【正文】 姿態(tài)，根據(jù)路面和行駛工況自動調(diào)整懸架剛度。 Truck目 錄第1章 緒 論 1第2章 懸架系統(tǒng)的結(jié)構(gòu)與分析 3 懸架的作用和組成 3 汽車懸架的分類 3 懸架的設(shè)計要求 4 懸架主要參數(shù) 4 懸架的靜撓度fc 4 懸架的動撓度 fd 5 懸架彈性特性 5 后懸架主、副簧剛度的分配 5 懸架側(cè)傾角剛度及其在前、后軸的分配 6第3章 前后懸架系統(tǒng)的設(shè)計 7 7 7 12 1副鋼板彈簧結(jié)構(gòu)參數(shù) 13 13第4章 平順性分析和編程 15 15 15 16第5章 結(jié)論 21參考文獻 22致 謝 23附 錄Ⅰ：外文資料 24附 錄 Ⅱ：中文翻譯 31附 錄 Ⅲ： 程序 36III第1章 緒 論隨著時代的發(fā)展，以及我國汽車行業(yè)的發(fā)展，人們對貨車的舒適性和穩(wěn)定性提出了新的要求。首先確定懸架的主要結(jié)構(gòu)形式，然后對主要性能參數(shù)進行確定。 懸架由彈性元件、減振裝置和導向機構(gòu)等三部分組成。在懸架垂直載荷一定時，懸架剛度越小，固有頻率就越低，但懸架剛度越小，載荷一定時懸架垂直變形就越大。范圍內(nèi)，汽車有一定不足轉(zhuǎn)向特性，前懸架側(cè)傾角剛度應(yīng)大于后懸架側(cè)傾角剛度。獨立懸架是左，右車輪通過各自的懸架與車架（或車身）連接，當一側(cè)車輪受沖擊，其運動不直接影響到另一側(cè)車輪，獨立懸架所采用的車橋是斷開式的。故本次設(shè)計選取的汽車前后部分的車身固有頻率nn2分別為n1=,n2=2Hz 懸架的靜撓度懸架的靜撓度 是指滿載靜止時懸架上的載荷Fw與此時懸架剛度c之比，即fc=Fw/c。側(cè)傾角過大或過小都不好。后者因為要傳遞縱向力，必須設(shè)置附加的導向傳力裝置，使結(jié)構(gòu)復雜、質(zhì)量加大，所以只在極少數(shù)汽車上應(yīng)用。該圖中實線所示的葉片長度是經(jīng)過圓整后的尺寸。后鋼板彈簧由主副兩副鋼板彈簧組成。為了保證所運輸貨物的完整性，車身振動加速度也不宜過大。汽車在一定路面上行駛時，其振動量（振幅、振動速度及加速度）的大小取決于汽車的質(zhì)量、懸架剛度、輪胎剛度等參數(shù)。另外，本文還對所設(shè)計的懸架系統(tǒng)運用時域方法進行了平順性分析，建立了整車系統(tǒng)二自由度的線性動力模型。根據(jù)實驗得到的形式分析可以作為時間基礎(chǔ)，和頻率范圍基礎(chǔ)測量法來計算形式參數(shù)。響應(yīng)發(fā)生在頻率范圍為0至50赫茲的光線之間?；旌掀髌脚_同樣有模擬路面情況的能力，這些路面在汽車上產(chǎn)生的負荷就像座椅。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 measured over a frequency range of 0 to 50 Hz with 200 spectral lines. Twenty averages were taken for each FRF measurement.The excitation signal was an 80% burst random function. Burst random excitation was chosen to excite the entire frequency range of interest uniformly and allow the system response to die out prior to the end of the measurement. A burst random signal is for FFT analysis, which assumes a periodic signal, because it ensures that the signal levels are zero at the beginning and at the end of the measurement.In addition to the excitation technique, it was important to test the seat systems in a representative environment. Modal analysis can be performed in a freefree environment where the seat system is pletely suspended, or in a variety of clamped positions. The samples tested for this paper were attached to a rigid bedplate in design position to simulate the boundary conditions present when the seat systems are installed in a vehicle. The rigid bedplate is exceptionally stiff representing the optimal vehicle floor pan.Once the measurements were taken and the frequency response functions calculated, the modal parameters were estimated. The least squares plex exponential method was used for estimating the frequency and damping characteristics, while the least squares frequency domain method was used to estimate the mode shapes. This method is accurate for systems with typical damping values below 5%, as seen in the seat systems that have been tested. Tables 1 through 4 detail the modal analysis results for the samples tested. The Mode Shape is a description of seat back response, unless the base is specifically mentioned. The Frequency and Damping of each mode shape are also prov