外文翻譯--具有車架結(jié)構(gòu)車輛的怠速震動(dòng)分析

1、<p><b>　　附　　錄</b></p><p>　　附錄A 外文文獻(xiàn)原文</p><p>　　An Analysis of Idling Vibration for a Frame Structured Vehicle</p><p><b>　　ABSTRACT</b></p><p&

2、gt;　　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 Souma'

3、;s 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 frequen

4、cy of the first-order vertical bending mode for the frame. It has bec</p><p>　　INTRODUCTION</p><p>　　One of the defects of a diesel vehicle, which has fuel and economical efficiency, is idling v

5、ibration 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

6、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</p><p>　　The idling

7、 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 Souma‘s method,and entering the

8、m to a vehicle model.</p><p>　　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 f

9、or 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 vertic

10、al bending for the frame had a significant effect.</p><p>　　ANALYSIS OF THE VEHICLE BODY VIBRATION</p><p>　　Figure 1 shows the results of analyzing the frequencies of the acceleration in vertica

11、l 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

12、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 res</p><p>　　IDE

13、NTIFICATION OF THE ENGINE EXCITING FORCE</p><p>　　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 pa

14、th 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, Souma’s method is used.</p><p&

15、gt;　　OUTLINE OF SOUMA’S METHOD</p><p>　　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 act

16、s 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 r

17、igid in this frequency domain. In this situation, the engine block will not vibrate if the piston crankshaft does</p><p>　　Accordingly, the direct cause of the engine block vibration is not the combustion pr

18、essure 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

19、works inside and outside of the engine block, may be calculated.</p><p>　　In Souma’s method, the non-uniform rotary motion in the crank system is found by measuring the pulse generated in a ring gear of the

20、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.</p><p>　　VERIFICATION OF THE ACCURACY

21、 IN THE EXCITING FORCE</p><p>　　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

22、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 considerin

23、g the idling rotation speed changes.</p><p>　　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

24、 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 af

25、fects the engine vibration. As the data of the experiment and the calculation is identified in this domain, the power p</p><p>　　However, around 12 Hz of 1st-orders, data is not much identified. In this freq

26、uency 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.</p><p>　　IMPROVEMENT OF

27、THE MEASURING ACCURACY IN LOW-FREQUENCY VIBRATION</p><p>　　The engine exciting force was determined using Souma’s method, and the vibration in each part of the engine was calculated by adding the exciting fo

28、rce. 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

29、 vibration, which causes the idling vibration, and the high-frequency vibration, which causes noise, are mixed. When the mixed vibra</p><p>　　In this paper, a strain gage acceleration pickup, which measures

30、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 dete

31、cting 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,</

32、p><p>　　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 exper

33、iment of movement performance, was used for an engine.</p><p>　　Fig. 1 Seat rail vertical vibration Fig. 2 Head cover lateral vibration</p><p>　　Fig. 3 Right engine foot vertical v

34、ibration Fig.4 Measurement with piezo element acceleration pickup</p><p>　　ENTIRE VEHICLE MODEL</p><p>　　Figure 6 shows the body model. Interior and exterior equipments such as doors and seat

35、 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 combi

36、ned, and a fuel tank and a bumper is added in the form of concentrated mass. The grid points are 39,262.</p><p>　　Combining the models shown in Figure 6 and 7 using cabmount makes the entire vehicle model. T

37、otal 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 c

38、alculated by the block structure.</p><p>　　Figure 8 shows the frequency response function, indicating the responses of the frame with the right back engine mount after exciting the driver’s seat rail. In the

39、 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.</p><p>　　Fig. 5 Measurement with strain gage acc

40、eleration pickup Fig. 6 Body mode</p><p>　　Fig.7 Frame,power plant and suspension model Fig.8 Frequency response function</p><p>　　CORRELATION ANALYSIS OF THE MODES</p><p&

41、gt;　　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 i

42、s not complete. When the entire vehicle is divided into block structures by each insulating mount and suspension, the body has 4 block structures:</p><p>　　(1) Block where interior equipment is added in the

43、form of concentrated mass to the body as shown in Figure 6, which is described as “body”, hereafter.</p><p>　　(2) Block where the fuel tank and the bumper are added in the from of concentrated mass to the fr

44、ame as shown in Figure 7, which is described as “frame,” hereafter. </p><p>　　(3) Power plant </p><p>　　(4) Suspension</p><p>　　Among the above block structures, (1) body and (2) fr

45、ame 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.</p><p>　　COMPARISON OF

46、NATURAL FREQUENCY</p><p>　　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, th

47、e 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 5

48、0 Hz.</p><p>　　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

49、) 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.</p><p

50、>　　Fig. 9 Natural modes in frequency domain</p><p>　　附錄B 外文文獻(xiàn)中文翻譯</p><p>　　具有車架結(jié)構(gòu)車輛的怠速震動(dòng)分析</p><p><b>　　摘要</b></p><p>　　建立全車架結(jié)構(gòu)SUV的有限元模型，用來(lái)評(píng)價(jià)車輛的怠速震動(dòng)特性。用S

51、ouma理論確定發(fā)動(dòng)機(jī)的動(dòng)力來(lái)模擬怠速震動(dòng)。發(fā)動(dòng)機(jī)和整車的模型通過(guò)實(shí)驗(yàn)和計(jì)算結(jié)果協(xié)調(diào)以后共同決定。注意力放在了車架一階縱向彎曲模型的頻率上。降低一階車架彎曲模型的頻率可以減少車輛的怠速震動(dòng)已經(jīng)變得明確。</p><p><b>　　簡(jiǎn)介</b></p><p>　　具有燃油經(jīng)濟(jì)性的柴油車的一個(gè)缺點(diǎn)就是車身的怠速震動(dòng)。在柴油發(fā)動(dòng)機(jī)里，由熱能積聚引起的壓力急劇上升會(huì)影響活

52、塞。在把直線運(yùn)動(dòng)轉(zhuǎn)換成旋轉(zhuǎn)運(yùn)動(dòng)的曲軸系統(tǒng)里，有兩種反作用力使得發(fā)動(dòng)機(jī)體振動(dòng)：由移動(dòng)部件運(yùn)動(dòng)換向引起的反作用力，和有限的氣缸不均勻的轉(zhuǎn)動(dòng)引起的。這個(gè)力傳遞到發(fā)動(dòng)機(jī)機(jī)體，發(fā)動(dòng)機(jī)底部，橡膠的發(fā)動(dòng)機(jī)支座，車架，橡膠駕駛室支架，最后到車身，引起乘客不舒服。</p><p>　　大型商用車的怠速震動(dòng)的平復(fù)處于發(fā)展的初期，用Souma理論模擬發(fā)動(dòng)機(jī)震動(dòng)，然后建立模型。</p><p>　　這篇論文中，將

53、發(fā)動(dòng)機(jī)置于車中來(lái)確定怠速震動(dòng)，因?yàn)檐嚰芎蛙嚿淼挠邢拊划?dāng)做一個(gè)小型休閑車。另外，在這篇文章中，也分析了車輛車架自然模式如何改變，并且指出車架一階縱向彎曲的自然頻率具有重要的影響。</p><p><b>　　車身震動(dòng)的分析</b></p><p>　　圖A1顯示了四缸柴油機(jī)RV怠速過(guò)程中座椅扶手處采集的加速過(guò)程中縱向震動(dòng)頻率的分析。怠速震動(dòng)的主要部分是二階發(fā)動(dòng)機(jī)轉(zhuǎn)動(dòng)，

54、第0.5，第1，和第1.5階同樣重要。但是，這些不同是由于不同氣缸的燃燒不同而引起的。完善噴射系統(tǒng)可以解決燃燒的差異。在這個(gè)實(shí)驗(yàn)中，只集中研究怠速轉(zhuǎn)速是720rmp時(shí)24Hz車架的二階震動(dòng)。此外，也研究了降低振動(dòng)的措施，因?yàn)樽蔚目v向振動(dòng)對(duì)人類的感覺有很大的破壞性影響。</p><p>　　發(fā)動(dòng)機(jī)引起作用力的判定</p><p>　　發(fā)動(dòng)機(jī)將振動(dòng)傳遞給車身的路線有三種：通過(guò)發(fā)動(dòng)機(jī)支座，驅(qū)動(dòng)

55、系統(tǒng)，和尾氣排放管。在這篇論文中，研究了起主要作用的發(fā)動(dòng)機(jī)支座的路線。研究方法有很多種，這里用Souma理論。</p><p>　　Souma理論的概要</p><p>　　考慮引起發(fā)動(dòng)機(jī)集體受力的有爭(zhēng)議的怠速振動(dòng)頻率范圍。首先，作用在活塞上的燃燒壓力被認(rèn)為引起這個(gè)振動(dòng)。但是，假設(shè)活塞曲軸并不隨飛輪移動(dòng)并且機(jī)體以某種方式固定，在這個(gè)頻率范圍發(fā)動(dòng)機(jī)的零件被認(rèn)為是完全剛性的。在這種情況下，如果

56、活塞曲軸不移動(dòng)，發(fā)動(dòng)機(jī)機(jī)體就不會(huì)振動(dòng)，盡管柴油燃燒引起壓力的迅速上升。</p><p>　　相應(yīng)地，引起發(fā)動(dòng)機(jī)機(jī)體振動(dòng)的直接原因不是燃燒壓力，而是活塞曲軸運(yùn)動(dòng)的反作用力。為了確定作用在發(fā)動(dòng)機(jī)機(jī)體上的這個(gè)力，需要計(jì)算在機(jī)體內(nèi)外都發(fā)揮作用的反作用力。</p><p>　　在Souma理論里，通過(guò)測(cè)量在飛輪齒圈上收集到的脈沖來(lái)發(fā)現(xiàn)曲軸系統(tǒng)的不協(xié)調(diào)旋轉(zhuǎn)運(yùn)動(dòng)。然后計(jì)算相連的活塞系統(tǒng)的縱向運(yùn)動(dòng)來(lái)確定發(fā)

57、動(dòng)機(jī)機(jī)體上的作用力。</p><p><b>　　作用力準(zhǔn)確性的驗(yàn)證</b></p><p>　　在整車模型里（后續(xù)描述），振動(dòng)力的增加和曲軸是對(duì)應(yīng)的。評(píng)估振動(dòng)主要影響的引擎蓋和發(fā)動(dòng)機(jī)右側(cè)底部。計(jì)算數(shù)據(jù)和實(shí)驗(yàn)結(jié)果的比較結(jié)論在圖A2和圖A3中表示了出來(lái)。在圖A2和圖A3中，表示出來(lái)5種不同的計(jì)算結(jié)果，因?yàn)橐紤]怠速轉(zhuǎn)速的變化。</p><p>　

58、　在圖A2和圖A3中，鑒定了在轉(zhuǎn)速為720rpm時(shí)第二第四和第六階的24Hz，48Hz和72Hz的計(jì)算數(shù)據(jù)和實(shí)驗(yàn)結(jié)果。發(fā)動(dòng)機(jī)左側(cè)底部的數(shù)據(jù)，在這篇論文中沒有顯示出來(lái)，但是也幾乎全部鑒定了出來(lái)。至于在這個(gè)頻率范圍內(nèi)，發(fā)動(dòng)機(jī)和車身的振動(dòng)被發(fā)動(dòng)機(jī)支座隔離開來(lái)。車身幾乎影響不到發(fā)動(dòng)機(jī)的振動(dòng)。因?yàn)閷?shí)驗(yàn)數(shù)據(jù)和計(jì)算結(jié)果的鑒定是在這一范圍內(nèi)，動(dòng)力模型和振動(dòng)力可以認(rèn)為是合理的。</p><p>　　但是在一階12Hz周圍，數(shù)據(jù)并沒

59、有鑒定出來(lái)。在這一頻率范圍內(nèi)，發(fā)動(dòng)機(jī)和車身的振動(dòng)被發(fā)動(dòng)機(jī)支座耦合到了一起，因此，車身模型的準(zhǔn)確定受到影響。</p><p>　　低頻振動(dòng)測(cè)量方式的改善</p><p>　　發(fā)動(dòng)機(jī)振動(dòng)力通過(guò)Souma理論來(lái)確定，通過(guò)增加振動(dòng)力，發(fā)動(dòng)機(jī)每個(gè)部分的震動(dòng)都被計(jì)算出來(lái)。至此，然而，計(jì)算數(shù)據(jù)并沒有和實(shí)際測(cè)量完全區(qū)分開來(lái)。因此，實(shí)際測(cè)量的準(zhǔn)確性得到提高。引起怠速振動(dòng)的低頻振動(dòng)和引起噪聲的高頻振動(dòng)在發(fā)動(dòng)機(jī)

60、表面混合到一起。當(dāng)通過(guò)壓力測(cè)量這個(gè)混合振動(dòng)，高頻率的振動(dòng)被加重，而作為研究目標(biāo)的低頻率表振動(dòng)則變得相對(duì)小了。舉個(gè)例子，測(cè)量發(fā)動(dòng)機(jī)右側(cè)底部的縱向振動(dòng)加速-時(shí)間波形如圖A4所示。</p><p>　　在這篇論文中，運(yùn)用了測(cè)量壓力作用在內(nèi)部的重量的加速壓力計(jì)。這個(gè)裝置比壓力元素加速機(jī)更大，對(duì)加速也更敏感。除此之外，內(nèi)部為了保護(hù)探測(cè)部分而填充的硅油阻止了高頻振動(dòng)。這個(gè)裝置測(cè)得的加速-時(shí)間縱向振動(dòng)波形如圖A5所示。和圖A4

61、相比，圖A5僅僅顯示出了低頻率，雖然測(cè)量的是相同的區(qū)域。通過(guò)這種方式，高頻率振動(dòng)被阻截掉，因此明暗度更高。這一次，使用了加速度測(cè)量范圍0到20m/s2的裝置。因?yàn)殪`敏度高，這個(gè)裝置很容易校準(zhǔn)，通過(guò)重力加速度。使用壓力加速度檢測(cè)計(jì)的時(shí)候，主階振動(dòng)計(jì)算數(shù)據(jù)和實(shí)驗(yàn)結(jié)果的差異是20-40%。因此，采用這一方式的Souma理論處于質(zhì)疑中。然而，圖A2和A3是采用流量計(jì)加速度檢測(cè)計(jì)鑒定出來(lái)的計(jì)算結(jié)果。</p><p>　　圖

62、B1 座椅扶手縱向振動(dòng) 圖B2 缸蓋橫向振動(dòng)</p><p>　　圖B3 右側(cè)發(fā)動(dòng)機(jī)底部縱向振動(dòng) 圖B4 壓力加速度計(jì)測(cè)量結(jié)果</p><p><b>　　整車模型</b></p><p>　　圖A6是整車模型。像車門和座椅等內(nèi)部和外部裝置以85點(diǎn)增加到詳細(xì)有限元結(jié)構(gòu)模型中。網(wǎng)格數(shù)是61,

63、912。圖A7是一個(gè)有懸架，發(fā)動(dòng)機(jī)，燃料箱和保險(xiǎn)杠的車架組合成一個(gè)整體，網(wǎng)格數(shù)是39,262。</p><p>　　把圖A6和圖A7組個(gè)到一起形成了一個(gè)整車模型，總的網(wǎng)格數(shù)是101,174。使用70.5.2版本的IBMSP2, MSC/NASTRAN計(jì)算時(shí)間是3,293。計(jì)算方法是打包計(jì)算。如果模型是更大規(guī)模，則必須通過(guò)整體結(jié)構(gòu)計(jì)算。</p><p>　　圖A8是頻率響應(yīng)函數(shù)，指示出振動(dòng)力

64、作用在發(fā)動(dòng)機(jī)支座時(shí)車架的響應(yīng)。在需要分析的20到30Hz頻率范圍里，實(shí)驗(yàn)數(shù)據(jù)相對(duì)于計(jì)算結(jié)果更好。</p><p>　　圖B5 流量計(jì)加速度檢測(cè)機(jī)的測(cè)量結(jié)果 圖B6 車身</p><p>　　圖B7 車架，發(fā)動(dòng)機(jī)和懸架模型 圖B8 頻率響應(yīng)函數(shù)</p><p>　　圖B9 自然模式的頻率

65、范圍</p><p><b>　　模型相關(guān)性分析</b></p><p>　　從振動(dòng)特性的角度來(lái)看，可以認(rèn)為整車振動(dòng)被發(fā)動(dòng)機(jī)支座和駕駛室支座隔離開來(lái)，因?yàn)橛袕椈蛇B接，雖然隔離并不徹底。如果整車被連接件和懸掛分開，車身有4大結(jié)構(gòu)：</p><p>　　(1)機(jī)體 增加了內(nèi)部零件，如圖A6所示，此后描述成機(jī)體。</p><p&g

66、t;　　(2)車架 車架上增加了燃料箱和保險(xiǎn)杠，如圖A7所示，此后描述成車架</p><p><b>　　(3)動(dòng)力系統(tǒng)</b></p><p><b>　　(4)懸架</b></p><p>　　在以上的結(jié)構(gòu)中，(1)機(jī)體和(2)車架怠速振動(dòng)的自然頻率在24Hz附近。機(jī)體，車架和整車模型的振動(dòng)特性被比較和研究。</

67、p><p><b>　　自然頻率的比較</b></p><p>　　圖A9顯示的是每個(gè)結(jié)構(gòu)和車輛不同狀態(tài)下自然振動(dòng)頻率的分布情況。車架有17個(gè)自然模式低于50Hz。在圖7中，裝有發(fā)動(dòng)機(jī)和懸架的車架，Y型底盤，有35個(gè)自然模式低于50Hz。Y型底盤加裝一個(gè)車身就形成了整車模型，具有94個(gè)自然模式低于50Hz。</p><p>　　當(dāng)Y型底盤的自然模式