【文章內(nèi)容簡(jiǎn)介】 性，車(chē)身振動(dòng)的固有頻率應(yīng)為人體所習(xí)慣的步行時(shí)，身體上、下運(yùn)動(dòng)的頻率。它約為60～85次/分(1HZ～)，～。為了保證所運(yùn)輸貨物的完整性，車(chē)身振動(dòng)加速度也不宜過(guò)大。如果車(chē)身加速度達(dá)到1g，未經(jīng)固定的貨物就有可能離開(kāi)車(chē)廂底板。所以，～。在綜合大量資料基礎(chǔ)上，國(guó)際標(biāo)準(zhǔn)化組織ISO提出了ISO 2631《人體承受全身振動(dòng)的評(píng)價(jià)指南》。該標(biāo)準(zhǔn)用加速度均方根值(rms)給出了在中心頻率1～80HZ振動(dòng)頻率范圍內(nèi)人體對(duì)振動(dòng)反應(yīng)的三種不同的感覺(jué)界限。我國(guó)參照ISO2631制定了國(guó)家標(biāo)準(zhǔn)《汽車(chē)平順性隨機(jī)輸入行駛試驗(yàn)方法》和《客車(chē)平順性評(píng)價(jià)指標(biāo)及極限》。ISO 2631用加速度均方根值給出了人體在1～80Hz振動(dòng)頻率范圍內(nèi)對(duì)振動(dòng)反應(yīng)的三個(gè)不同感覺(jué)界限：舒適－降低界限、疲勞－工效降低界限和暴露極限。舒適－降低界限與保持舒適有關(guān)。在此極限內(nèi)，人體對(duì)所暴露的振動(dòng)環(huán)境主觀(guān)感覺(jué)良好，并能順利完成吃、讀、寫(xiě)等動(dòng)作。疲勞－工效降低界限與保持工作效率有關(guān)。當(dāng)駕駛員承受振動(dòng)在此極限內(nèi)時(shí)，能保持正常地進(jìn)行駕駛。暴露極限通常作為人體可以承受振動(dòng)量的上限。當(dāng)人體承受的振動(dòng)強(qiáng)度在這個(gè)極限之內(nèi)，將保持健康或安全。三個(gè)界限只是振動(dòng)加速度容許值不同?！氨┞稑O限”值為“疲勞－工效降低界限”的2倍(增加6dB)；“舒適－降低界限”為“疲勞工效降低界限的1/(降低10dB)；而各個(gè)界限容許加速度值隨頻率的變化趨勢(shì)完全相同。為了改善車(chē)內(nèi)乘員的舒適感，必須降低汽車(chē)行駛中的振動(dòng)，即提高汽車(chē)的行駛平順性能。汽車(chē)在一定路面上行駛時(shí)，其振動(dòng)量（振幅、振動(dòng)速度及加速度）的大小取決于汽車(chē)的質(zhì)量、懸架剛度、輪胎剛度等參數(shù)。但是，汽車(chē)振動(dòng)是一個(gè)極為復(fù)雜的空間多自由度振動(dòng)系統(tǒng)。由于車(chē)身基本不動(dòng)，所以將車(chē)身與車(chē)輪2個(gè)自由度系統(tǒng)簡(jiǎn)化圖為如圖41所示車(chē)輪部分的單質(zhì)量系統(tǒng)，來(lái)分析車(chē)輪部分在高頻共振區(qū)的振動(dòng)。圖41 汽車(chē)振動(dòng)系統(tǒng)模型根據(jù)力學(xué)定理，可列出圖41所示系統(tǒng)的振動(dòng)微分方程： （41）式中，為簧載質(zhì)量；為非簧載質(zhì)量； 為左右兩側(cè)懸架的合成剛度；為左右兩側(cè)懸架的合成當(dāng)量阻尼系數(shù)；為左右兩側(cè)懸架的合成輪胎剛度；為路面不平度賦值函數(shù)，即路面不平度對(duì)汽車(chē)的實(shí)際激勵(lì)。解式（1）可得該系統(tǒng)振動(dòng)的兩個(gè)主頻率： （42） 式中。由上式可知，汽車(chē)振動(dòng)存在兩個(gè)主頻和，它們僅為系統(tǒng)結(jié)構(gòu)參數(shù)的函數(shù)而與外界的激勵(lì)條件無(wú)關(guān)，是表征系統(tǒng)特征的固有參數(shù)。一般地說(shuō)，其中較小值的一階主頻，且接近由彈簧質(zhì)量和懸架剛度所決定的頻率，而較大值的二階主頻率，較接近主要由輪胎剛度和非簧載質(zhì)量所決定的頻率。方程的解是由自由振動(dòng)齊次方程的解與非齊次方程特解之和組成。令，則齊次方程為 式中的稱(chēng)為系統(tǒng)固有頻率，而阻尼對(duì)運(yùn)動(dòng)的影響取決于和的比值變化ζ，ζ稱(chēng)為阻尼比 ，屬于小阻尼，此時(shí)微分方程的通解為 根據(jù)上面的式子可以得到車(chē)身加速度的功率譜公式：其中（為車(chē)速）根據(jù)路面不平度分類(lèi)標(biāo)準(zhǔn)選擇G級(jí)路面，可得：=，（其中=）則=圖42 車(chē)身加速度的幅頻特性曲線(xiàn)圖也可以得到：懸架動(dòng)撓度f(wàn)d對(duì)q的幅頻特性： 將 與 代入上式，得： 式中其中為阻尼比；為剛度比；為質(zhì)量比。圖4—3 懸架動(dòng)撓度的幅頻特性曲線(xiàn)圖通過(guò)分析，當(dāng)阻尼比時(shí)，本懸架系統(tǒng)的平順性特性較好，符合ISO026311：1997 （E）標(biāo)準(zhǔn)。相對(duì)動(dòng)載Fd/G對(duì)q的幅頻特性： ，頻率響應(yīng)函數(shù) 將 代入上式，得： 式中 圖4—4相對(duì)動(dòng)載的幅頻特性曲線(xiàn)圖第5章 結(jié)論本文進(jìn)行了EQ1092中型貨車(chē)的懸架系統(tǒng)設(shè)計(jì)和平順性評(píng)價(jià)。前懸架系統(tǒng)采用鋼板彈簧和減振器的非獨(dú)立懸架，后懸架采用了主副鋼板彈簧式非獨(dú)立懸架。在前懸架系統(tǒng)設(shè)計(jì)中，對(duì)鋼板彈簧的參數(shù)進(jìn)行了確定，確定鋼板彈簧的片數(shù)為8片等厚，厚度為9mm，并確定了簧的斷面形狀；主簧的長(zhǎng)度為1270mm，用作圖法確定出各片的長(zhǎng)度。接著對(duì)鋼板彈簧的剛度和強(qiáng)度進(jìn)行了校核，使它們充分滿(mǎn)足要求。最后對(duì)減振器進(jìn)行了選擇，工作缸直徑50mm，型號(hào)選用HH型。在后懸架系統(tǒng)設(shè)計(jì)中主要對(duì)鋼板彈簧的剛度比進(jìn)行了分配并確定主副簧的各項(xiàng)參數(shù)，然后進(jìn)行校核。另外，本文還對(duì)所設(shè)計(jì)的懸架系統(tǒng)運(yùn)用時(shí)域方法進(jìn)行了平順性分析，建立了整車(chē)系統(tǒng)二自由度的線(xiàn)性動(dòng)力模型。利用MATLAB軟件進(jìn)行時(shí)域計(jì)算，根據(jù)所列微分方程得到車(chē)身加速度功率譜、相對(duì)動(dòng)載Fd/G對(duì)q的幅頻特性和懸架動(dòng)撓度f(wàn)d對(duì)q的幅頻特性，利用MATLAB軟件作出曲線(xiàn)圖。最后得出的結(jié)論為：人對(duì)該車(chē)在相應(yīng)工況下的主觀(guān)感覺(jué)為沒(méi)有不舒適。參考文獻(xiàn)[1] :人民交通出版社,2004[2] 齊志鵬．汽車(chē)懸架和轉(zhuǎn)向系統(tǒng)的結(jié)構(gòu)原理與檢修．北京：人民郵電出版社，2000[3] 張正智．中國(guó)貨車(chē)叢書(shū)．北京：北京理工大學(xué)學(xué)報(bào)，． [4] 龔為寒．：人民交通出版社， [5] 曹永堂．汽車(chē)底盤(pán)維修．北京：人民交通出版社，1999[6] 屠衛(wèi)星．汽車(chē)底盤(pán)構(gòu)造與維修．北京：人民交通出版社，．[7] :上?？茖W(xué)技術(shù)出版社,2003．[8] Yu F.,Crolla .,A State Observer Design for an Adaptive Vehicle Suspension,Vehicle Suspension Dynamic,1999[9] 汽車(chē)工程手冊(cè)，2001[10] . 北京:機(jī)械工程出版社,2002[11] :北京總后汽車(chē)試驗(yàn)場(chǎng)，.[12] ：電子工業(yè)出版社，2003.[13] ，（第四卷）.中國(guó)汽車(chē)技術(shù)研究中心標(biāo)準(zhǔn)化研究所出版社，2000[14] 整車(chē) 底盤(pán)卷（，）.長(zhǎng)春汽車(chē)研究所，[15] （第1版）.北京理工文學(xué)出版社，1999[16] 中國(guó)汽車(chē)車(chē)型手冊(cè)（上卷） 中國(guó)汽車(chē)技術(shù)研究中心 2003年 第四版附 錄Ⅰ：外文資料Comparison of Seat System Resonant Frequency Testing MethodsA seat system developed without an accurate structural dynamics model has a higher probability of squeaks, rattles, excessive seat back motion, and poor ride characteristics. If these issues are not addressed during development testing and are allowed to go into production, engineering changes are more costly and difficult to implement. Because today’s seat systems are more plex, engineers must use the latest technology to determine the seat system response characteristics.Modal analysis is the process of developing a dynamic model of a structure or a mechanical system which will be used for problem solving and trouble shooting, simulation, prediction，and optimization. The dynamic model is a set of modal parameters consisting of natural frequencies, damping factors, and mode shapes. These parameters are based on the structure or system. Experimental modal analysis can use either time based, or frequency domain based measurements to calculate the modal parameters. This method provides the most thorough definition of the dynamic response characteristics of the isolated seating system.Resonant Impact Analysis is used to determine the approximate dynamic response of a seating system. This method provides frequency response functions which describe the natural frequencies of the system. Resonant impact analysis provides information quickly, but does not define the dynamic response characteristics as pletely as modal analysis.Multiaxis shaker table testing is another tool used to determine resonant frequencies in the seat system. The shaker table is able to input sine sweep and random inputs into the seating system. The amplitude of the sine sweep or random input can be controlled in acceleration or displacement control. The shaker table is also capable of simulating road conditions of a customer’s proving grounds in the laboratory. These roads generate loads in vehicle ponents such as seats. Controlled laboratory tests allow duplication of plex multichannel time histories of a test specimen. The shaker table can reproduce road inputs in six degrees of freedom: vertical, lateral, longitudinal, pitch, roll, and yaw motions.EXPERIMENTALA correlation study of seat resonant frequencies involved the parison of seat resonant frequency data acquired by: Resonant Impact Analysis, Modal Analysis, and Shaker Table Testing using a sixaxis simulation reproducing both sinusoidal sweeps and simulated road data. All seat were installed in the OEM design position and rigidly attached to either the shaker table or modal bedplate for testing.MODAL ANALYSISModal analysis was one method used to characterize the dynamic properties of the seats. This involved collecting frequency domain measurements, more specifically frequency response functions, to describe the dynamic characteristics. An H1 estimator was used to calculate the frequency response functions of the seat systems. The seat structures were excited with two electrodynamic shakers, one mounted laterally at the top of the seat back and one mounted fore/aft at the bottom of the seat back. The response was meas