5外文翻译汽车噪声的控制策略研究

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1、毕业设计(论文) 外文翻译 英文翻译题目一： New analytical method to evaluate the powerplant and chassis coupling in the improvement vehicle NVH英文翻译题目二： 学 院 名 称： 机械工程 专 业： 汽车服务工程 班 级： 汽车 071 姓 名： 陈晔 学 号 07405050104 指 导 教 师： 李发宗 定稿日期： 2011 年2月18日英文题目一New analytical method to evaluate the powerplant and chassis coupling i

2、n the improvement vehicle NVH翻译内容第929-934页指导教师评语外文翻译与毕业选题相关，译文质量较好，字数达标，符合本科毕业论文外文翻译要求。指导教师签字年 月 日英文题目二翻译内容指导教师评语指导教师签字 年 月 日 New analytical method to evaluate the powerplant and chassis coupling in the improvement vehicle NVHE. Courteille a,b, L. Lotoing b, F. Mortier a, E. Ragneau bAbstractThe des

3、ign of an automotive powerplant mounting system is an essential part in vehicle safety and improving the vehicle noise, vibration and harshness (NVH) characteristics. One of the main problems encountered in the automotive design is isolating low frequency vibrations of the powerplant from the rest o

4、f the vehicle. The significant powerplant mass makes the choice of frequency and mode arrangements a critical design decision. Several powerplant mounting schemes have been developed to improve NVH properties concentrating on the positioning and design of resilient supports. However these methods ar

5、e based on decoupling rigid body modes from a grounded powerplant model which ignores chassis and suspension system interactions.But it cannot be stated that decoupling the grounded rigid body modes of the powerplant will systematically reduce chassis vibrations. In this paper, a new analytical meth

6、od is proposed to examine the mechanisms of coupling between the powerplant and the vehicle chassis and subsystems. The analytical procedure expands the equation of motion of the vehicle components to such that a domain of boundary conditions used in the 6 degrees-of-freedom powerplant mounting mode

7、l can be defined. An example of this new procedure is given for improving NVH chassis response at idle speed using the torque roll axis decoupling strategy.Keywords: Powerplant mounting system; Optimization; Dynamic isolation; Coupled systems1. IntroductionIn vehicles, the engine mounts play an esse

8、ntial role for the noise, vibration and harshness (NVH) comfort. The main functions of these mounts (rubber or hydraulic) are to provide static supports for the powerplant and to isolate the vibrations of the powerplant from the rest of the vehicle. To provide design characteristics necessary for th

9、e NVH improvement in terms of rigidity and damping it is essential to simulate the responses of the powerplant mounting system to low frequency vibrations. Is is essential that the model includes the primary interactions between the powerplant mounting system and each of the vehicle subsystems. In t

10、he early stages of the vehicle design most of the necessary data needed from the subsystems are not yet fully described. Thus, to begin a theoretical layout of the powerplant mounting system, some reasonable assumptions of the vehicle components must be made. Specifically, the model includes rigid b

11、ody representations of the powerplant and the chassis; with appropriate values for the location of the centers of gravity, masses, and moments of inertia. This simulation model enables the assessment of the rigid body modes of the powerplant in the vehicle. As well, the motions of the powerplant and

12、 the chassis under various engine operating conditions (idle, full load speed sweep) and road/wheel inputs can be analyzed.Equations of motion for the powerplant mounting system include parameters for a rigid chassis. On the contrary, the chassis flexibility may have a significant effect on powerpla

13、nt vibrations and mounting forces transmitted from the powerplant to the structure, especially when flexible vibration modes of the chassis are excited. The dominant vibration modes of body structure at idle speed should be the first longitudinal bending mode and the first torsional mode, normally a

14、bove 3035 Hz. Experimental verification of the simulation models assumptions through measurements of the vibration modes of the chassis should be included in future work.The current industrial strategies use a model approach to analyze the harmonic response of the powerplant on resilient supports at

15、tached to ground (Brach, 1997; Khajepour and Geisberger, 2002). The 6 degrees-of-freedom model used in the modal analysis is interesting insofar as the response to an excitation is calculated and interpreted according to the position in frequency and to the form of the modes.Typical design strategie

16、s move input source frequencies away from the rigid body natural frequencies of the powerplant in order to avoid resonances (Gray et al., 1990; Kano and Hayashi, 1994). Vibrations are minimized within this design approach by manipulating the rigid body modes of the grounded powerplant and shaping th

17、e response through the torque roll axis decoupling and the elastic axis decoupling methods attempt. The background theory of these techniques is widely described in literature (Patton and Geck, 1984; Singh and Jeong, 2000; Brach, 1997). However, by considering the powerplant to be grounded these des

18、ign strategies neglect the influences of the chassis, exhaust subsystem, drive-shaft, wheel suspension . . . .Lately, researches have focused on the significance of the rigid body modes alignment for grounded powerplant to its invehicle behavior (Sirafi and Qatu, 2003; Hadi and Sachdeva, 2003). Thes

19、e studies deal with the accuracy of NVH vehicle models and raise the problem of interactions between the different subsystems. Various powertrain models have been studied and their accuracy was discussed through a full vehicle model. By the evaluation of actual cases, the existence of these interact

20、ions have been clearly demonstrated. Nevertheless, no general formalism have been introduced to evaluate the limits of the modeling assumptions made during the development of the classical 6 degrees-of-freedom powerplant mounting schemes.The aim of the proposed method is to highlight and identify, t

21、hrough an analytical procedure, the relationships between the powerplant mounting schemes and the vehicle response characteristics. In the second section, the general equations of motion are reformulated using an original matrix, the coupling matrix introduced for coupled plates (Bessac and Guyader,

22、 1996). With the characteristics of the coupling matrix, acceptable boundary conditions used in the traditional 6 degrees-of-freedom mounting strategies can be defined for different engine operating conditions. As an example in Section 3, these parameters are defined for engine models in the idle st

23、ate. In the last section of the paper, the issue of the torque roll axis decoupling strategy is analyzed using the coupling parameters in terms of improvement of the dynamic chassis responses at idle speed.2. Formulation of the coupling problem2.1. Modelling of the vehicle systemDerivation of the eq

24、uations of motion to simulate dynamic behaviors of powerplant mounting systems with supporting structures, a good modelling of the total vehicle system can consist of four subsystems: the powerplant which includes engine and transmission, the engine mounts, the chassis and the suspension. Since smal

25、l displacements can be assumed, the powerplant is modelled as rigid body of time-invariant inertial matrix of 6 dimensions. The powerplant is supported by an arbitrary number of mounts on the vehicle chassis, that is modelled as an elastically suspended rigid body as shown in Fig. 1.The mounts class

26、ically used in powerplant mounting application are bonded metal-rubber construction. It is possible to get better isolation effects than conventional rubber mount systems with hydraulic engine mount. Hydraulic engine mount control the damping characteristics by using the fluid viscosity. Elastomeric

27、 materials behave visco-elastically, thus engine mounts are represented by three sets of .mutually perpendicular of linear springs and viscous dampers in parallel. No rotational stiffness of the mounts has been considered. The stiffness matrix Kmi and damping matrix Cmi of a mount i can be written i

28、n the local coordinate system as: ，and (1)Fig. 1. Powerplant mounting modelFig. 2. Translational u and rotational displacements of the powerplant center of gravity.The subscript mi corresponds to the mount frame coordinates Rmi (Fig. 1) of the ith mount. The stiffness and damping matrices must be tr

29、ansformed from the local mount coordinate system Rmi to the global coordinate system R by the following linear transformation: ,and (2)The element mi is the transformation matrix from the local coordinate system Rmi to the global one R. The elements of mi consist of directional cosines of the local

30、frame with respect to R defined from Euler angles.2.2. Equations of motionAnother transformation is necessary to express the equations of motion of the powerplant and chassis centers of gravity in terms of displacements and rotations. This transformation relates the displacements of each mount with

31、respect to the displacements and rotations of the powerplant and chassis centers of gravity. The superscripts (e) and (c) stands for powerplant and chassis respectively. The superscript (b) may refer to either the powerplant or the chassis. A generalized vector q (Eq. (3) is defined by combining tra

32、nslational u and rotational displacements of the centers of gravity of the powerplant (Fig. 2) and of the chassis. (3)The position vector of the ith mounts center of elasticity with respect to the center of gravity of the powerplant and the chassis are given in terms of global coordinate system as:

33、(4)and each has a corresponding skew asymmetric matrix defined as: (5)with a generalized form: (6)Let ui(b) be the translational displacement vector at the mounting point i for the rigid body (b) side. The relative translational displacement vector i for the ith mount for small motions is related to

34、 the rigid body center of gravity motions and the translational displacements at the mounting point according to Eq. (7). (7)The translational reaction force fi(e) and fi(c) and moment reactioni(e) andi(c) resulting from the application of the elastic forces of mounting i on the powerplant and the c

35、hassis centers of gravity can be expressed in the R frame as: (8) At idle speed, the connection to the ground is simply represented by four systems of linear spring and viscous damper in parallel at each wheel, characterized by their stiffness and damping coefficients following the three directions

36、of the vehicle frame coordinates R. The translational reaction force fk(c) and moment reactionk(c) from the kth suspension applied to the chassis can be expressed in the frame R with the displacement of the chassis u k(c) at the supporting point as: (9)Similarly, for the road/wheel inputs, a simple

37、model can be used for the wheel-suspension system, with a single degree of freedom. This can be represented by a mass and a spring accounting for the wheel mass and the tires stiffness in parallel with a spring and a damper accounting for the suspension system. The dynamic interaction between the ve

38、hicle suspension and the powerplant mounting system should be included in future work.Assuming all elastic loadings from all mounts and suspension, the total elastic loadings on the powerplant and chassis centers of gravity can be expressed through a generalized square stiffness matrix K of 12 dimen

39、sions (10), resulting from the assembly of the elementary stiffness matrices (mounts and suspensions).(10)The matrix K(ec) is the powerplants matrix of influence on the chassis and the reciprocal, K(ce) is the chassiss matrix of influence on the powerplant. Using a similar assembly procedure to the

40、elastic loadings, the total damping loadings on the powerplant and chassis centers of gravity can be expressed by a generalized square stiffness damping matrix C of 12 dimensions (11).(11)Since all component reactive forces are derived in terms of the generalized coordinates, and assuming small osci

41、llations, the equations of motion of the powerplant and the chassis can be written as the matrix form in the frequency domain: (12)The vector F =tF(e) F(c) is the generalized external load vector. The external excitations are harmonics with known frequencies,amplitudes and phases. Engine excitation

42、forces are applied to the powerplant at the center of the crankshaft location.The response to road inputs can be studied by applying forces or displacements at the suspensions location of ground contact.The matrix M is the generalized mass matrix of the system (13).(13)Withand m(b) is the mass of th

43、e rigid body (b) and M(b) its inertia matrix. C is the generalized viscous damping matrix assuming a proportionally damped system. If a structural damping matrix H is considered, viscous damping term jC may be replaced by the structural damping term jH. For the following developments, a complex stif

44、fness is used to model the dynamic behavior of the isolators. The bar indicates that the stiffness term is complex ().2.3. Introduction of the coupling matrixThe response of the powerplant and chassis centers of gravity can be calculated through the solving of Eq. (12). Then, the complex matrix inve

45、rsion of Eq. (14) is classically used.(14)The inversion of the impedance matrix can be numerically resolved. Nevertheless, this method hinders the understanding of the coupling phenomena between the powerplant and the chassis. From the traditional equation of motion (14), one can isolate a matrix pr

46、esenting only terms related to the coupling from the two bodies (15).(15)For the sake of physical meaning of the coupling mechanism, the term (2M(e) 1F(e) in Eq. (15) represents the displacement of the powerplant subjected to its own excitation when the chassis is blocked (suspensions with infinite

47、stiffnesses).This configuration represents the typical industrial model of the grounded behavior of the powerplant (Fig. 3(a). The term (2M(c) 1F(c) represents the displacement of the chassis subjected to his own excitation when the powerplant is blocked (null displacements) (Fig. 3(b). This configu

48、ration, however, does not represent a realistic behavior. We can express the two configurations by the generalized vector displacement of the uncoupled blocked bodies tq0(e) q0(c) (16).(16)Fig. 3. Uncoupled blocked bodies.While revealing the displacement vector of the coupled systems, Eq. (15) takes

49、 the form of a coupling matrix D (Bessac and Guyader, 1996) (Eq. 17).(17)WithEach coupling matrix term represents the action of the powerplant mass displacement (respectively chassis) on the chassis mass displacement (respectively powerplant). The matrix of coupling describes the exchange between th

50、e masses independent of the external excitation. The coupling matrix, studied in more details in Section 3, is a practical solution to predict the global behavior of a system starting from the behavior of the isolated subsystem.评估改进车辆振动噪音中发动机和底盘耦合的全新分析法摘要汽车发动机装备系统的设计是车辆安全以及汽车振动噪音改善不可或缺的重要部分。汽车设计中遇到的

51、主要问题之一是将发动机的低频振动与其它交通工具隔离开来。频率和模式安排的选择成为了质量显著的发动机的一个重要的设计决策。几个发动机装备方案已被开发出来，并且应用于改善有关定位以及弹力支撑设计的噪音性能中。将刚体模式从接地发动机模式中解耦，这些方法是以此基础的，然而，接地发动机模式忽略了底盘与悬架系统之间相互作用。但是我们不能说发动机接地刚体模式去耦将会减弱底盘系统的振动。本论文提出了一个全新的分析法来检测发动机、车辆底盘以及子系统的耦合机制。分析程序对汽车零部件的运动方程进行了扩展，如此一来，就能对应用在自由度为6度的发动机装备模式中的边界条件域进行定义了。本论文通过扭矩辊轴耦战略，给出了一个

52、全新程序实例，本实例诣在改善低速底盘噪音反应。关键词：发动机装备系统，最优化，动态隔离，耦合系统1.简介发动机架在缓和交通工具的振动噪音方面起着重要的作用。这些装备（橡胶或液压）的主要作用是为发动机提供静态支撑，以及将其振动与其它交通工具隔离。如果要想从刚性和阻尼方面得到改善振动噪音所需的设计特征，就很有必要将发动机装备系统的反应刺激至低频振动，这一点非常重要。这个模式包括发动机装备系统与每个车辆子系统之间的基本相互作用。在车辆设计的初级阶段，大多数子系统所需的必要数据尚未被完全描述出来。因此，想要开始发动机装备系统的理论设计，就必须对车辆零部件做出一些合理设想。具体来讲，该模式包含发动机和底

53、盘的刚体再现，以及具有重心位置，质量和惯性矩的恰当值。该仿真模型有助于发动机刚体模式的评估。同样，不同发动机条件下（闲置或满载负载速度下）的发动机和底盘的运动都能被分析出来。发动机装备系统的运动方程包含一个刚性底盘的参数。与之对应的，底盘灵活性可能会对发动机震动和由发动机传输至结构的装备力量，尤其是在底盘的弹性震动模式被激活的时候，产生强烈作用。阀体结构在空转速度时的显性运动模式应该是第一纵弯曲模式和第一扭转模式，通常是在3035赫兹之上。通过测量底盘振动模式，得出经实验鉴定的仿真模型假设，这一点需被列入今后的工作中。现有的工业战略用模型方法来分析发动机在接地弹性支撑上的和声反应（布拉齐，19

54、97；Khajepour and Geisberger, 2002）。模式分析中所采用的6度自由模型是有趣的，它对刺激的反应可以被计算出来，也能根据频率定位和模式形式被解释出来。典型的设计战略将输入端电源频率从发动机刚性主体自然频率中移除以避免共振（Gray 等，1990；卡诺和林文夫，1994）。该设计方法能通过两种方式将振动最小化，一种是手动操作接地发动机的刚体模式，另一种是令转矩轧辊轴线解耦和弹性轴解耦形成反应的方式。很多文献中都讨论了这些技术的背景理论（Patton and Geck, 1984; Singh and Jeong, 2000; Brach, 1997）。但是，考虑到将要

55、接地的发动机，这些设计策略忽略了底盘、排气子系统、驱动器轴和车轮悬浮等的影响。最近，很多研究都集中在接地发动机的刚体模式校准对其车辆行为的影响上（Sirafi and Qatu, 2003; Hadi and Sachdeva, 2003）。这些研究论述了振动噪音车辆模型的准确性，并引发了不同子系统之间的相互作用的问题。目前已有人研究出了各种动力系统模型，并使用完整的车辆模型讨论其准确性。对实际案例进行评估发现：这些相互作用是存在的。然而，目前还没有引入任何形式体系来评估在6度自由发动机装备方案的发展中制定的关于模型假设的限制性。所提出的方法诣在通过分析程序强调和确认发动机装备方案和车辆反应特

56、点之间的关系。在第二部分中，作者采用原始矩阵模型以及为耦合平板引进的耦合矩阵模型，再次形成了运动的一般方程（Bessac and Guyader, 1996）。通过耦合矩阵模型的特点，就能为不同的引擎操作条件定义传统6度自由装备战略所采用的可接受的边界条件。正如第三部分的例子所示：这些参数是为处于静止状态的引擎模型定义的。该论文的最后一部分通过使用改善空转速度下动态底盘反应的耦合参数，分析了转矩轧辊轴线去耦战略。2. 耦合问题构想2.1.车辆系统模型化为了引出运动方程来刺激具有支撑结构的发动机装备系统的动态行为，总车辆系统的好模型由4个子系统构成：包含引擎和传动装置的发动机、发动机架、底盘和悬

57、架。由于可对小排量做出假设，发动机被设计成了6维不变时的惯性矩阵刚体模型。发动机被车辆底盘上任意数量的装备支撑，那就形成了如图1所示的弹性悬浮刚体模型。发动机装备应用系统通常使用的挂架是由金属和橡胶制作而成的。有了液压发动机架，就有可能带来比传统的橡胶装备系统更好的隔离效果。液压发动机架通过采用流体粘度来控制减幅特征。弹性材料的运动也是有弹性的，因此，发动机架是由三组相互垂直的线性弹簧以及相互平行的粘性减震器为代表的。没有考虑挂架的旋转劲度。在局部坐标系中，可通过如下方式表达挂架i的刚性矩阵Kmi和阻尼矩阵Cmi： ，and (1)图1：弹性悬浮刚性模型图2：发动机重心平移u和旋转位移,and

58、 (2)相当于框架的下标mi能与ith挂架的Rmi（图1）相协调。刚体和阻尼矩阵必须按照以下线性转化，从局部挂架坐标系Rmi转换至总坐标系R。mi是从局部坐标系Rmi转换至总坐标系R的转换矩阵。mi的构成元素包含与由欧拉角定义的与R相关的局部框架的方向余弦。2.2. 运动方程很有必要用另外一个变形来表达发动机和底盘重心在移位和旋转方面的运动方程。该变形描述了有关发动机和底架重心的移位和旋转。下标(e)和(c)分别代表发动机和底盘。下标(b)可能指的是发动机和底盘二者之一。通过结合发动机（图2）和底盘重心平移的u和旋转的位移来定义一个广义矢量q （方程3）。 (3)依照总坐标系，这里给出了有关发

59、动机和底盘重心的ith挂架弹性中心的位矢： (4)每一个都具有对应的斜交非对称矩阵，定义如下：, (5)广义形式为： (6)将ui(b)看成刚体(b)侧装备点i的平行位移矢量。根据方程7，ith挂架的小幅度运动相对平行位移矢量i是与刚体重心运动以及装备点的平行位移相关的。 (7)由发动机和底盘重心i装备的弹力所产生的平移反作用力fi(e) 和fi(c)以及瞬间作用力i(e) 和i(c)可在R框架内按如下方程表达： (8)在空转速度，接地仅由每个车轮4组线性弹簧和平行粘性减震器代表，其特点为：它们的刚性和阻尼系数遵循车架坐标R的3个方向。底盘Kth悬架的平行反作用力fk(c)和瞬间作用力k(c)

60、可在总坐标系R（在该坐标系中，底盘在支撑点有位移：u k(c)中表达为： (9)同样，路段/车轮输入可在车轮悬浮系统中使用一个自由度单一的简单模型。这是以能够解释车轮质量的弹簧和质量与能够解释悬浮系统的弹簧和减震器相互平行为代表的。车辆悬浮和发动机装备系统间的动态相互作用需被列入到未来工作中。假设所有弹性负荷均来自挂架和悬浮，发动机和底盘重心的弹性负荷可通过12维一般方形刚性矩阵K来表达，该矩阵由基础刚性矩阵（挂架和悬浮）组装而成。(10) K(ec)矩阵是发动机对底盘的影响力矩阵，它的求逆K(ec)是底盘对发动机的影响力矩阵。对弹性负荷使用类似的组装程序，发动机和底盘重心的总阻尼负荷可通过12维一般方形刚性阻尼矩阵C来表示。(11)由于所有组件反作用力是由总坐标系派生出来的，假设振幅很小，发动机和底盘运动方程可在频域内写成矩阵形式： (12)F =tF(e) F(c)这一矢量是广义的外加载矢量。外部刺激和已知的频率、振幅以及阶段都是和声。发动机刺激力被运用到了发动机机轴中心。对路段输入的反应可通过在接地悬浮位置施加外力或者位移来研究。M矩阵是该系统（13）的广义质量矩阵。 (13)与