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机械毕业设计英文翻译
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机械毕业设计38英文翻译外文文献翻译132,机械毕业设计英文翻译
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附录 II：外文文献翻译原文及其译文 Effects of structure elastic deformations of wheelset and track on creep forces of wheel/rail in rolling contact Abstract In this paper the mechanism of effects of structure elastic deformations of bodies in rolling contact on rolling contact performance is briefly analyzed. Effects of structure deformations of wheelset and track on the creep forces of wheel and rail are investigated in detail. General structure elastic deformations of wheelset and track are previously analyzed with finite element method, and the relations, which express the structure elastic deformations and the corresponding loads in the rolling direction and the lateral direction of wheelset, respectively, are obtained. Using the relations, we calculate the influence coefficients of tangent contact of wheel and rail. The influence coefficients stand for the occurring of the structure elastic deformations due to the traction of unit density on a small rectangular area in thecontact area of wheel/rail. They are used to revise some of the influence coefficients obtained with the formula of Bossinesq and Cerruti in Kalkers theory of three-dimensional elastic bodies in rolling contact with non-Hertzian form. In the analysis of the creep forces, the modified theory of Kalker is employed. The numerical results obtained show a great influence exerted by structure elastic deformations of wheelset and track upon the creep forces. 2002 Elsevier Science B.V. All rights reserved. Keywords: Wheel/rail; Rolling contact; Creep force; Structure elastic deformation nts Introduction During running of a train on track the fierce action between wheelset and rails causes large elastic deformations of structure of wheelset and track. The large structure deformations greatly affect performances of wheels and rails in rolling contact, such as creep forces, corrugation 13, adhesion, rolling contact fatigue, noise 4,5 and derailment 6. So far rolling contact theories widely used in the analysis of creep forces of wheel/rail are based on an assumption of elastic half space 712. In other words, the relations between the elastic deformations and the traction in a contact patch of wheel/rail can be expressed with the formula of Bossinesq and Cerruti in the theories. In practice, when a wheelset is moving on track, the elastic deformations in the contact patch are larger than those calculated with the present theories of rolling contact. It is because the flexibility of wheelset/rail is much larger than that of elastic half space. Structure elastic deformations (SED) of wheelset/rail caused by the corresponding loads are shown in Figs. 1 and 2. The bending deformation of wheelset shown in Fig. 1a is mainly caused by vertical dynamic loads of vehicle and wheelset/rail. The torsional deformation of wheelset described in Fig. 1b is produced due to the action of longitudinal creep forces between wheels and rails. The oblique bending deformation of wheelset shown in Fig. 1c and the turnover deformation of rail shown in Fig. 2 are mainly caused by lateral dynamic loads of vehicle and wheelset/rail. The torsional deformations with the same direction of rotation around the axle of wheelset (see Fig. 1d), available for locomotive, are mainly caused by traction on the contact patch of wheel/rail and driving torque of motor. Up to now very few published papers have discussions on the effects of the SED on creepages and creep forces between wheelset and track in rolling contact. In fact, the SED of wheelset/rail mentioned above runs low the normal and tangential contact stiffness of wheel/rail. The normal contact stiffness of wheel/rail is mainly lowed by the subsidence of track. The normal contact stiffness lowed doesnt affect the normal pressure on the contact area much. The lowed tangential contact stiffness affects the status of stick/slip areas and the traction in the contact area greatly. If the effects of the SED on the rolling contact are taken into account in analysis of rolling contact of wheel/rail, the total slip of a pair of contacting particles in a contact area is different from that calculated with the present rolling contact ntstheories. The total slip of all the contacting particles and the friction work are smaller than those obtained under condition that the SED is ignored in the analysis of creep forces of wheel/rail. Also the ratio of stick/slip areas in a contact area is larger than that without consideration of the effects of the SED. In this paper the mechanism of effects of structure elastic deformations of bodies in rolling contact on rolling contact performance is briefly analyzed, and Kalkers theoretical model of three-dimensional elastic bodies in rolling contact with non-Hertzian form is employed to analyze the creep forces between wheelset and track. In the numerical analysis the selected wheelset and rail are, respectively, a freight-car wheelset of conical profile, China “TB”, and steel rail of 60 kg/m. Finite element method is used to determine the SED of them. According to the relations of the SED and the corresponding loads obtained with FEM, the influence coefficients expressing elastic displacements of the wheelset and rail produced by unit density traction acting on the contact area of wheel/rail are determined. The influence coefficients are used to replace some of the influence coeffi- cients calculated with the formula of Bossinesq and Cerruti in Kalkers theory. The effect of the bending deformation of wheelset shown in Fig. 1a and the crossed influences among the structure elastic deformations of wheelset and rail are neglected in the study. The numerical results obtained show marked differences between the creep forces of wheelset/rail under two kinds of the conditions that effects of the SED are taken into consideration and neglected. 2. Mechanism of reduced contact stiffness increasing the stick/slip ratio of contact area In order to make better understanding of effects of the SED of wheelset/track on rolling contact of wheel/rail it is necessary that we briefly explain the mechanism of reduced contact stiffness increasing the ratio of stick/slip area in a contact area under the condition of unsaturated creep-force. Generally the total slip between a pair of contact particles in a contact area contains the rigid slip, the local elastic deformation in a contact area and the SED. Fig. 3a describes the status of a pair of the contact particles, A1 and A2, of rolling contact bodies and without elastic deformation. The lines, A1A_1 and A2A_2 in Fig. 3a, are marked in order to make a good understanding of the description. After the deformations of the bodies take place, the positions and deformations of lines, A1A_1 and A2A_2, are shown in Fig. 3b. The displacement difference, w1, between the two dash lines in Fig. 3b is caused by the ntsrigid motions of the bodies and (rolling or shift). The local elastic deformations of points, A1 and A2, are indicated by u11 and u21, which are determined with some of the present theories of rolling contact based on the assumption of elastic-half space, they make the difference of elastic displacement between point A1 and point A2, u1 = u11 u21. If the effects of structure elastic deformations of bodies and are neglected the total slip between points, A1 and A2, can read as: S1 = w1 u1 = w1 (u11 u21) (1) The structure elastic deformations of bodies and are mainly caused by traction, p and p_ acting on the contact patch and the other boundary conditions of bodies and , they make lines, A1A_1 and A2A_2 generate rigid motions independent of the local coordinates (ox1x3, see Fig. 3a) in the contact area. The u10 and u20 are used to express the displacements of point A1 and point A2, respectively, due to the structure elastic deformations. At any loading step they can be treated as constants with respect to the local coordinates for prescribed boundary conditions and geometry of bodies and . The displacement difference between point A1 and point A2, due to u10 and u20, should be u0 = u10 u20. So under the condition of considering the structural elastic deformations of bodies and , the total slip between points, A1 and A2, can be written as: S1 = w1 u1 u0 (2) It is obvious that S1 and S1 are different. The traction (or creep-force) between a pair of contact particles depends on S1 (or S1 ) greatly. When |S1| 0 (or |S1 | 0) the pair of contact particles is in slip and the traction gets into saturation. In the situation, according to Coulombs friction law the tractions of the above two conditions are same if the same frictional coefficients and the normal pressures are assumed. So the contribution of the traction to u1 is also same under the two conditions. If |S1| = |S1 | 0, |w1| in (2) has to be larger than that in (1). Namely the pairs of contact particles without the effect of u0 get into the slip situation faster than that with the effect of u0. Correspondingly the whole contact area without the effect of u0 gets into the slip situation fast than that with the effect of u0. Therefore, the ratios of stick/slip areas and the total traction on contact areas for two kinds of the conditions discussed above are different, they are simply described with Fig. 4a and b. Fig. 4a shows the situation of stick/slip areas. Sign in Fig. 4a indicates the case without considering the effect of u0 and indicates that with the effect of u0. Fig. 4b expresses a relationship law between the total tangent traction F1 of a contact area and the creepage w1 of the bodies. Signs and in Fig. 4b have the same meaning as those in Fig. 4a. From Fig. 4b it is known that the tangent traction F1 reaches its ntsmaximum F1max at w1 = w_1 without considering the effect of u0 and F1 reaches its maximum F1max at w1 = w_1 with considering the effect of u0, and w_1 0 when the wheelset shifts towards the left side of track and 0 if it is inclined, in the clockwise direction, between the axis of wheelset and the lateral direction of track pointing to the left side. The parameters depend on the profiles of wheel and rail, y and . But if profiles of wheel and rail are prescribed they mainly depend on y 7. Detailed discussion on the numerical method is given in 7,8 and results of contact geometry of wheel/rail. When a wheelset is moving on a tangent track the rigid creepages of wheelset and rails read as where i = 1, 2, it has the same meaning as subscript i in (3). The undefined parameters in (4) can be seen in the Nomenclature. It is obvious that the creepages depend on not only the parameters of contact geometry, but also the status of wheelset motion. Since the variation of the parameters of contact geometry depend mainly on y with prescribed profiles of wheel/rail some of their derivatives with respect to time can be written as Putting (5) into (4), we obtain: In the calculation of contact geometry and creepage of wheel/rail, the large ranges of the yaw angle and lateral displacement of wheelset are selected in order to make the creepage and contact angle of wheel/rail obtained include the situations producing in the field as completely as possible. So we select y = 0, 1, 2, 3, . . . , 10 mm, = 0.0, 0.1, 0.2, 0.3, . . . , 1.0, y/v = 0, 0.005 and r0 /v = 0, 0.001. riy, /y and i/y are calculated with center difference method and the numerical results of ri , and i versus y. l0 = 746.5mm, r0 = 420mm.Using the ranges of y, , y/v and r0 /v selected above we obtain that i 1 ranges from 0.0034 to 0.0034, i 2 ranges from 0.03 to 0.03, i 3 ranges from 0.00013 to 0.00013 (mm1), and contact angle i is from to 2.88 to 55.83. Due to length limitation of paper the detailed numerical results of creepage and contact geometry are not shown in this paper. 4. Conclusion (1) The mechanism of effects of structure elastic deformation of the bodies in rolling contact on rolling contact performance is briefly analyzed. It is understood that ntsthe reduced contact stiffness of contacting bodies increases the stick/slip area of a contact area under the condition that the contact area is not in full-slip situation. (2) Kalkers theoretical model of three-dimensional elastic bodies in rolling contact with non-Hertzian form is employed to analyze the creep forces between wheelset and track. In the analysis, finite element method is used to determine the influence coefficients expressing elastic displacements of wheelset/rail produced by unit traction acting on each rectangular element, which are used to replace some of the influence coefficients calculated with the formula of Bossinesq and Cerruti in Kalkers theory. The numerical results obtained show the differences of the creep forces of wheelset/rail under two kinds of conditions that effects of structure elastic deformations of wheelset/rail are taken into consideration and neglected. (3) The structure elastic deformations of wheelset and track run low the contact stiffness of wheelset and track, and reduce the creep forces between wheelset and track remarkably under the conditions of unsaturated creep force. Therefore, the situation is advantageous to the reduction of the wear, rolling contact fatigue of wheel and rail. (4) In the study the effect of the bending deformation of wheelset shown in Fig. 1a is neglected, and the crossed influence coefficients AIiJj(i _= j ; i, j = 1, 2) are not revised. So, the accuracy of the numerical results obtained is lowed. In addition, when the lateral displacement of center of the wheelset, y 10mm, the flange action takes place. In such situation the contact angle is very large and the component of the normal load in the lateral direction is very large. The large lateral force causes track and wheelset to produce large structure deformations, which affect the parameters of contact geometry of wheel/rail and the rigid creepages. Therefore, the rigid creepages, the creep forces, the parameters of contact geometry, the SED and the motion of wheelset have a great influence upon each other. It is necessary that they are synthetically put into consideration in the analysis. Numerical results of them can be obtained with an alternative iterative method. Probably conformal contact or two-point contact between wheel and rail take place during the action of flange. Such phenomenon of wheelset and rails in rolling contact is very complicated, and can be analyzed with a new theory of rolling contact, which may be a FEM model including effects of structure deformations and all boundary conditions of wheelset and track in the near future. nts nts轮和轨道的结构弹性变形对滚动接触的轮 /轨蠕变力的影响 摘要 本文简要分析了机构的结构弹性变形对滚动接触时滚动接触性能的影响。详细研究了轮和轨道结构变形对轮轨滚动接触时的蠕变力的影响。对轮和轨道的一般性结构弹性变形进行了有限元分析，以 及分别获得了表示结构弹性变形和相应的滚动方向负荷和横向方向轮的关系。利用这些关系，我们计算了轮轨切线接触的影响系数。这些影响系数说明结构发生弹性变形与轮 /轨接触面上一个小矩形面积内的单位密度牵引力有关。它们被用来修整一些由在 Kalker 以非赫兹形式的三维弹性体滚动接触理论中提出的 Bossinesq 和 Cerruti 公式得出的影响系数。在分析爬行力时就应用了修正后的 Kalker 理论。获得的数值结果表明轮和轨道的结构性弹性变形对蠕变力存在很大的影响。 2002 爱思唯尔科技有限公司保留所有权利。 nts关键词：轮 /轨 ;滚动接触 ;蠕变力 ;结构弹性变形 1.导言 在轨道上运行的火车轮和铁轨之间的激烈行动引起轮和轨道的结构出现大量弹性变形。大量结构变形将大大影响车轮和钢轨的滚动接触性能，如蠕变力，起皱 1-3 ，粘附，滚动接触疲劳，噪音 4,5 和脱轨 6 。到目前为止，广泛应用于分析轮 /轨蠕变力的滚动接触理论基于假设的弹性半空 7-12 。换言之，轮 /轨弹性变形和牵引点的关系可用该理论的 Bossinesq 和切瑞蒂公式表示。在实践中，当轮正在轨道上运动时，接触处的弹性变形大于 按现有的滚动接触理论所计算出的值。这是因为轮 /轨的弹性远大于半弹性空间。相应的负载造成轮 /轨的结构弹性变形（ SED）于图 1 和 2 所示 。在图 1A 中显示的轮辐的弯曲变形，主要是由车辆和轮对 /轨道的纵向动态载荷引起的。图。图 1b 中所描述的轮辐扭变形是由车轮和钢轨之间纵向蠕变力作用产生的。引起图 1C 所示的轮辐斜弯曲变形和图 2 所示铁路的倾覆变形的主要原因是辆和轮对轨道的横向动荷载。可用于机车运动的与旋轴轮转向同一方向的扭变形（见图。 1 ），主要是由轮 /轨接触处的牵引力和电机驱动力矩引起的。直至目前为止很少有发表论文 讨论 SED 对轮和轨道之间的滚动接触的蠕动和蠕变力的影响。 事实上，上面提到的轮 /轨 SED 降低了轮 /轨的法向和切向接触刚度。轮 /轨的法向的接触刚度，主要是因轨道下沉而减小。法向的接触刚度降低并不会影响接触面的法向压力很大。该切线接触刚度降低对粘附 /滑移区的境况和接触面的牵引力的影响很大。如果考虑到滚动接触中对轮 /轨的滚动接触分析，接触面一对接触粒子的总滑动系数与按本滚动接触理论计算的是不同的。取得的所有接触粒子的总滑动系数和摩擦功，小于在忽略 SED 的影响条件下分析轮 /轨蠕变力时所得值。接触面粘 /滑区的比例也大 于不考虑 SED 的影响时的。本文简要分析了机构的结构弹性变形对滚动接触时滚动接触性能的影响，并在分析轮和轨道蠕变力时就应用了 Kalker 的非赫兹形式三维弹性机构滚动接触理论模型。在分析时选定的轮和铁路数值分别是，一列货运汽车的锥形剖面轮，中国“ TB” ，和 60公斤 /米的钢轨。有限元方法是用来确定他们的 SED 。根据 SED 和通过有限元获得的相应的载荷的关系，确定能表示由接触面单位密度牵引力产生的轮轨弹性位移的影响系数。这些影响系数是用来取代一些由 Kalker 的理论中的 Bossinesq 和切瑞蒂公式计算出的影响系数。轮弯曲变形的影响如图 1A 示，轮和铁路的结构弹性变形的交叉影响研究时被忽视。数值结果表明，在 SED 的影响是否被考虑的两种情况下，轮 /轨的蠕变力有明显区别。 nts2.减少接触刚度增加接触面粘 /滑率的机械装置 为了更好地了解轮 /轨滚动接触的轮 /轨 SED 的影响，我们有必要简要地了解不饱和蠕变力条件下减少接触刚度增加接触面粘 /滑率的机械装置。一般来说，接触面的一对接触粒子之间的总滑动，包含刚性滑移，接触面接触处的弹性变形和 SED。图 3A 描述接触对粒子的情形， A1 和 A2，滚动接触体且没有弹性 变形。线 A1-A1 和 A2-A2 标记于图 3A 中，以便更好的理解说明。机构发生变形后的位置和变形线， A1-A1 和 A2-A2，列于图 3A 中。位移差异， W1，图 3B 中两个破折号之间的线是由机构的硬性的运动和滚动或滑动所造成的 。该处的弹性变形点， A1 和 A2，是靠 u11 和 u21 表示的，这是由一些依据弹性半空间假设的滚动接触理论确定的，他们导致了点 A1 和点 A2 的弹性位移之间的差异 ， U1= u11 - u21。如果机构的结构弹性变形的影响被忽视，总滑点之间， A1 和 A2 ，可以理解为： S1= w1u1=w1(u11 u21)（ 1）。机构的结构弹性变形的主要由牵引力所造成的， p 和 p_作用于接触点和机构的其他边界条件，它们导致线， A1_A1和 A2_A2 产生不受接触面的坐标（ ox1x3，见图 3A）约束的刚性运动。 u10 和 u20是用来分别表示点 A1 和点 A2 由于结构弹性变形的位移。在任何载荷下，他们可以视为与该处给定边界条件下的坐标和机构的几何形状保持一致。点 A1 和点A2 位移差异，取决于 u10 和 u20，应为 u0 = u10 - u20 。这样的条件下，考虑机构的结构弹性变形，总滑点之间， A1 和 A2 ，可以写成： S1= w1u1u0（ 2）。很明显 S1 和 S*1 是不同的。接触粒子对之间的牵引力（或蠕变力），极大地取决于 S1（或 S * 1 ）。当 |S1| 0 (or |S1 | 0)接触粒子对是打滑且牵引进入饱和。在这种情况下，根据库仑摩擦定律，如果摩擦系数与假设的法向压力相同，上述两个条件下牵引力相同。这样牵引力对 U1 的作用在上述两个条件下也是相同的。如果 |S1| = |S1 | 0, |w1| 在（ 2）中要大于（ 1）中。即接触粒子对在没有 u0的影响时进入滑动形势快于有 u0 的影响时。相应的整个接触面在没有 u0 的影响时进入滑动形势快于有 u0 的影响时。因此，粘 /滑区比率和接触处的总牵引力在上述两种条件下是不同的，在图 4a 和 b 对他们进行了简单的描述。 4A 表明了粘 /滑区的情况。图 4A 中的标志表明了考虑与不考虑 u0 的影响的情形。图 4B表示接触面的总切线牵引 F1 积和 1 机构的蠕动 W 之间关系。图 4A 中的标志和图 4B 中的具有相同的含义。从图 4b 可知，切线牵引力 F1 达到最大值 F1max 在W1= w_1 而不考虑 u0 作用时和 F1 达到最大值 F1max 在 W1= w_1 考虑 u0 的影响，并 w_1 w_ 1 。 u0 主要取决于机构的 SED 和接触面的牵引力。大的 SED导致大的 u0 和
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