外文翻译--具有车架结构车辆的怠速震动分析

附 录 附录 A 外文文献原文 An Analysis of Idling Vibration for a Frame Structured Vehicle ABSTRACT A finite element model for an entire frame-structured sports utility vehicle was made to evaluate the characteristics of the idling vibrations for the vehicle. The engine exciting forces were determined by Soumas method to simulate the idling vibrations. The modeling of the power plant and the entire vehicle was verified by the reasonable agreement of the experiment and calculation results. Attention was focused on the frequency of the first-order vertical bending mode for the frame. It has become clear that the idling vibration level of the vehicle is lowered by decreasing the frequency of the first-order frame bending mode. INTRODUCTION One of the defects of a diesel vehicle, which has fuel and economical efficiency, is idling vibration for a vehicle body. In a diesel engine, sharp pressure rise caused by the generation of the thermal energy affects the pistons. In the crank system, which converts the linear motion into the rotary motion, two types of reaction forces excite the engine block: the reaction caused by the alternation of the velocity vector in each moving parts, and by the non-uniform rotary motion generated by the finite number of cylinders. The forces transmit to an engine block, an engine foot, a rubber engine mount, a frame, a rubber cab-mount, and then a vehicle body, which make occupants uncomfortable. The idling vibration for large-sized commercial vehicles was estimated at the early development stage, and the measures against the vibration were taken by simulating the engine exciting forces with Soumas method,and entering them to a vehicle model. In this paper, the idling vibration was determined by entering the engine exciting forces to the vehicle model, which was made of the finite element of the frame and the body for a small-sized recreational vehicle (RV). Also in this paper, how the natural modes for the frame changes in the vehicle condition is analyzed, and it was indicated that the natural frequency of the first-order vertical bending for the frame had a significant effect. ANALYSIS OF THE VEHICLE BODY VIBRATION Figure 1 shows the results of analyzing the frequencies of the acceleration in vertical vibration generated on the seat rail while idling in small-sized RV powered by 4-cylinder diesel engine. The main part of the idling vibration is the second-order engine rotation. The 0.5th, 1st, and 1.5th -orders are also critical. However, these orders are caused by the varied combustion between cylinders. A measure against the varied combustion can be expected by improving the injection system. In this research, only 24Hz of the second-order at the idling rotation speed of 720rpm is focused on as a measure in the vehicle structure. Besides, a measure for lowering the vibration is studied because the vertical vibration on seats has a great damaging effect on human sense. IDENTIFICATION OF THE ENGINE EXCITING FORCE There are three paths for the engine to excite vibration to a vehicle body: through an engine mount, a driving system, and a tail pipe. In this paper, the path through an engine mount, which has a greatest effect, is studied. The various types of methods to identify the exciting force through an engine mount are known. In this paper, Soumas method is used. OUTLINE OF SOUMAS METHOD The cause of the exciting force to an engine block in the controversial frequency domain of the idling vibration is considered. First, the combustion pressure that acts on the pistons is considered to cause the vibration. However, assuming that a piston crankshaft does not move with a flywheel and an engine block fixed in some way, the engine components are supposed to be completely rigid in this frequency domain. In this situation, the engine block will not vibrate if the piston crankshaft does not move in spite of the rapid pressure rise in a combustion chamber due to the diesel combustion. Accordingly, the direct cause of the engine block vibration is not the combustion pressure but the reaction against the piston crankshaft movement. To determine the exciting force to the engine block, the reaction forces against the movement of the mass (mainly in crank system and piston system), which works inside and outside of the engine block, may be calculated. In Soumas method, the non-uniform rotary motion in the crank system is found by measuring the pulse generated in a ring gear of the flywheel. Then, the vertical motion in the connected piston system is calculated to determine the exciting force to the engine block using each mass specification value. VERIFICATION OF THE ACCURACY IN THE EXCITING FORCE The exciting forces are added at the point corresponding to the crankshaft on the entire vehicle model (described later). The vibration on the head cover and the right engine foot, which the exciting forces mostly affect, is estimated. The results of comparing the calculation with the experiment are shown in Figure 2 and 3. In Figure 2 and 3, 5 types of calculated results are shown considering the idling rotation speed changes. In Figure 2 and 3, the calculation and the experiment are identified around 24 Hz, 48 Hz, and 72 Hz of 2nd, 4th, and 6th-orders at the speed of 720 rpm. The data of the left engine foot, which is not shown in this paper, is also almost identified. In this frequency domain, as for the vibration, the engine and the vehicle body are insulated by the engine mount. The body hardly affects the engine vibration. As the data of the experiment and the calculation is identified in this domain, the power plant modeling and the exciting force can be considered reasonable. However, around 12 Hz of 1st-orders, data is not much identified. In this frequency domain, the vibration of the engine and the vehicle body are mutually coupled through the engine mount. Therefore, the accuracy of the vehicle body model has a damaging effect. IMPROVEMENT OF THE MEASURING ACCURACY IN LOW-FREQUENCY VIBRATION The engine exciting force was determined using Soumas method, and the vibration in each part of the engine was calculated by adding the exciting force. So far, however, the calculated data has not been much identified with the actual measurement. Therefore, the accuracy of the actual measurement is improved. In the surface vibration of the engine, the low-frequency vibration, which causes the idling vibration, and the high-frequency vibration, which causes noise, are mixed. When the mixed vibration is measured with a piezo element acceleration pickup, the high-frequency order is emphasized and the target low-frequency order becomes relatively small. For example, the measured acceleration to time waveform for the vertical vibration in the right engine foot is shown in Figure 4. In this paper, a strain gage acceleration pickup, which measures force acting on the inner weight by strain, is used. This device, which is larger than a piezo element acceleration pickup, is more sensitive to the acceleration. Besides, silicon oil is filled inside to protect the detecting parts in this device, which mechanically blocks off the high-frequency order. The measured acceleration to time waveform for the vertical vibration with the device is shown in Figure 5. Compared with Figure 4, Figure 5 shows only the low-frequency order although the same area was measured. In this way, the high-frequency order is blocked off, which results in the higher sensitivity with the device. This time, the device, which measures the acceleration ranging from 0 to 20m/s2,was used. This device is easily calibrated using G-forces because it has the higher sensitivity. When a piezo element acceleration pickup was used, the differences between the calculation and the experiment were 20-40% in the main order of the vibration, and a few times in other orders. Therefore, the principle of Soumas method using a piezo element acceleration pickup has been in doubt. However, the data of the experiment and the calculation has been identified as shown in Figure 2 and 3 since a strain gage acceleration pickup, which has been used in the experiment of movement performance, was used for an engine. Fig. 1 Seat rail vertical vibration Fig. 2 Head cover lateral vibration Fig. 3 Right engine foot vertical vibration Fig.4 Measurement with piezo element acceleration pickup ENTIRE VEHICLE MODEL Figure 6 shows the body model. Interior and exterior equipments such as doors and seat are added in the form of 85 mass points to the main structure modeling detailed with sheet metal finite elements. The grid points are 61,912. Figure 7 shows the model where a frame, a suspension, and an engine are combined, and a fuel tank and a bumper is added in the form of concentrated mass. The grid points are 39,262. Combining the models shown in Figure 6 and 7 using cabmount makes the entire vehicle model. Total grid points mounts to 101,174. The calculation time is 3,293 seconds using IBMSP2, MSC/NASTRAN Version 70.5.2. The calculating method is package calculation. If the model becomes on larger scale, the model must be calculated by the block structure. Figure 8 shows the frequency response function, indicating the responses of the frame with the right back engine mount after exciting the drivers seat rail. In the frequency ranging from 20 to 30 Hz, which is required for the analysis, the data of the experiment is qualitatively identified with that of the calculation. Fig. 5 Measurement with strain gage acceleration pickup Fig. 6 Body mode Fig.7 Frame,power plant and suspension model Fig.8 Frequency response function CORRELATION ANALYSIS OF THE MODES From the viewpoint of vibration characteristics, it can be considered that an entire vehicle is insulated by the engine mount and the cabmount, which have relatively small spring constants, although the insulation is not complete. When the entire vehicle is divided into block structures by each insulating mount and suspension, the body has 4 block structures: (1) Block where interior equipment is added in the form of concentrated mass to the body as shown in Figure 6, which is described as “body”, hereafter. (2) Block where the fuel tank and the bumper are added in the from of concentrated mass to the frame as shown in Figure 7, which is described as “frame,” hereafter. (3) Power plant (4) Suspension Among the above block structures, (1) body and (2) frame have the natural frequency around 24 Hz in the idling vibration. The vibration characteristics for the body, the frame and the entire vehicle model are compared and investigated. COMPARISON OF NATURAL FREQUENCY Figure 9 shows the distribution of the natural vibration frequency in each block structure and in the vehicle condition. The frame has 17 natural modes below 50Hz. In Figure 7, the model mounting a power plant and a suspension on the frame, is called Y chassis, which has 35 natural modes below 50 Hz. Y chassis makes the entire vehicle model by mounting the body, which has 94 natural modes below 50 Hz. When the number of natural modes of Y chassis is added to 61 natural modes of the body, total number of the modes amounts to 96. The number of the natural modes of the entire vehicle model (94) is less than the above total number by 2 modes. This is because 2 natural modes became above 50 Hz by combining Y chassis and the body, as the result of analyzing the mode correlation described later. Fig. 9 Natural modes in frequency domain 附录 B 外文文献中文翻译 具有车架结构车辆的怠速震动分析 摘要 建立全车架结构 SUV 的有限元模型，用来评价车辆的怠速震动特性。用Souma 理论确定发动机的动力来模拟怠速震动。发动机和整车的模型通过实验和计算结果协调以后共同决定。注意力放在了车架一阶纵向弯曲模型的频率上。降低一阶车架弯曲模型的频率可以减少车辆的怠速震动已经变得明确。 简介 具有燃油经济性的柴油车的一个缺点就是车身的怠速震动。在柴油发动机里，由热能积聚引起的压力急剧上 升会影响活塞。在把直线运动转换成旋转运动的曲轴系统里，有两种反作用力使得发动机体振动：由移动部件运动换向引起的反作用力，和有限的气缸不均匀的转动引起的。这个力传递到发动机机体，发动机底部，橡胶的发动机支座，车架，橡胶驾驶室支架，最后到车身，引起乘客不舒服。 大型商用车的怠速震动的平复处于发展的初期，用 Souma 理论模拟发动机震动，然后建立模型。 这篇论文中，将发动机置于车中来确定怠速震动，因为车架和车身的有限元被当做一个小型休闲车。另外，在这篇文章中，也分析了车辆车架自然模式如何改变，并且指出车架一阶纵向弯 曲的自然频率具有重要的影响。 车身震动的分析 图 A1 显示了四缸柴油机 RV 怠速过程中座椅扶手处采集的加速过程中纵向震动频率的分析。怠速震动的主要部分是二阶发动机转动，第 0.5，第 1，和第1.5 阶同样重要。但是，这些不同是由于不同气缸的燃烧不同而引起的。完善喷射系统可以解决燃烧的差异。在这个实验中，只集中研究怠速转速是 720rmp 时24Hz 车架的二阶震动。此外，也研究了降低振动的措施，因为座椅的纵向振动对人类的感觉有很大的破坏性影响。 发动机引起作用力的判定 发动机将振动传递给车身的路线有三种：通过发动机支座， 驱动系统，和尾气排放管。在这篇论文中，研究了起主要作用的发动机支座的路线。研究方法有很多种，这里用 Souma 理论。 Souma 理论的概要 考虑引起发动机集体受力的有争议的怠速振动频率范围。首先，作用在活塞上的燃烧压力被认为引起这个振动。但是，假设活塞曲轴并不随飞轮移动并且机体以某种方式固定，在这个频率范围发动机的零件被认为是完全刚性的。在这种情况下，如果活塞曲轴不移动，发动机机体就不会振动，尽管柴油燃烧引起压力的迅速上升。 相应地，引起发动机机体振动的直接原因不是燃烧压力，而是活塞曲轴运动的反作用力。为了确 定作用在发动机机体上的这个力，需要计算在机体内外都发挥作用的反作用力。 在 Souma 理论里，通过测量在飞轮齿圈上收集到的脉冲来发现曲轴系统的不协调旋转运动。然后计算相连的活塞系统的纵向运动来确定发动机机体上的作用力。 作用力准确性的验证 在整车模型里（后续描述），振动力的增加和曲轴是对应的。评估振动主要影响的引擎盖和发动机右侧底部。计算数据和实验结果的比较结论在图 A2 和图A3 中表示了出来。在图 A2 和图 A3 中，表示出来 5 种不同的计算结果，因为要考虑怠速转速的变化。 在图 A2 和图 A3 中，鉴定了在转速为 720rpm 时第二第四和第六阶的 24Hz，48Hz 和 72Hz 的计算数据和实验结果。发动机左侧底部的数据，在这篇论文中没有显示出来，但是也几乎全部鉴定了出来。至于在这个频率范围内，发动机和车身的振动被发动机支座隔离开来。车身几乎影响不到发动机的振动。因为实验数据和计算结果的鉴定是在这一范围内，动力模型和振动力可以认为是合理的。 但是在一阶 12Hz 周围，数据并没有鉴定出来。在这一频率范围内，发动机和车身的振动被发动机支座耦合到了一起，因此，车身模型的准确定受到影响。 低频振动测量方