental tests that were performed suggested that the overall system of the ControlledBioBall developed, that it includes the mechanical part, motor DC brushless out runner, commercial R/C brushless controller and USB data acquisition board with LabVIEW software is adequate to overcome the limitations of the Powerball\u00ae existent in the market, and this way, improving the effectivity of the wrist rehabilitation processes. However they will be necessary to perform more experimental tests, with the insertion of an incremental encoder to allow to measure and to control in closed-loop the rotation speed of the motor DC. Also, it will be necessary afterwards to be checked and validated in real conditions, following systematic tests established by specialized medical specialists and physiotherapists\nFigures 14 and 15 show the developed ControlledBioBall,\nrespectively, with separate and mounted main components.\nV. CONCLUSIONS AND FUTURE WORK\nAn investigation was carried out in order to design and develop a new type of Powerball\u00ae named ControlledBioBall\nto improve the effectivity of the wrist rehabilitation processes. This will have adjustable speed given by an electric motor inside the ball, and by this manner the manual start and the requirement of minimum torque will be eliminated.\nAccording to the performed conceptual design a mechanical versions for the system was proposed for the new ControlledBioBall device due to the easy-to-implement characteristics of its mechanical design. Regarding the electronic design, the actuation system will be performed based on a sensor less motor, more complex to control than if an hall sensor motor was used, but the performed simulation of the developed electronic circuits, proved that they were well appropriate to the performance of the adjustable speed motor required for the device.\nThe first experimental tests performed suggested that the overall system of the ControlledBioBall developed, is well suitable to reduce the limitations of the existing type of Powerball\u00ae in the market, and this manner, increases significantly the wrist rehabilitation performance.\nThe future work will be focus initially on the insertion of an incremental encoder to allow to measure and to control in closed-loop the rotation speed of the motor DC. Afterwards, it will be necessary to be checked and validated in real conditions, following systematic tests established by specialized medical specialists and physiotherapists. During this validate phase will be carried out all the necessary corrections and modifications to ensure the proper operation of the ControlledBioBall.\nREFERENCES\n[1] http://www.powerballs.com/tour.php?m=Tour [accessed June 5th, 2012]\n[2] R. Deimel. Mechanics of the gyroscope - The dynamics of rotation, Dover publications, 1950, USA.\n[3] http://www.aviationgroundschool.com/sample_pages/gyro.html\n[accessed May 5th, 2012].\n[4] http://www.mydfx.com/ [accessed May 5th, 2012].\n[5] http://www.kernpower.de/ [accessed May 5th, 2012]. [6] http://www.hobbyking.com/hobbyking/store/uh_viewitem.asp?idproduct\n=8474 [accessed July 25th, 2012].\n[7] G.Yarrish. Getting Started in Radio Control Airplanes. Model Airplanes News, 2000, USA.\n[8] M. Benfield. Radio-Control Car Manual: The complete guide to buying,\nbuilding and maintaining. Haynes Publications, Inc. Newbury Park, California, 2008, USA.\n[9] http://www.youtube.com/watch?v=CWItbhW7b6c&NR=1 [accessed\nMay 5th, 2012].\n[10] http://www.hobbyking.com/hobbyking/store/uh_viewItem.asp?idProduc t=659 [accessed January 5th, 2013].\n[11] http://www.proteussoftware.com/ [accessed December 29th, 2012].\n[12] G. Johnson, R. Jennings. LabVIEW Graphical Programming. McGrawHill Professional Publishing, 2001, USA.\n[13] B. Mihura. LabVIEW for Data Acquisition. Prentice Hall PTR, 2001,\nUSA.\n336\nPowered by TCPDF (www.tcpdf.org)" ] }, { "image_filename": "designv11_101_0001172_ijmee.39.4.7-Figure5-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001172_ijmee.39.4.7-Figure5-1.png", "caption": "Fig. 5 Free-body diagram.", "texts": [ "comDownloaded from International Journal of Mechanical Engineering Education, Volume 39, Number 4 (October 2011), \u00a9 Manchester University Press exerted by the surface onto the cylinder acts to the left with a larger magnitude than that of the applied force to the right. Let us consider the equivalent situation of a rope pulling on a large cylinder, as in Fig. 4. The rope lies in a groove on the outer edge of the cylinder. A small lip on the cylinder supports the cylinder on the surface. Since we will assume that the cylinder rolls, we must include a force of friction, Ff, in the free-body diagram, as shown in Fig. 5 [4]. The cylinder, of mass m, has a weight mg acting at the center of mass, O. A rope with an applied force, Fa, at the outer surface pulls the cylinder. A frictional force, Ff, and a normal force, N, act on the cylinder at the point of contact, C, where the cylinder touches the surface. The radius from the center of mass to the outer surface of the cylinder (where the force of the rope is applied) is ra. The radius from the center of mass to the lip of the cylinder (which touches the surface where the frictional force is applied) is rf", "comDownloaded from International Journal of Mechanical Engineering Education, Volume 39, Number 4 (October 2011), \u00a9 Manchester University Press Substituting equation 1 into equation 2, we have: a r F r r I O x C , ( ) = \u2212\u239b \u239d\u239c \u239e \u23a0\u239fa a a f (3) The summation of forces in the x-direction is equal to the cylinder\u2019s mass times the acceleration component aO,x (i.e. \u2211 =F max ). Taking positive x to the left: F F m r F r r IC f a a a a f\u2212 = \u2212\u239b \u239d\u239c \u239e \u23a0\u239f ( ) (4) Therefore we have: F m r F r r I F F mr r r I F C C f a a a f a a a a f a= \u2212\u239b \u239d\u239c \u239e \u23a0\u239f + = \u2212( ) +\u239b \u239d\u239c \u239e \u23a0\u239f > ( ) 1 (5) since all quantities in the fi nal pair of brackets are positive. This shows that Ff is greater than Fa (as depicted by the magnitudes shown in the free-body diagram, Fig. 5). Therefore the cylinder moves to the left because the frictional force is greater than the applied force! Finally, it is interesting to consider the torques about O: F r F ra a f f\u2212 > 0 (6) This torque causes the cylinder to rotate counterclockwise (just as a torque about point C did). Since we have already shown that Ff > Fa, then we see that the relationship Fara > Ffrf is possible only because ra > rf, which is a rather unusual condition. When considering how an object will roll in response to an applied force, it is critical to consider the frictional force, which can exceed the applied force" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001647_amm.52-54.583-Figure8-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001647_amm.52-54.583-Figure8-1.png", "caption": "Figure 8. Stress of model before and after optimization", "texts": [ " So that, we choose design variables are thickness of disc and radius of the rigid link two discs of brake as figure 6. We choose design variables as the radius initial of rigid equal 2mm and thickness of two discs equal 7mm but when we limit the optimization by minimum stress constrains, that the result shown that new structure of disc brake which radius of rigid equal 2.8mm and thickness of a disc equal 6.2mm. The result also shown the stress on the model is decrease from 9.76N/mm 2 to 7.52 N/mm 2 as on the figure 7. The stress occurs at rigid are shown on figure 8. In this paper, a study about sensitivity analysis of pressure and friction coefficient effect to stress of disc brake was presented. It is comfortable for new design or problem of structure optimization. The result was shown that the stress of model is reduced approximate 22.95%. And critical locations on the disc brake geometry are located at the rigid link between two discs. The actual model is used for analysis as on figure 9. This is a study for pre studies about dynamic or vibration and noise of disc brake" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002682_msec2011-50018-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002682_msec2011-50018-Figure1-1.png", "caption": "Fig. 1: Detail of sealing wiper mounted in a steel profile.", "texts": [ " The model takes into account the deformation of the cover segments due to wiper preload, and the sliding resistance of the guiding elements. The main input data for the model include cover dimensions, cover travel and maximum running speed. The output information is the passive force of the cover, depending on the position and moving speed of the cover. Component submodels constitute important parts of the model. These will be described in the following chapters. Sealing wipers are usually made of polyurethane. A typical wiper profile consists of three parts: lip, head and body (see Fig. 1). The lip is the main seal part of the wiper. The lip is flexible and it is able to seal the contact with the lower cover segment (not shown in the figure) if the surface of the latter is not perfectly planar. The wiper head can shift within the profile. It increases the real lip stroke for better sealing. The wiper is fastened in its supporting steel profile through the body. If the upper cover segment is overloaded, this results in maximum lip and head deformation and the body comes into contact with the lower cover segment" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002790_ijeee.50.1.5-Figure4-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002790_ijeee.50.1.5-Figure4-1.png", "caption": "Fig. 4 Cart-pole testbed.", "texts": [ " For the linear system (16) with a cost function defi ned as J Q u Ru tT T= + \u221e \u222b ( )x x d 0 (17) the state feedback control law that minimises the value of the cost J is u T= \u2212k x (18) here k is a coeffi cient vector, given by k = R\u22121BTW and W is found by solving the Riccati equation ATW + WA \u2212 WBR\u22121BTW + Q = 0. With the prescribed Q and R, k can easily be solved by a MATLAB command LQR(A, B, Q, R). In order to confi rm and evaluate the designed method, experiments of automatically setting an initial angular deviation of the pole will be illustrated. The physical experiments are demonstrated by the cart-pole testbed in Fig. 4 produced by Googol Technology (HK) Ltd. In Fig. 4, the control input is offered by a motor and applied to the system through a gearbelt component. Physical parameters of this equipment are determined as M = 1.096 kg, m = 0.109 kg, l = 0.25 m, g = 9.81 m\u00b7s\u22122, and Jp = 0.0034 kg m2. From the initial downward position to the fi nial upright one, energy deviation of the pole is 2mgl, so that we have E0 = 2mgl. Another controller parameter for swinging up the pole is picked up as \u03b1 = 0.41. Predefi ning at University of Hawaii at Manoa Library on June 15, 2015ije" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001173_indin.2012.6301231-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001173_indin.2012.6301231-Figure1-1.png", "caption": "Figure 1. Basic structure of rolling bearings", "texts": [ " Support vector machine classifier is used to carry out multi-class fault classification of the extracted feature vectors. Experimental results show that support vector machine can realize agency risk minimization and overcome problems such as local minimum points, and it has a high fault classification accuracy rate [5]. II. FAILURE MECHANISIM OF ROLLING BEARINGS 1) Basic structure of rolling bearings A rolling bearing commonly consists of inner race, outer race, rolling element and retainer, as shown in Fig. 1. 2) Rolling bearing fault characteristics and corresponding vibration frequency If the rolling surface is working under poor lubrication, rolling bearing might fall into damage in a short 978-1-4673-0311-8/12/$31.00 \u00a92012 IEEE 546 period of time, especially when it is overloaded or working at high speed. In addition, when load changes, the tiny crack on rolling element produced in manufacturing or in use can also make the cracks expand or even produce local falling off, leading to fatigue pitting" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002165_icnc.2011.6022262-Figure6-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002165_icnc.2011.6022262-Figure6-1.png", "caption": "Figure 6. moving of robot along local path", "texts": [ "375 and 6.00 respectively. Given that 2 2 2 2 ( ) ( ) 0.79r r o oS t v S t v , smaller than 2t , and 3 3 3 3 ( ) ( ) 3.80r r o oS t v S t v , smaller than 3t , but 4 4 4 4 ( ) ( ) 12.63r r o oS t v S t v , larger than 4t , the robot will collide with obstacle 2 and 3, so local robot path planning is needed and the robot should collide free obstacle 4 when moving along the local path. The local robot path after planning according to the method in subsection 3.2 is represented with the broken line in Figure6 (a). The scenario of the moving of the robot along the local path is shown as Figure 6 (b) (c) (d); we can observe that the robot passes the intersection point before the obstacles, avoiding a collision between the robot and the obstacles. VI. CONCLUSION We study the problem of robot path planning in an environment with dynamic obstacles and propose a method of local path planning with the prediction of the trajectory of an obstacle being a band. Our performance analysis and simulation results in various scenarios confirm that the band trajectory of an obstacle decreases the probability of collision between the robot and the obstacle greatly" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001194_1.1716244-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001194_1.1716244-Figure1-1.png", "caption": "FIG. 1. Schematic drawing of vacuum terminal.", "texts": [ " This note is published with the permission of the South African Council for Scientific and Industrial Research. Seattle 5, Washington (Received February 7, 1958) CERTAIN applications require a brakin\u00a3 action to be applied to a motor-driven device upon completion of a work cycle, and to be maintained until initiation of the following cycle. Although mechanical brakes and Geneva movements are excellent for this purpose, there are times when their cost or weight renders them undesirable, and magnetic braking is used. The conventional half-wave rec tifier braking circuit of Fig. 1 is probably the most econom- AC FIG. I. R ical of parts, but, because of the pulsating nature of the direct current, the braked motor tends to \"creep,\" and has very limited holding power. Also, a certain amount of 6O-cy noise is developed, which may be objectionable. The circuit of Fig. 2, using one additional rectifier, overcomes the above disadvantages. When the switch is in the \"stop\" position, alternate pulsations of line current flow through SRI and the motor windings, as in Fig. 1. When the line current reverses, however, the collapsing magnetic field of the motor maintains the current flow in the same direction through the windings and through SR2\u2022 Thus, a much steadier flow of direct current is main tained through the windings. Creepage is eliminated, hold ing power is greatly increased, and noise is minimized. Using this circuit with small motors, such as the Barber Colman series, it was found that current in the SR2 leg was about 10% lower than in the SRI leg. Therefore, the two rectifiers may be of equal value, with R selected to limit the current to a value that will prevent overheating", " This procedure ensures that no oil will accumulate at the joint of the capillary with the chamber. However, oil which has stuck in the closed capillary can easily be re moved. While the gauge is under high vacuum, the plunger is immersed until the oil rising in the compression capillary reaches the oil pushed to the closed end, Any bubble of gas will rise to the top, therefore causing the oil column to become uninterrupted. A quick increase in the magnet izing current will give the plunger a sudden downward FIG. 1. Oil McLeod gauge. (a) TO VACUUM SYSTEM GLASS ENVELOPE PLUNGER IRON~ ZERO LEVEL COMPARISON--CAPILLARY COMPRESSION (b) CAPILLARY SCALECOMPRESSION CHAMBER APIEZON-e OIL (100 ee) HEATER motion, which will help the upward removal of the gas bubble. The readings are quite reproducible. The instrument has been used mainly for calibrations, and covered the range from 10--5 to 10--1 tor. If kept under vacuum, no difficulty is experienced because of gas ab sorption. The fact that the gauge is small in size, and is light, and does not require a liquid-air trap and an addi tional pump, highly recommends its use" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000659_amm.312.42-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000659_amm.312.42-Figure3-1.png", "caption": "Figure 3 Engagement principle of elliptic gear Figure 4 The series combined mechanism of elliptic gear and geneva", "texts": [ " Kinematic analysis on elliptic gear The shift gear of elliptic gear was shown in Figure 2, and the driving force of this mechanism was input by the axis O1, elliptic gear 1 and elliptic gear 2 were two different ellipses with major axis 2a1=2a2=2a, minor axis 2b1=2b2=2b, focal length 2c1=2c2=2c, and their focus was O1, O2, O1O2=2a respectively, then the centrifugal rate: 222111 // aceace === , when the elliptic gear 1 rotated around point O1 at constant speed, and elliptic gear 2 rotated around point O2 at variable speed, then this elliptic gear driven instantaneous transmission ratio was as below: )cos2(1/)(1/i 2 1 2 1221 eee \uff0b\uff0b\uff0d \u03d5\u03c9\u03c9 == (4) The instantaneous angular acceleration of elliptic gear 2: 22 11 22 12 )c2(1/sin)-(12 eoseee \uff0b\uff0b \u03d5\u03d5\u03c9\u03b5 = (5) The kinematic analysis on series combined mechanism of elliptic gear and geneva The series combined mechanism of elliptic gear and geneva was shown in Figure 3. The driving force of this mechanism was input by axis O1. When elliptic gear rotated with constant speed around O1, due to the same axis of drive plate 3 and driven elliptic gear 2, drive plate 3 rotated at variable speed around axis O2. Assuming that the centering distance O2 O4 of Geneva mechanism was L, the radius of drive plate was r3, the distance between the coincidence point on the geneva and the centering of the geneva was r4. When the drive plate just entered the slot, r3 and r4 was perpendicular to each other, then \u03bb\u03c0 == )/sin(/3 ZLr \uff0c ))2/((2))2/((2 10 1 1 1 30202303 \u03d5\u03d5\u03d5\u03d5\u03d5\u03d5\u03d5 tgtgtgtg \u2212\u2212 +\u2212=+\u2212= " ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001638_s1068798x11110128-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001638_s1068798x11110128-Figure1-1.png", "caption": "Fig. 1. Model of chevron transmission.", "texts": [ " 11 2011 REDUCING THE VIBRATION OF A CHEVRON GEAR TRANSMISSION 1053 The generalized model of vibration in a helical coupling is incorporated into the selected simulation method, which also includes a dynamic model of the transmission; a set of modifications to the equations of motion during the engagement period; a logical algo rithm for regulating the selection of these modifica tions within the engagement period; and appropriate software. In helical couplings, on account of the inclination of the teeth, three dimensional vibrations are excited at engagement. To simplify the calculations, while retaining the basic relation between the excitation of torsional, transverse, and axial vibrations, we consider a model in which each gear is a solid with four degrees of freedom: axial and radial motion, rotation in the engagement plane, and torsional vibrations. The dynamic model of the chevron transmission (Fig. 1) consists of two absolutely solid bodies: gears 1 and 2, which are mounted on elastic bearings with rigidities Cxi and Czi (i = 1, 2) in the transverse and axial directions and are connected by springs Ck. Each spring Ck simulates the rigidity of engagement of tooth pair k. The springs lie in the engagement plane; each one corresponds to a specific contact line and moves together with this line over the engagement field as the gears turn. Each spring is spatially oriented along the line of action of normal force Pk in the coupling" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001101_1369-4332.16.11.1871-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001101_1369-4332.16.11.1871-Figure2-1.png", "caption": "Figure 2. Photographs of test samples ready for testing", "texts": [ " 3rd batch samples semi apical angles were changed to 5 degree and they were designated as A5_B1_S1\u2026. etc. The 4th batch samples were radial scaled down of 3rd batch samples. They were designated as A5_B2_S1\u2026 etc. The top cone has 2 mm difference for its bottom and top diameter and hence it looks almost like cylinder. This difference was kept ~8 mm for batch A5_B2 to facilitate smooth inversion and to avoid any frictional contact between top and middle cone. Few aluminum samples from each batch are shown in Figure 2, after spinning and ready for testing. The dimensions of tested samples are listed in Table 1 along with their results. The dimensions schematic is shown in Figure 3(a). The typical dimensions from each batch sample are shown in Figure 3(c). After spinning, the thickness may vary along with height in a sample. A typical cut section is shown in Figure 4, for sample A10_B1_S1. It was noticed that the cap end and bottom conical end (height 25 mm to 35 mm from bottom) had more thickness. Other portion had comparatively less thickness and its variation is also less" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000966_sces.2012.6199029-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000966_sces.2012.6199029-Figure1-1.png", "caption": "Fig. 1. Twin rotor control system (TRCS) The dynamics of the TRCS can be approximately represented in the state-space form as follows.", "texts": [ " The simulation results obtained reveal that the proposed controller gives good tracking performance. The paper is arranged as follows. In Section II, the modeling of TRCS is introduced. The problem statement is introduced in Section III. In Section IV, neuro sliding mode controller design is proposed. Section V presents the simulation results. Finally conclusions are given in the Section VI. II. MODELING OF TRCS The mechanical setup of the TRCS is a multivariable, nonlinear and strongly coupled system with degrees of freedom on the pitch and yaw angles as shown in Fig. 1. The complete dynamics of TRCS is given in [12]. Neuro Sliding Mode Controller for Twin Rotor Control System Bhanu Pratap, Student Member, IEEE S ( ) ( ) ( ) ( ) ( ) [ ] 2 21 1 1 1 1 1 1 1 1 2 1 1 1 1 1 1 12 22 2 2 2 1 1 1 1 2 2 2 2 10 1 1 1 1 11 11 20 2 2 2 2 21 21 0.0326sin sin 2 2 cos 1.75 g gy c T d dt Ma bd dt I I I I kB a b I I d dt Ba bd k a b dt I I I I T kd u dt T T T kd u dt T T y \u03c8 \u03d5 \u03c8 \u03c8 \u03c8 \u03c4 \u03c4 \u03c8 \u03c8 \u03d5 \u03c8 \u03c8 \u03d5 \u03c4 \u03c4 \u03d5 \u03d5 \u03d5 \u03c4 \u03c4 \u03d5 \u03c4 \u03c4 \u03c4 \u03c4 \u03c4 \u03c4 \u03c8 \u03d5 \u23ab= \u23aa \u23aa \u23aa= + \u2212 + \u23aa \u23aa \u23aa \u2212 \u2212 + \u23aa \u23aa \u23aa = \u23ac = + \u2212 \u2212 + = \u2212 + = \u2212 + = \u23aa\u23aa \u23aa \u23aa \u23aa \u23aa \u23aa \u23aa \u23aa \u23aa \u23aa \u23aa \u23aa\u23ad (1) where \u03c8 & \u03d5 are pitch (elevation) & yaw (azimuth) angle; \u03c8 & \u03d5 are angular velocity across pitch & yaw angle; 1\u03c4 & 2\u03c4 are Momentum of main & tail rotor and 1u & 2u are voltage applied to main & tail rotor respectively" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001300_s1068798x13100079-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001300_s1068798x13100079-Figure2-1.png", "caption": "Fig. 2. Rotation of sample around center P.", "texts": [ "= The position of the skew axis in the xOy plane may expediently be determined by specifying the point D, at the end of the vector t, on the axis xt = xD = \u2013\u03b1\u03b4c/\u03b8 2; yt = yD = \u2013\u03b2\u03b4c/\u03b8 2. The distance from the coordinate origin to the skew axis is equal to the length of vector t: |t| = \u03b4c/\u03b8. If the application of force Fz and torques Mx and My to the sample is followed by the application of rotary torque T in the junction plane, the sample will turn by an angle \u03d5 around the center of rotation P with coor dinates xP, yP. The center of rotation (Fig. 2) is shifted relative to the center of gravity of the contact surface in the direction of the large pressure, by an amount that increases with increase in Mx and My. On the assumption that the sample\u2019s lower surface is not deformed by torque T, the tangential displace ment \u03b4 \u03c4 at each point (x, y) may be determined as \u03b4 \u03c4 (x, y) = \u03d5r(x, y), where r(x, y) is the length of the radius vector from the center of rotation to point (x, y) r(x, y) = [(x \u2013 xp) 2 + (y \u2013 yp) 2]0.5, With no local slip, the tangential stress \u03c4 at each point depends linearly on the tangential displacement [3] (4) The tangential stress at each point is perpendicular to the radius vector from the center of rotation" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000552_amm.86.688-Figure5-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000552_amm.86.688-Figure5-1.png", "caption": "Fig. 5 Residual stress on machined gear surface( feed of 1mm)", "texts": [ " Meanwhile, the origin of the cutter rotated around the origin of the whole machine tool. The roughcast is also defined to rotate around axis of the gear. Their rotating velocities honor the roll ratio. In this paper, analysis was performed with cutting speed and cutting depth as variables. Velocities of 35m/min, 30m/min and 45m/min were considered, with a total cutting depth of 1mm. Then Von Mises stress of every simulation was analyzed. The schemas are listed in Table 3. Simulation result and analysis of milling surface residual stress on Spiral bevel gear Representatively, Fig. 5 shows equivalent residual stress on machined gear surface from the simulations in cutting depth, and the equivalent residual stress on gear surface of each results in simulation group1 were read and averaged, the data were shown in Fig. 6 and Fig. 7. Fig. 8 shows the plastic flow on chip and workpiece. The tangent stress along the cutting direction is mainly compression in the shear zone. In the area closest to the tip of cutter, where was surrounded by the line A, tangent stress reached the highest value, roughly in -2000Mpa" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003001_detc2013-12231-Figure9-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003001_detc2013-12231-Figure9-1.png", "caption": "Fig. 9 Detailed contact situation", "texts": [ " In this way the instantaneous contact load distributions in a whole mesh cycle are obtained as shown in Fig. 8. Then, the boundary condition of gears is interchanged and the loaded tooth contact analysis is conducted again. \u9f7f1 \u9f7f2 \u9f7f3 \u9f7f1 \u9f7f2 \u9f7f3 (a) 0 \u00b0 (b) 2 \u00b0 From the results, the simultaneous meshing gear pairs are alternatively two or three during one mesh cycle. The maximum contact pressure, 3969 MPa, is located in the tip region when a tooth meshes out. It is caused by stress concentration phenomenon. From the local enlarged view in Fig. 9, it can be seen that the edge contact and stress concentration appeared during meshing and the maximum contact pressure mainly locates in the edge region of the toe. Once the contact pressure of one single gear is extracted along the contact line during the whole mesh cycle, the contact pattern and the distribution trends of the contact pressure can be obtained, as shown in Fig. 10 and 11. From the results, the contact pressure is decreased from the edge region to the middle part of the tooth surface" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001101_1369-4332.16.11.1871-Figure4-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001101_1369-4332.16.11.1871-Figure4-1.png", "caption": "Figure 4. Sample cut section for thickness variation measurement and for making material testing specimen", "texts": [ " This difference was kept ~8 mm for batch A5_B2 to facilitate smooth inversion and to avoid any frictional contact between top and middle cone. Few aluminum samples from each batch are shown in Figure 2, after spinning and ready for testing. The dimensions of tested samples are listed in Table 1 along with their results. The dimensions schematic is shown in Figure 3(a). The typical dimensions from each batch sample are shown in Figure 3(c). After spinning, the thickness may vary along with height in a sample. A typical cut section is shown in Figure 4, for sample A10_B1_S1. It was noticed that the cap end and bottom conical end (height 25 mm to 35 mm from bottom) had more thickness. Other portion had comparatively less thickness and its variation is also less. The averaged out thickness along height has been adopted for calculations. Similar thickness trend was noticed for other batches samples also. The corners of steps have minimal radius (~1 mm) and are modeled in FE (Finite Element) accordingly. The strips of 20 mm width were cut from samples at the mid location either from cut section (refer Figure 4) or from intact sample, for making material test samples. The compression is accomplished with INSTRON universal test machine of 4 ton capacity. The fixtures to position the samples are made of two steel thick platens (thickness = 20 mm) with arrangement to attach to INSTRON machine. Top fixture platen is attached to machine load cell. The ram on which bottom fixture platen is kept, moved up in speed of 5 mm/min to ensure the quasi static condition. The photograph along with samples and fixtures of test setup is shown in Figure 5" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001194_1.1716244-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001194_1.1716244-Figure2-1.png", "caption": "FIG. 2. AC4-~-------R~U~N-- I-ST_O_P ____ ~", "texts": [ " Although mechanical brakes and Geneva movements are excellent for this purpose, there are times when their cost or weight renders them undesirable, and magnetic braking is used. The conventional half-wave rec tifier braking circuit of Fig. 1 is probably the most econom- AC FIG. I. R ical of parts, but, because of the pulsating nature of the direct current, the braked motor tends to \"creep,\" and has very limited holding power. Also, a certain amount of 6O-cy noise is developed, which may be objectionable. The circuit of Fig. 2, using one additional rectifier, overcomes the above disadvantages. When the switch is in the \"stop\" position, alternate pulsations of line current flow through SRI and the motor windings, as in Fig. 1. When the line current reverses, however, the collapsing magnetic field of the motor maintains the current flow in the same direction through the windings and through SR2\u2022 Thus, a much steadier flow of direct current is main tained through the windings. Creepage is eliminated, hold ing power is greatly increased, and noise is minimized" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003294_978-3-642-39128-6-Figure2.8-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003294_978-3-642-39128-6-Figure2.8-1.png", "caption": "Fig. 2.8 The Cytocentering technology. a schematic drawing of the cytocentering site with suction and patch channels; b SEM image of two concentric opening formed with focused ion beam milling in a 10 \u03bcm thick quartz layer [19]", "texts": [ " The idea of bringing cells to the patching site rather than bringing the pipette to the cells has drawn the attention of many researchers and hundreds of different designs have been developed, each with some advantages over the others. In planar patch clamping the same hole is used for both cell positioning and gigaseal 2.3 Attempts to Improve Patch Clamping 10 2 Development of Patch Clamping formation. This greatly affects the gigaseal quality since debris in the solution may block or contaminate the pore. To overcome this problem, Stett et al. [13] developed a concentric double pipette-like structure (Fig. 2.8). The outer channel is used for cell positioning and the inner channel for current measurements. Positive pressure is initially applied in the inner channel to prevent debris from approaching its surface. Suction in the outer channel directs a cell to the top of the measurement site, which looks exactly like the tip of a pipette. When a cell is placed at the measurement site, suction is used in the inner channel to encourage seal formation. 11 Planar patch clamping systems increase throughput mainly by taking advantage of their potential to be parallelised and integrated with microfluidic systems" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002168_amr.328-330.755-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002168_amr.328-330.755-Figure2-1.png", "caption": "Fig. 2 Schematic diagram of test scheme", "texts": [ " To determine friction pair reasonable hardness match, the experiments investigate the resistance of both sides under the same conditions, and then compared with each other. This includes material matching, hardness matching, contact forms, specimens\u2019 forms, physical forms, speed, contact stress, lubrication condition, test environment. To study the effect of hardness on the fretting wear behavior of steel worm gear, and minimize the error caused by specimen shape, position, movement, measurement, etc. In this study, specially test method was designed as it is shown in Fig.2. To measure the amount of fretting wear of the worm material E match with worm wheel material A, the fretting wear test was made on the mach of E as the upper specimen and A as the lower specimen, then A as the upper specimen and E as the lower specimen. After the test, measure the fretting wear volume loss of the lower specimen A and lower specimen E as the fretting wear volume loss of the worm material E and worm wheel material A match. Worm material listed in Table 1 was matched with four kinds of hardness worm wheel material, and tested by the above method, then obtain the required test data" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000258_icma.2012.6285740-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000258_icma.2012.6285740-Figure1-1.png", "caption": "Fig. 1 Two degrees of freedom SJQ12 forearm EMG controlled prosthesis structure", "texts": [], "surrounding_texts": [ "OF ANALYSIS 2504978-1-4673-1278-3/12/$31.00 \u00a92012 IEEE Proceedings of 2012 IEEE International Conference on Mechatronics and Automation August 5 - 8, Chengdu, China prosthetic dimensional map A. Intelligent prosthetic works The composition of smart prostheses, in addition to the prosthetic body, following part of: 1) sensitive components, namely a variety of sensors. Their role is to changes in external conditions to convert the signal can be extracted, usually analog signal. 2) the information processing unit: it is usually a microcomputer. Its role is to read the signals of sensitive components, identification and decision-making, and issue control commands to the control elements. 3) control elements: generally installed inside the prosthetic body, used to adjust the parameters of the artificial limb movement, force parameters, structural parameters. forearm two degrees of freedom B. Intelligent prosthetics the main sources of vibration and noise analysis, mainly from the following reasons: 1)The motor itself vibration 2) the installation of the reducer shaft loose or strong enough 3) The structure unreasonable cause the resonance of the whole prosthetic 4) there are air holes." ] }, { "image_filename": "designv11_101_0000069_wcica.2012.6359142-Figure4-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000069_wcica.2012.6359142-Figure4-1.png", "caption": "Fig. 4 Communication link in scenarios B and C Fig.5 shows the trajectory coordination result in scenario B, where four UAVs have the same initial positions and target positions as scenario A but different flight paths. Four UAVs average collision avoidance decision-making times are 182.4ms, 177.9ms, 183.1ms and 179.8ms respectively. Fig.6 shows the flight paths and the relative distances in scenario C. We can find that the method failed in coordination, and UAV 1 collided with UAV 3 finally.", "texts": [ " In scenario A, four UAV initial positions are (20, 30), (30, 40), (40, 33), (40, 20), and their desired positions are (45, 30), (15, 10), (10, 20), (14, 45). The simulation experiment last 150s, and the collision avoidance trajectory coordination results are shown in Fig. 3. B. Partial Information Condition Simulation In partial information condition simulation, the communication network connectivity is local and incomplete. In order to illustrate the capability of our algorithm, we randomly choose two kinds of communication topologies as shown in Fig.4. In scenarios B and C, each UAV can only get its neighbor\u2019s state information. Noticed that our method have different performance in different network communication topologies. In scenario B, four UAVs are weakly connected that each pair of UAVs have one directional communication, thus, our method can support coordination. However, in scenario C, there are no communication between UAV 1 and UAV 3, UAV 2 and UAV 3, they can not exchange information with each other, and results show that UAV 1 collided with UAV 3" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000562_amm.201-202.517-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000562_amm.201-202.517-Figure1-1.png", "caption": "Fig. 1 Grinding disk definition and applied coordinate systems", "texts": [ " According to Hermann\u2019s proposition, the face gear is possible to be grinded by grinding disk [4]. It is possible to realize the face gear longitudinal crown by motion modification of the grinding disk center without reducing processing efficiency. The grinding disk manufactured the double crowned face gear and pinion in this paper and the method of TCA (tooth contact analysis) and LTCA (load tooth contact analysis) demonstrated the performance of the double crowned face gear. The face gear driver consists of pinion and face gear. Fig.1 represents the grinding disks definition by the pinion and the shaper that was applied by Litvin in face gear generation. Fig.1(a) are the rack cutters (i=p,s) generate the pinion (i=p) and the shaper (i=s) respectively, and the rack cutters traditional straight profile was substituted with a parabola one. Parameters aci is the parabola coefficient and u0i controls the point of tangency of parabolic profile to the traditional straight flank surface, uci and lci are the rack cutters surface parameters. The rack cutters surfaces position and normal vectors representing in coordinate system Sci by equation (1) and (2). Applying coordinate systems transformation and equations of meshing the surfaces of the pinion and shaper are described in Si Fig.1(b) and the details in literature [3]. T 2 c c c c c c 0 c( , ) 1i i i i i i i iu l a u u u l = \u2212 R (1) c c c c c c c c c i i i i i i i i iu l u l \u2202 \u2202 \u2202 \u2202 = \u00d7 \u00d7 \u2202 \u2202 \u2202 \u2202 R R R R n (2) All rights reserved. No part of contents of this paper may be reproduced or transmitted in any form or by any means without the written permission of Trans Tech Publications, www.ttp.net. (ID: 128.210.126.199, Purdue University Libraries, West Lafayette, USA-13/07/15,05:52:51) The grinding disks (i=p,s) whose teeth are surfaces of revolution are the tools applied to crown the pinion (i=p) and face gear (i=s) in longitudinal direction. The generating line of the grinding disk is the transversal section of the above-mentioned pinion and shaper, and the transversal sections revolving about xki in coordinate system Ski Fig.1(c) represents the grinding disk generation. Reminding that the profiles of the pinion and the shaper are out of involutes hence the grinding disk are capable of profile crowning of the pinion and the face gear. The pinion and the face gear are represented in coordinate system Stp and S2 respectively. Fig 2(a) shows that the grinding disks manufacturing the proposed longitudinal crown pinion (i=p) and face gear (i=s) move along the parabolic line DBE otherwise move along the straight line ABC for grinding the traditional non-longitudinal crowning pinion and face gear" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002016_55909-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002016_55909-Figure2-1.png", "caption": "Figure\u00a0 2\u00a0 shows\u00a0 the\u00a0 dynamics\u00a0 of\u00a0 the\u00a0 walking\u00a0 bipedal\u00a0 robot\u00a0model\u00a0 in\u00a0 the\u00a0 x \u2010direction,\u00a0with\u00a0 a\u00a0walking\u00a0motion\u00a0 that\u00a0 consists\u00a0 of\u00a0 two\u00a0 steps\u00a0 and\u00a0 two\u00a0 stance\u00a0 phases.\u00a0 The\u00a0 stance\u00a0phase\u00a0dynamics\u00a0are\u00a0calculated\u00a0by:\u00a0", "texts": [ "\u00a0 \u00a0 The\u00a0 state\u00a0 of\u00a0 the\u00a0 walking\u00a0 bipedal\u00a0 robot\u00a0 model\u00a0 can\u00a0 be\u00a0 described\u00a0 by\u00a0 the\u00a0 state\u00a0 variables\u00a0 of\u00a0 Tq (p,p) ,\u00a0 with\u00a0 Tp (x,y) \uff0cand\u00a0 the\u00a0 x \u00a0and\u00a0 y \u00a0 representing\u00a0 the\u00a0position\u00a0 of\u00a0 the\u00a0point\u00a0mass\u00a0with\u00a0respect\u00a0 to\u00a0a\u00a0 local\u00a0reference\u00a0 frame\u00a0 of\u00a0 x y z(e ,e ,e ) .\u00a0The\u00a0dynamics\u00a0 of\u00a0 the\u00a0model\u00a0 consists\u00a0 of\u00a0 a\u00a0 stance\u00a0phase\u00a0and\u00a0A\u00a0 transition\u00a0 from\u00a0one\u00a0 stance\u00a0phase\u00a0 to\u00a0 the\u00a0next.\u00a0\u00a0 \u00a0 2 0p p \u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0(1)\u00a0 \u00a0 (a)\u00a0 \u00a0 \u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0(b)\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0(c)\u00a0 Figure\u00a02.\u00a0Dynamics\u00a0of\u00a0the\u00a03D\u2010Model\u00a0in\u00a0the\u00a0x\u2010direction\u00a0\u00a0 Figure\u00a0 2(a)\u00a0 is\u00a0 shown\u00a0 in\u00a0 the\u00a0 state\u00a0 space\u00a0 and\u00a0 in\u00a0 the\u00a0 time\u00a0 domain\u00a0 in\u00a0 Figure\u00a0 2\u00a0 (b)\u00a0 and\u00a0 Figure\u00a0 2\u00a0 (c).\u00a0 For\u00a0 the\u00a0 state\u00a0 space\u00a0 plot,\u00a0 the\u00a0 boxed\u00a0 model\u00a0 pictures\u00a0 illustrate\u00a0 the\u00a0 model\u2019s\u00a0configuration\u00a0 for\u00a0 the\u00a0 four\u00a0quadrants\u00a0of\u00a0 the\u00a0plot.\u00a0 The\u00a0 continuous\u00a0 dynamics\u00a0 describe\u00a0 the\u00a0 evolution\u00a0 of\u00a0 the\u00a0 initial\u00a0state\u00a0 0,iq \u00a0to\u00a0state\u00a0 0,fq \u00a0over\u00a0time\u00a0 0t .\u00a0The\u00a0transition\u00a0 dynamics\u00a0describe\u00a0the\u00a0instantaneous\u00a0change\u00a0of\u00a0state\u00a0 1,fq \u00a0 to\u00a0state\u00a0 1,iq \u00a0due\u00a0to\u00a0step 0S .\u00a0This\u00a0sequence\u00a0is\u00a0repeated\u00a0for\u00a0 consecutive\u00a0 stance\u00a0phases\u00a0 and\u00a0 steps.\u00a0Any\u00a0 state\u00a0 that\u00a0 lies\u00a0 2 Int J Adv Robotic Sy, 2013, Vol", ",\u00a0the\u00a0distance\u00a0between\u00a0subsequent\u00a0point\u00a0foot\u00a0locations:\u00a0 \u00a0 n maxS l \u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0(5)\u00a0 Due\u00a0to\u00a0these\u00a0constraints,\u00a0the\u00a0model\u00a0can\u00a0only\u00a0operate\u00a0in\u00a0a\u00a0 subset\u00a0of\u00a0the\u00a0state\u00a0space.\u00a0The\u00a0subset\u00a0consists\u00a0of\u00a0all\u00a0states\u00a0 for\u00a0which\u00a0the\u00a0model\u00a0has\u00a0the\u00a0ability\u00a0to\u00a0come\u00a0to\u00a0a\u00a0stop.\u00a0For\u00a0 states\u00a0 outside\u00a0 of\u00a0 this\u00a0 subset,\u00a0 the\u00a0model\u00a0 will\u00a0 accelerate\u00a0 without\u00a0 the\u00a0 possibility\u00a0 of\u00a0 deceleration.\u00a0 The\u00a0 subset\u00a0 is\u00a0 spanned\u00a0 by\u00a0model\u00a0 states\u00a0 n,iq \u00a0 that\u00a0 lie\u00a0within\u00a0 the\u00a0 1\u2010step\u00a0 viable\u2010capture\u00a0basin,\u00a0as\u00a0given\u00a0by:\u00a0 \u00a0 0 min maxi s t 0 lp p d e 1 \u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0\u00a0(6)\u00a0 In\u00a0 Figure\u00a0 2,\u00a0 the\u00a0 basin\u00a0 boundary\u00a0 is\u00a0 indicated\u00a0 by\u00a0 the\u00a0 dashed\u00a0lines\u00a0labelled d .\u00a0 3.\u00a0Walking\u00a0Gait\u00a0 Our\u00a0proposed\u00a0foot\u00a0placement\u00a0algorithm\u00a0has\u00a0the\u00a0ability\u00a0to\u00a0 bring\u00a0 the\u00a0 model\u00a0 to\u00a0 any\u00a0 desired\u00a0 feasible\u00a0 state.\u00a0 In\u00a0 this\u00a0 paper,\u00a0 we\u00a0 select\u00a0 a\u00a0 more\u00a0 practical\u00a0 application\u00a0 of\u00a0 the\u00a0 controller\u00a0by\u00a0bringing\u00a0the\u00a0model\u00a0to\u00a0a\u00a0state\u00a0that\u00a0is\u00a0part\u00a0of\u00a0a\u00a0 desired\u00a0walking\u00a0gait.\u00a0This\u00a0requires\u00a0that\u00a0the\u00a0model\u00a0should\u00a0 not\u00a0only\u00a0arrive\u00a0at\u00a0the\u00a0desired\u00a0state,\u00a0but\u00a0that\u00a0it\u00a0should\u00a0also\u00a0 be\u00a0 capable\u00a0 of\u00a0 maintaining\u00a0 the\u00a0 desired\u00a0 gait\u00a0 cycle" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001869_amm.226-228.427-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001869_amm.226-228.427-Figure3-1.png", "caption": "Fig. 3 Assembly clearance", "texts": [ " Therefore, if the meshing noise of bevel gear is reduced, it must be increase the machining accuracy and reduce eccentricity. Improve the quality of shaft gear and assembly.[5] To reduce the assemble error which is produced from defective gear meshing, the assemble quality of bevel gears and shaft should be pay attention specially and ensure the center distance constant, so as to keep ideal contact area in a lower level noise in the transmission. The noise will be increased when the center distance changes larger than 0.01 mm. From Fig. 3 it can be seen that if the center distance of driving gear 1 decrease (pitch cone shift to the left) \u2206 1H , and for guarantee the specifies clearance of gear meshing in assembling; the driven gear 2 must be increased (pitch cone shift to the upper) \u2206 2H , 112 tan\u03b4HH \u2206=\u2206 . It follows that the assembly quality is also an efficient way to reduce noise of bevel gears noise. To reduces the noise of gears by noise mufflers. To shapes the gears by chamfer machines. Improve the natural frequency of the gear or improve gear structure" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003263_b978-0-08-097016-5.00007-3-Figure7.14-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003263_b978-0-08-097016-5.00007-3-Figure7.14-1.png", "caption": "FIGURE 7.14 Cross-section contour with elliptical shape.", "texts": [ " pNL8 were assessed through a manual fitting process: pNL1 \u00bc 2.5, pNL2 \u00bc 0.8, pNL3 \u00bc 0, pNL4 \u00bc 3, pNL5 \u00bc 1, pNL6 \u00bc 2, pNL7 \u00bc 2.5, pNL8 \u00bc 10. With these values, the responses computed for the nonlagging part of the side force show reasonable correspondence with the experimental results of Figure 7.13. See Exercise 7.1 for parameters of a motorcycle tire. The vertical load has been calculated using a tire model with a circular contour of the cross section with radius rc. For the more general case of an elliptic contour, cf. Figure 7.14, the following equations apply. We have, for the coordinates of the lowest point, z \u00bc b= ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 1\u00fe \u00f0a=b\u00de2tan2 g q h \u00bc a\u00f0a=b\u00detan g= ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 1\u00fe \u00f0a=b\u00de2tan2 g q (7.45) The vertical compression rz, which is the distance of the lowest point of the ellipse to the road surface if this distance is non-negative, now reads with ro the free tire radius and rl the loaded tire radius: rz \u00bc max\u00f0\u00f0ro rl b\u00fe z\u00decos g\u00fe h sin g; 0\u00de (7" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000285_s1068798x1306018x-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000285_s1068798x1306018x-Figure2-1.png", "caption": "Fig. 2. Planetary transmission.", "texts": [ " As a result, the frictional forces between the rim 1 and the intermediate rings 3 adjacent to the crowns 2 with the greatest load will be less than the azimuthal force, and those crowns will be displaced azimuthally, thereby redistributing the load between all the crowns. At the same time as the azimuthal slip, axial displace ment of the corresponding rings 3 is possible. In brief overloading, the composite gear acts as a safety clutch. Such correction of the tooth inclination is used in a multiple planetary transmission designed for borehole immersion [2]. In this transmission (Fig. 2), housing 1 contains a central shaft with drive gear 2 (common to all the satellites 4) and a shaft 3 to which a series of sat ellites 4 is hinged. The central gear with internal engagement consists of individual gear crowns 5, whose conical lateral surfaces are characterized by frictional interaction with the conical surfaces of intermediate frictional rings 6. Radial slots 7 in the ring permit frictional interaction of their external cylindrical surfaces with the internal cylindrical sur faces of the housing 1" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000762_1.3555500-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000762_1.3555500-Figure1-1.png", "caption": "Fig. 1. Vertical force frame.", "texts": [ " Due to safety considerations, we generally restrict masses to 10 kg or less. The frame also can be placed in a horizontal position and used like a conventional force table. As the orientation of the frame can be anything from horizontal to vertical, it can also be used to produce a three-dimensional system of forces. In addition, the frame provides a support system for many other physics experiments. The frame, the support base, and the upper and lower supports for attaching springs, weights, and pulleys are shown in detail in Fig. 1. The frame and base are made of standard size 1-in-x-4-in and 2-in-x-4-in lumber (preferably a fairly dense wood such as poplar), and all connections are made using deck screws. The frame is 5 ft high x 20 in wide. The upper and lower supports are each 5 ft long \u00bd-in diameter galvanized steel electric conduit, supported on the frame via pairs of screw eyes. With this conduit the effective space for experiments is almost 5 ft x 5 ft; this can be expanded to 5 ft x 10 ft using standard 10-ft conduit" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000312_amr.291-294.2970-Figure3-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000312_amr.291-294.2970-Figure3-1.png", "caption": "Figure 3. The three-dimensional entity model of casting model", "texts": [ " Next, this paper takes the forming of a high voltage switch casting model in the AFS-3000 rapid prototyping system as example to introduce the casting model rapid prototyping technology based on SLS. Three-dimensional entity model and Slice technology Before the rapid prototyping casting model by SLS method, the three-dimensional entity model of casting should be first generated in computer modeling system. The software for generating a three-dimensional entity model have AutoCAD, UG, Solidwork and Pro/E, etc. The three-dimensional entity model of a high voltage switch casting model in Pro/E is shown in Figure 3. After the generation of three-dimensional entity model of casting, it needs to translate solid model into STL file by computer interface module, as shown in Figure 4. STL file are universal standardized input format in rapid prototyping manufacturing system, which is mainly used to connect three-dimensional entity model and rapid prototyping manufacturing system. This file is composed of many data representing a series of small triangle which used to express the shapes and sizes of three-dimensional entity model" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0001744_iros.2011.6095082-Figure15-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0001744_iros.2011.6095082-Figure15-1.png", "caption": "Fig. 15 Calculation of the theoretical value of the steering radius", "texts": [ " In the case of the SCV I, it can be steered by using the traction belt which is in contact with the ground. However, a different point besides the traction belt is contacted in many cases, too. For examples, the rim of the end pulleys and the slack of the flexible crawler belt. They prevent the steering performance to be the same as the theoretical value due to the slippage of the belt. In this experiment, the steering radius was measured and compared with the theoretical value. The theoretical value is determined as follows by the use of Fig. 15. R is the steering radius. L1 is the length between the center of the end pulley and the axis of the steering arm. L2 is between the axis of steering arm and the center of the center body. \u0398 is the steering angle. R is obtained from the following calculating formula. R = L1cot\u0398 + L2 csc\u0398 In Fig. 16, the blue line is steering at forward, the red line is at reverse, and the green line is the theoretical value. From the Fig.17 Climbing the slope results, the measurement values are almost the same as the theoretical values" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002274_978-3-642-40840-3_19-Figure5-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002274_978-3-642-40840-3_19-Figure5-1.png", "caption": "Fig. 5. 3D Model of Quad-copter", "texts": [ " The platform independence feature can be used for different types of robotic platforms, such as mobile vehicles, flying quad-copters, etc. We have decided to use the final results from research simulation to test the virtual quad-copters. Virtual reality and 3D animation provides better representation of the final result of the simulations. This was our main reason to develop the 3D visualization tool. We have started with designing of 3D model of the quadcopter, which is converted to several 3D formats. The VRML model can be seen in Figure 5. After spinning the four screws of quad-copter we can start the 3D visualization of the final results from the simulations. Final data outputs from simulations are stored in the output storage elements. Data are input to the module and sorted. One module is responsible to create VRML-syntax from the sorted data and to compute and regulate the routes with respect on size of the 3D quad-copter models. The other modules are responsible for the background simulation. The background is shown as a 3D model of the objective environment in which the quad-copters can fly in [3]" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000252_memsys.2012.6170186-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000252_memsys.2012.6170186-Figure1-1.png", "caption": "Figure 1: Micro-object manipulation by AC-EWOD-driven twin bubbles: (a) Initial state; (b) When twin bubbles actuated by AC-EWOD at a certain frequency (100 Hz) generate streaming, a micro-object is pushed by the streaming and transported to the next EWOD electrodes; (c) The object is carried by the sequential AC-EWOD operations with the twin bubbles to the right end of EWOD electrodes; (d) When the twin bubbles actuated by AC-EWOD at 1 Hz are transported to the original position without streaming, the object is released from the bubbles and remains where it is. Note that the bubble-induced streaming is generated only at a certain frequency.", "texts": [ " To minimize the damage of manipulating objects, the effects of the shear stresses from the microstreaming were reduced by controlling the intensity of the applied acoustic signals and the operation distances between the bubbles and objects using a three-dimensional (3-D) traverse system combined with the U-shape rod. In this configuration, however, two different actuation systems are required: (1) a piezoactuator for bubble oscillation and (2) a 3-D traverse system for bubble transportation, which makes the total system bulky and integration with micro total analysis systems (\u03bc\u03a4\u0391S) difficult. In contrast to the previous work, we propose a novel approach that AC-EWOD is solely used for both bubble oscillation and transportation to manipulate micro/bio-objects in a microfluidic chip, as shown in Fig. 1. The proposed method is simple, yet efficient, and allows non-contact manipulation between the bubbles and objects with minimizing physical damage. Electrowetting-on-dielectric (EWOD) is one of the methods of manipulating the interfacial tension between two fluids using an external electrical potential[12-15]. In the early 1990s, Berge et al.[16, 17] deposited a thin dielectric material on top of a metal electrode to prevent direct contact between the electrode and a conducting droplet. Hence, when an electrical potential was applied between the electrode and the droplet, the dielectric layer 978-1-4673-0325-5/12/$31" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0003060_ijnsns-2012-0172-Figure9-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0003060_ijnsns-2012-0172-Figure9-1.png", "caption": "Fig. 9: Effect of parametric excitations on the frequency-response curve", "texts": [ " Instead of using the frequency of the parametric excitation we as a parameter, we introduce a detuning parameter s = o(1), which quantitatively describes the nearness of we to w0. Accordingly, we write 0ew w s\u2248 + (23) Substituting Eq. (23) into Eq. (4), the frequencyresponse curves are obtained from HAM. Let us consider the case of w0 = 0.2, a3 = 0.1, B2 = B3 = 0, we \u2248 w0 + s, Fah = 0, weh = w0, FaT = 0, a = 0.1, m = 0.02. Fig. 8 shows that the effect of the nonlinearity of the frequency-response curve is to bend the amplitude curve under different values of static loads in the case of Fm = 0.3. Similarly, Fig. 9 shows that the multivalued regions of the\u00a0frequency-response curve are formed under different parametric excitations in the case of B1 = 0.2. From Fig. 8 and Fig. 9 we can draw the following conclusions: 1. The multivaluedness of the frequency-response curves due to the strong nonlinearity has a significance from the physical point of view because it lead to jump phenomena. 2. The frequency-response curves are divided into upper and lower branches respectively under the static loads. The resonance amplitude a has obviously increased with the increase of Fm. 3. The frequency-response curves are also divided into upper and lower branches respectively under the parametric excitations" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000562_amm.201-202.517-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000562_amm.201-202.517-Figure2-1.png", "caption": "Fig. 2 Coordinates systems for derivation of pinion and face gear", "texts": [ "199, Purdue University Libraries, West Lafayette, USA-13/07/15,05:52:51) The grinding disks (i=p,s) whose teeth are surfaces of revolution are the tools applied to crown the pinion (i=p) and face gear (i=s) in longitudinal direction. The generating line of the grinding disk is the transversal section of the above-mentioned pinion and shaper, and the transversal sections revolving about xki in coordinate system Ski Fig.1(c) represents the grinding disk generation. Reminding that the profiles of the pinion and the shaper are out of involutes hence the grinding disk are capable of profile crowning of the pinion and the face gear. The pinion and the face gear are represented in coordinate system Stp and S2 respectively. Fig 2(a) shows that the grinding disks manufacturing the proposed longitudinal crown pinion (i=p) and face gear (i=s) move along the parabolic line DBE otherwise move along the straight line ABC for grinding the traditional non-longitudinal crowning pinion and face gear. Applying grinding disk coordinates transformation from system Sgp to system Stp and meshing equation between the grinding disk and the pinion defines the surface of the double-crowned pinion. The envelope to the family of grinding disk and the meshing equations between grinding disk and the machined face gear determine the surface of face gear, the family represented in S2 by transforming grinding disk coordinates in Sts to S2 Fig.2(b), and the meshing equations are the dot products of grinding disk normal vector and relative velocity between the points of grinding disk surface tangency to the face gear surface. TCA that provides under-load contact information must be simulated prior to LTCA. Fig.3 (a) and (b) respectively represent installation process of the pinion and face gear in coordinate system Sf for TCA simulation. Where \u03b3f=\u03c0-\u03b3m+\u2206\u03b3 and parameters \u2206\u03b3, \u2206q, and \u2206E are errors of alignment. The surfaces of the pinion and face gear will be in tangency contact in Sf if the following vectors equations are observed: 2f cs gs g g 2 1f cp gp 1( , , , , ) ( , , )s su L u L\u03c8 \u03d5 \u03a6 \u03a6=R R (3) 2f cs gs gs gs 2 1f cp gp 1( , , , , ) ( , , )u L u L\u03c8 \u03d5 \u03a6 \u03a6=n n (4) g2(1) 1 cs gs gs gs gs( , , ) 0f u L v\u03c8 = \u22c5 =n (5) g2(2) 2 cs gs gs gs gs gs( , , , ) 0f u L v\u03c8 \u03d5 = \u22c5 =n (6) Where R2f, n2f, R1f, and n1f are respective position and normal vectors of the pinion and face gear in Sf, f1 and f2 are respective meshing equations in the grinding disk feeding and face gear generating motion direction, \u03c6gs=\u03c6s+\u2206\u03c6s, \u03c62+\u2206\u03c62=m2s\u22c5\u03c6gs, m2s is the contact ratio" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000663_icssem.2012.6340798-Figure2-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000663_icssem.2012.6340798-Figure2-1.png", "caption": "Figure 2 the structure of the steering system", "texts": [ " INTRODUCTION TO ORIGINAL STEERING SYSTEM STRUCTURE The original steering system structure is composed mainly of the steering wheel, the link system, the drive system, and other components. Two steerable wheels respectively mounted on the girders of the vehicle steering sleeve, two steering wheel connected by a crossbar, Therefore it can make the steering operation coordinated, steering hydraulic cylinder one end is mounted on the frame and the other end is mounted on the crossbar. As shown in Figure 2. Original steering system contains hydraulic power unit, hydraulic steering cylinder, four-way three-position solenoid valves, flow priority valve, relief valve and other major components. As shown in Figure 3. Its working principle is: Workers on work platform open the main power, operating steering button, Electronic Control Unit through the can bus received the steering signal, then putout two signals: one is pump motor control signal the other is solenoid control signal. The pump motor controller receives the control signal to the pump motor, control the Compound excited motor rotation" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0000026_amr.430-432.1524-Figure8-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0000026_amr.430-432.1524-Figure8-1.png", "caption": "Fig. 8 The third order modes and vibration displacement without defining contact", "texts": [ " Division grid control relevance is 0, element size is system default and control smooth is medium. Node number is 258165, element number is 156450. Meshing schemes can be shown in Fig. 3. Using cylindrical support constraints fixed two gear within hole node all freedom, due to conduct modal analysis, the natural frequency and mode has nothing to do with the load, and negligible damping [5]. Influence of system frequency is generally low frequency [5], so we can extract the 3 order mode for analysis. There are two cases. (2) Not defining contact. The analysis can be shown in Fig. 6-Fig. 8. By comparing the defining contact and without defining contact, it can be found the natural frequency in the defined contact case is higher than that of without defining contact. Through observing both former three order modes figure, it can be found that big gear vibration is larger than the small gear. Defining contact between the two meshing gears is equivalent to add a mesh stiffness. There have an impact on the two gears axial runout and torsional vibration, especially on the torsional vibration, and the modes may also be affected at the same time" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002030_ijhvs.2013.053008-Figure23-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002030_ijhvs.2013.053008-Figure23-1.png", "caption": "Figure 23 Scheme for correction of contact force calculation", "texts": [ " It consists in correcting the creep forces calculated by the standard subroutines of VI-Rail according to the SHE model (Curve 1, Figure 7) in order to take into account the effects of temperature variation in the contact area. A User Defined Function (UDF) has been implemented at the purpose. As it will be explained in detail later on, within this subroutine, temperature in the contact area is determined as a function of the creep forces (Fx,tot, Fy,tot) and the sliding velocities (equations (8)\u2013(9)). Then the friction coefficient is corrected accounting for the calculated contact area temperature (equation (10)). An equivalent system of forces (Figure 23) is finally introduced to correct the creep forces calculated from the SHE model (Fx, Fy) so that the effect of a friction coefficient decreasing with temperature (equation (11)) is included. Note that, since temperature in the contact area (and thus the friction coefficient) is a function of the creep forces, a feedback is introduced in the contact forces calculation algorithm (Figure 22). In the following, the details of the implemented procedure for creep forces calculation are provided. As already mentioned, the first step of the procedure consists in estimating the temperature in the contact area due to the friction dissipated power P", "4 has been considered for the friction coefficient at ambient temperature \u00b50 (dry rail conditions), while coefficient K has been assumed equal to 4.417 \u00d7 10\u20134 \u00b0C\u20131 in accordance with (Ertz and Bucher, 2002). An equivalent system of forces has then been introduced to correct the creep forces calculated by the standard subroutines of VI-Rail for the SHE model and to take the effect of a friction coefficient varying with temperature in the contact area (equation (10)) into account. It is in fact to point out that the subroutines implemented in VI-Rail to calculate the contact forces cannot be directly modified by users. Making reference to Figure 23, the equivalent system of forces introduced to correct the creep forces is given by: Fx ,therm = K \u2206T Fx Cx ,therm = K \u2206T Fy R Fx ,tot = Fx \u2212 Fx ,therm Fy ,therm = K \u2206T Fy Cy ,therm = K \u2206T Fx R Fy ,tot = Fy \u2212 Fy ,therm (11) Fx,therm, Fy,therm, Cx,therm, and Cy,therm represent the variation of the creep force/torque components due to the decrease of the friction coefficient with increasing temperature in the contact area. As already mentioned, note that through equations (8)\u2013(11) the creep forces calculated according to the proposed procedure (Fx,tot, Fy,tot) are a function (by means of a temperature dependent friction coefficient) of the sliding velocities and of the creep forces themselves" ], "surrounding_texts": [] }, { "image_filename": "designv11_101_0002052_emeit.2011.6024109-Figure1-1.png", "original_path": "designv11-101/openalex_figure/designv11_101_0002052_emeit.2011.6024109-Figure1-1.png", "caption": "Figure 1 The tooth surface relative sliding diagram Two gears meshing can be treated as pitch circles pure rolling. In the jerkwater time, the pitch circles travel equals to arc length.", "texts": [ " The main factors, relative sliding and contact compressive stress, which influence the wear instance of the tooth profile is progressed in this paper, and the equation are established. Simulation is progressed with numerical method, the changing laws of relative sliding coefficient and contact compressive stress are found based on the simulation curve. The wear instance of the tooth profile is analyzed on the base in order to get the means to improve the wear instance of the tooth profile. II. THEORETICAL ANALYSIS When two pitch wheels turning, the relative sliding velocities are different at the different position of the tooth profile. Figure 1 is the tooth surface relative sliding diagram of the two pitch wheels, P is the instantaneous centre of velocity, N\\N2 is the theoretic meshing line, K is the meshing point. du = PPX = PP2 But the lengths of the tooth profiles travelling are unequal. dsx = KKX ds2 = KK2 So the relative sliding length is dsx \u2014 ds2 . The relative sliding coefficient is induced as this: 978-1-61284- -8/ll/$26.00 \u00a92011 IEEE 4788 12-14 August, 2011 7 = (1) Where, / is the length of PK, the length of NXP is p H - / .The right-angle triangle N\\PQ and PP\\Q are used to carry numeration through, the length of PQ can be gotten: PQ = PNX sin d