新型微型拖拉机外观及主要部件结构设计含开题及9张CAD图
收藏
资源目录
压缩包内文档预览：
编号：91865860
类型：共享资源
大小：12.24MB
格式：ZIP
上传时间：2020-08-10
上传人：QQ14****9609
认证信息
个人认证
郭**（实名认证）
陕西
IP属地：陕西
35
积分
- 关 键 词：
-
新型
微型
拖拉机
外观
主要
部件
结构设计
开题
CAD
- 资源描述：
-
新型微型拖拉机外观及主要部件结构设计含开题及9张CAD图,新型,微型,拖拉机,外观,主要,部件,结构设计,开题,CAD
- 内容简介：
-
Finite Elements in Analysis and Design 45 (2009) 456-462Contents lists available at ScienceDirectFiniteElementsinAnalysisandDesignjournal homepage: /locate/finelSimplified modelling of joints and beam-like structures for BIW optimization in aconceptphaseofthevehicledesignprocessD. Mundoa, R. Hadjitb, S. Dondersb, M. Brughmansb, P. Masb, W. DesmetcaDepartment of Mechanical Engineering, University of Calabria, 87036 Arcavacata di Rende, ItalybLMS International, Interleuvenlaan 68, B-3001 Leuven, BelgiumcDepartment of Mechanical Engineering, Katholieke Universiteit Leuven, Division PMA, B-3001, Leuven, BelgiumA R T I C L EI N F OA B S T R A C TArticle history:Received 3 March 2008Received in revised form 3 December 2008Accepted 10 December 2008Available online 7 February 2009Keywords:BeamJointConceptual designNVHVehicle bodyThe paper proposes an engineering approach for the replacement of beam-like structures and joints ina vehicle model. The final goal is to provide the designer with an effective methodology for creating aconcept model of such automotive components, so that an NVH optimization of the body in white (BIW)can be performed at the earliest phases of the vehicle design process. The proposed replacement method-ology is based on the reduced beam and joint modelling approach, which involves a geometric analysis ofbeam-member cross-sections and a static analysis of joints. The first analysis aims at identifying the beamcenter nodes and computing the equivalent beam properties. The second analysis produces a simplifiedmodel of a joint that connects three or more beam-members through a static reduction of the detailedjoint FE model.In order to validate the proposed approach, an industrial case-study is presented, where beams and jointsof the upper region of a vehicles BIW are replaced by simplified models. Two static load-cases are definedto compare the original and the simplified model by evaluating the stiffness of the full vehicle undertorsion and bending in accordance with the standards used by automotive original equipment manufac-turer (OEM) companies. A dynamic comparison between the two models, based on global frequenciesand modal shapes of the full vehicle, is presented as well. 2009 Elsevier B.V. All rights reserved.1. IntroductionIn highly competitive markets, design engineers face the chal-lenging problem of developing products, which must fulfil complexand even conflicting design criteria. In the field of automotive indus-try, the task of improving various functional performance attributes,such as safety, noise and vibrations, ecological impact etc., is mademore and more difficult by the necessity of launching new productsor renewing existing models in an increasingly short time frame. Inorder to make the complexity of the design criteria compatible withthe necessity of reducing the time-to-market, predictive computer-aided engineering (CAE) methods must be already available in theearly phases of the design process.Traditional computer-aided design (CAD) software packages havea very limited applicability in early design stages, since they requiredetailed data of the vehicle. Besides, they are based on the traditionaldefinition of geometry via points, lines and surfaces, thus makingCorresponding author. Fax: +39984494673.E-mail address: d.mundounical.it (D. Mundo).0168-874X/$-see front matter2009 Elsevier B.V. All rights reserved.doi:10.1016/j.finel.2008.12.003the parameterization of models difficult and time-consuming 1. Asa result, the experience of engineers is a key factor for the selectionof proper structural concepts at the beginning of the design process.Recently, research efforts have been spent to enable designers touse CAE as a support in the conceptual phase of the design process,when functional performance targets are defined, while detailed ge-ometrical data are still unavailable. The objective is to improve theinitial CAD design, hence shortening the design cycle 25.In the field of NVH and crashworthiness prediction, several con-cept modelling approaches have been proposed by researchers. Theycan be classified into three categories: methods based on predecessorFE models, methods from scratch, and methods concurrent with CAD.Methods belonging to the first category, which includes meshmorphing and concept modification approaches 68, are used to de-sign a variant or incremental improvement of an existing vehiclemodel. By using a predecessor FE model, early CAE predictions canbe performed to identify issues and to include possible countermea-sures already in the initial CAD design.If a new car concept is to be designed and a predecessor FE modelis not available, methods “from scratch” can be used to support thedesign process during the early design phases. Two classes of meth-ods are distinguished. The first class is topology design optimization,D. Mundo et al. / Finite Elements in Analysis and Design 45 (2009) 456-462457where material is eliminated from an initial admissible designdomain in order to make the structure lighter without violatingfunctional requirements 911. Performance based on the opti-mized topological information is usually improved by optimizingshape and size. The second class of methods “from scratch”, knownas functional layout design, aims at building a simplified conceptmodel, consisting of beams, joints and panels, which represents thefunctional layout and which is used to predict the performance of themodel 12.Methods concurrent with CAD are CAE tools available in an earlyphase of the design process. These methods provide simulation re-sults as soon as component-level CAD models are available, whilevehicle-level models are still unavailable 13.Among the methods based on predecessor FE models, the“reduced beam and joint modelling” approach has been recentlyproposed by Donders et al. 14 to improve the fundamental NVHbehavior of a vehicle BIW. The proposed approach creates a reducedmodal model at the beam center nodes, to which beam elementsand joint superelements can be added, thus enabling a conceptmodification of the body and an accurate prediction of dynamicNVH performance. The commercial software program LMS VirtualLab. 15 includes a user-friendly implementation of the reducedbeam and joint modelling approach. Design engineers can define abeam and joint layout, calculate the body reduced modal model andperform efficient design modification and optimization of the bodybeam-like sections and joint connections.In this paper, the reduced beam and joint modelling approach isemployed to replace beams and joints of the predecessor FE modelwith concept models. After identifying the beam center nodes asthe geometric center of the cross-sections, the equivalent beamproperties are calculated through a geometric analysis and appliedto simplified beam elements that connect the beam center nodes.The stiffness parameters of thin-walled beams, as computed bymeans of a geometric approach, need a correction that takes intoaccount section variations and discontinuities (holes, spot-welds,stiffeners) 16,17. For this purpose, proper correction factors aredefined and estimated for each beam-member by means of aniterative model updating procedure. In a next step, a simplifiedmodel of joints, connecting two or more beam-members, is thenobtained through a static reduction of the detailed FE model of thejoint.In order to validate the proposed approach, a case-study is pre-sented, in which beams and joints of the upper region of a vehicleBIW are replaced by simplified models. A static comparison betweenthe original and the simplified model is performed by evaluating thestatic stiffness of the full FE vehicle BIW under torsion and bend-ing. A dynamic comparison between the two models, based on theglobal frequencies and mode shapes of the full vehicle, is performedas well.2. The reduced beam and joint modelling approachThe reduced beam and joint modelling approach is proposed byDonders et al. 14 for efficient modification of beams and jointsof a vehicle, based on the reduced modal model of the nominalvehicle. The basic idea is to identify the so-called beam center nodes,and to create a reduced modal model at these beam center nodes.Subsequently, the mass and stiffness properties of the structure aremodified by connecting the beam center nodes through simple beamelements and joint superelements. In this paper, simplified beamand joint models are created to completely replace the original FEmodel (without the necessity of the reduced modal model), so thatan optimization of the vehicle can be performed in the early phaseof conceptual design, when a detailed model of the structure is notyet available.yizixixB.C.N.zyFig. 1. Schematic representation of a beam end-section.In this section, an overview of the procedure that is used to es-timate the mass and stiffness properties of the simplified beam andjoint models is provided.2.1. Beam property estimationBeam-like members, i.e. structures for which the dimension inthe longitudinal direction is much larger than the characteristic di-mension of the cross-sectional area, are the primary structural ele-ments in a BIW. They strongly influence the natural frequencies ofthe vehicle body.In the FE model of a vehicle, beam-like members are typicallythin-walled structures, formed by shell elements.In order to replace the detailed mesh of such components by sim-plified beam elements, a number of beam cross-sections are consid-ered and the equivalent beam properties are computed for each ofthem. For this purpose the following procedure is implemented:(1) a cut node is selected in the region of the beam-member wherean intersection plane is to be applied,(2) an axis system that defines the approximate beam direction andintersection plane is defined,(3) the primary members shell elements along the intersectionplane are cut and analyzed to locate the beam center node inthe geometric center of the original cross-section,(4) the following equivalent beam properties w.r.t. the beam centernode are computed: A: cross-section area; Ixx: torsional moment of inertia; Iyy, Izz: moments of inertia of area; and Iyz: product of inertia of area.Here, x denotes the beam direction, and the yz plane is theintersection plane, as shown in Fig. 1. For an arbitrary cross-section, the calculation of the properties can be implementedby computing the equivalent beam properties for each shellelement that belongs to the cross-section, according to thelocal principal axes (xi, yi, zi). Then, a transformation fromthe local axis system to that of the intersection plane (x, y, z)is performed. Finally, a summation over all shell elements isperformed to find the global properties for that cross-section.(5) the beam center node is connected to the surrounding mesh bymeans of interpolation relations (Nastran superelements RBE3).These relations are defined between each beam center node anda particular node group, formed by all nodes of the shell ele-ments that are defined at the intersection plane at the consid-ered cross-section.Typically, along each primary beam-member a number of intersec-tion planes are defined, for which equivalent beam properties arecomputed. The entire beam member can then be represented as aseries of linear beam elements taken from a standard FE library. Anexample is shown in Fig. 2, where both the original detailed and thesimplified FE model of B-pillars of a vehicle BIW are represented.458D. Mundo et al. / Finite Elements in Analysis and Design 45 (2009) 456-462Fig. 2. (a) Original and (b) conceptual models of a BIW B-pillars.Fig. 3. Original FE model of a joint group, extracted from the vehicle model forstatic reduction.2.2. Joint property estimationComplementary to the simplified beam modelling approach de-scribed in Section 2.1, a procedure for simplifying joints connectingbeam-like structural members in a vehicle body is proposed. Afterevaluating the equivalent beam properties of all beam-members con-nected by the joint, a joint group is created that includes the inter-polation elements to the beam center nodes at the joint ends 15. InFig. 3 an example is shown, in which the mesh of the joint that con-nects the left B-pillar of the vehicle to the roof-rails is extracted fromthe rest of the vehicle body. For this isolated joint model, Guyan re-duction is used to calculate a small-sized representation of the joint.Guyan reduction 18, also known as static condensation, is amethod to reduce the finite element stiffness and mass matricesof structures. For an arbitrary structure, the basic static FE matrixequation is given byK x = F(1)where K is the stiffness matrix, F and x are the force and the dis-placement vectors, respectively. By identifying ntboundary degreesof freedom (DOFs), which must be retained in the solution, and nointernal DOFs, which are to be removed by static condensation, thesystem of Eq. (1) can be partitioned as follows:?KooKotKtoKtt?xoxt?=?FoFt?(2)where subscripts t and o are used to designate the boundary and theinternal DOFs, respectively. From the first line of Eq. (2), the internaldisplacement vector can be determined asxo= K1oo(Fo Kot xt)(3)By introducing the static reduction matrix Got=K1ooKotand substi-tuting Eq. (3) into the second line of Eq. (2), the following equationis obtained:Ktt,red xt= Ft,red(4)where Ft,red=Ft+GTotFois the reduced loading vector, while Ktt,red=KtoGot+ Kttis the ntx ntreduced stiffness matrix. Physically thismatrix represents the stiffness values between each pair of boundaryDOFs. This way, the stiffness of the structure has been condensed tothe boundary DOFs.The same transformation can be used to condense the mass ma-trix on the boundary DOFs, to obtain a reduced system also for dy-namic analyses. However, while exact for the stiffness matrix, theGuyan reduction is an approximation for the mass matrix. By re-ducing the mass matrix, it is assumed for the considered structurethat inertia forces on internal DOFs are less important than elasticforces transmitted by the boundary DOFs. This is true for very stiffcomponents or in cases where local dynamic effects can be ignored.Therefore, the accuracy of the result is case dependent.For each isolated joint model, a Guyan reduction is performed,with the DOF of the joints end nodes (i.e. beam center nodes) asthe boundary DOFs to be retained in the solution. The FE modelof the joint is thus reduced to a small superelement, consisting ofa reduced stiffness and mass matrix. For typical automotive joints,the stiffness relations between the end points of the joint have amuch stronger influence on the global body behavior than the exactdistribution of mass on the joint. For this reason, Guyan reductionof the joint structure to its joint end-nodes (i.e. beam center nodes)seems an appropriate choice to create a small-sized representationof the actual joint 14.D. Mundo et al. / Finite Elements in Analysis and Design 45 (2009) 456-4624593. Case-study3.1. Model descriptionFig. 4 shows an industrial BIW model, consisting of 123 panelsthat are modelled with linear shell elements. The constituent panelsare assembled by means of about 3000 spot-weld connections 19,which are represented in the FE model by means of Hexa solid ele-ments15.Inordertovalidatethereducedbeamandjointmodellingapproach, as described in the previous section, a group of beam-likestructures, labelled in Fig. 4 as B1.B5, are selected and replaced byequivalent simple beams.In total, 10 beams are selected for the replacement, namely theA and B-pillars and the longitudinal and transversal roof-rails. Fourjoints, symmetrically arranged w.r.t. the longitudinal plane of thevehicle, connecting these beams are labelled in Fig. 4 as J1, J2, J3and J4,are statically reduced. Fig. 5 shows the simplified BIW model,where the detailed shell models of the beam-like structures havebeen replaced by simple two-node beam elements. The numberand length have been selected based on the geometric characteris-tics (i.e. length and cross-section variations) of the original mesh.The original FE joint models have been removed from the BIW FEmodel, and the joints have been represented by static superelements(i.e., the equivalent mass and stiffness matrices of each joint).3.2. Static comparisonTo validate the proposed approach, static and dynamic indica-tors of the full vehicle performance are considered. These indicatorsare evaluated for both the original BIW model and the simplified(or conceptual) model. To assess the static behavior, the torsionaland bending stiffness of the BIW are calculated. The body is clampedat the rear suspensions, while static vertical forces are applied at theFig. 4. Original FE model of the BIW.Fig. 5. Conceptual FE model of the BIW. The original meshes of 10 beam-membersand four joints are replaced by simplified beam elements and joint superelements.Fig. 6. Static load-cases defined to estimate the BIW stiffness under (a) torsion and(b) bending.front suspensions (points A and B in Fig. 6). Based on the estimationof the vertical displacements vAand vBat the excitation points, thebending and torsion deflection angles?band?tare determined as?b= arctan?vA+ vB2L?(5)?t= arctan?vA vBW?(6)where L and W denote the wheelbase and the width of the car,respectively, measured at the front suspension points.Based on the torsional deflection?t, the torsional stiffness Ktisdetermined asKt=M?t(7)where M = F W is the moment applied at the front suspension,resulting from two oppositely oriented forces F.Similarly, the bending stiffness Kbis determined from?basKb=2FL?b(8)where F is the vertical force applied at the frontal suspension loca-tion.The stiffness properties of the BIW are estimated for both themodels in Figs. 4 and 5, by performing a static FE analysis (Nastran-Sol 101) with both models. In Table 1, the torsional and bendingstiffness indicators are listed, as well as the approximation involvedby the simplified model w.r.t. the original model. The results showthat the bending stiffness of the original vehicle model is accuratelypredicted by the model with the replaced simplified beams andjoints, while a significant discrepancy between the original and the460D. Mundo et al. / Finite Elements in Analysis and Design 45 (2009) 456-462simplified models is obtained for torsion. In the latter case, the stiff-ness of the full vehicle body is overestimated by 10.15%, which sug-gests that the definition of correction factors is indeed required. Thiswill be evaluated and assessed in Section 4.3.3. Dynamic comparisonIn order to compare the simplified and the original model in termsof dynamic behavior, the frequencies and mode shapes of the BIWare estimated through an FE modal analysis (Nastran-Sol 103) in thelow-frequency range of 050Hz. When a normal mode analysis inTable 1Torsional and bending stiffness of the original and of the conceptual FE model.TorsionBendingOriginal modelConcept modelOriginal modelConcept modelStiffness(Nm/rad)1.456E+051.603E+055.013E+045.036E+04?(%)10.150.45Table 2Dynamic comparison between the original and the conceptual FE model in termsof global frequencies and modal shapes.Mode n Modal shapeFrequency (Hz)MACiiOriginal modelConcept model?(%)11st Torsion18.2219.285.820.9922nd Torsion26.1327.886.700.9683Lateral bending39.3640.011.650.9894Vertical bending41.7342.120.930.9895Mixed torsion+bending47.8547.920.150.9910.80.60.40.20Id 7-47.9Id 5-42.0Id 4-40.0Id 2-27.9Id 1-19.3Id 1-18.2Id 3-26.1Id 7-39.4Id 8-41.7Id 10-47.9Discrete valuesMode Set.2 - uncorrectedDiscrete valuesMode Set.1 - Original model 10.90.80.70.60.50.40.30.20.1I12for all beams(12)Fig. 9. History of the goal function as defined in Eq. (13).Table 3Optimal values of the correction factors for bending stiffness parameters.Correction factorsB1B2B3B4B5Ixx0.2600.1570.3110.0070.001Iyy0.9330.3910.7710.0790.011Izz0.7610.4550.8930.1920.219Ixy0.8090.3710.8120.1130.001where the inequality constraint ensures that the actual set of cor-rection factors defines a feasible solution.An optimal and feasible solution can be searched for by means ofa genetic algorithm 22. For this purpose, the constrained problemdefined above is transformed into an unconstrained problem by us-ing a penalty formulation: a large cost-value is added to the objec-tive function in case that the constraint is violated. Such a procedureensures that an unfeasible solution has a larger goal function thanany feasible solution. This enables the convergence of the algorithmtowards a global optimum, which fulfils all constraints. The originalconstrained optimization problem is then replaced by the followingunconstrained problem:Minimizef(c1, .,cN) + k(c1, .,cN) C(13)where C is a penalty cost-value, while k(c1, .,cN) is a Boolean func-tion defined ask(c1, .,cN) =?0if I1 I2I121otherwise(14)The optimization problem is solved by a genetic algorithm imple-mented in Matlab. In Fig. 9 the history of the goal function definedby Eq. (13) is shown.Table 3 lists the optimal values of the correction factors evalu-ated for both bending stiffness-parameters of all the replaced beam-members.The geometric approach used to compute the equivalent beamparameters considers each cross-section as closed, even for the roofcross members. This is the main reason why correction factors arequite low (significantly lower than one), especially for these mem-bers. Roof cross members, in fact, are formed by two panels con-nected to each other by glue connections, which are much more462D. Mundo et al. / Finite Elements in Analysis and Design 45 (2009) 456-462Table 4Torsional and bending stiffness of the original and the final conceptual FE model.TorsionBendingOriginalmodelFinal conceptmodelOriginalmodelFinal conceptmodelStiffness(Nm/rad)1.456E+051.462E+055.013E+045.024E+04?(%)0.460.22Table 5Dynamic comparison between the original and the final conceptual FE model interms of global frequencies and modal shapes.Mode nFrequency (Hz)MACiiOriginal modelFinal concept model?(%)118.2218.220.010.99226.1326.090.150.998339.3639.310.120.998441.7341.990.620.978547.8547.520.680.984flexible than spot-weld connections. Therefore, the stiffness overes-timation involved by the geometric approach is much bigger thanfor the other beam-members.Finally, Tables 4 and 5 provide a comparison between the originalmodel and the simplified model, corrected by the optimal factors,in terms of static stiffness, natural frequencies and modal shapes.The results show that the final conceptual model correlates verywell with the detailed model, both in terms of static and dynamicperformance.5. ConclusionAn engineering approach for the replacement of beam-like struc-tures and joints in a vehicle model has been presented in this paper.In order to validate the proposed approach, a case-study has beendefined, where A-pillars, B-pillars and roof-rails of a vehicles BIWhave been replaced by equivalent beam models. Four joints, con-necting the above-mentioned beam-like structures to each other,have been replaced as well through a static reduction of the detailedmesh. Two static load-cases have been defined to apply torsion andbending to the full vehicle and compare the original and the simpli-fied models. The stiffness of the full vehicle under the two loadingconditions has been evaluated and a difference of +0.46% and +0.22%between the simplified and the original models has been obtainedfor the torsion and bending load-case, respectively. A dynamic com-parison between the two models, based on the first 10 frequenciesand modal shapes of the full vehicle, has been performed as well.The dynamic behavior of the full vehicle is accurately predicted bythe simplified model. More specifically, in the comparison of the twomodels, a maximum eigenfrequency difference of 0.68% and a MACvalue difference of 2.2% have been obtained. The quantitative resultsdescribed above are well in line with OEM requirements for conceptmodification predictions in an early design stage.In summary, a proof-of-concept of the feasibility of a stand-alonebeam and joint replacement layout has been realized, which enablesan accurate approximation of the global static and dynamic charac-teristics. The stiffness correction factors can be derived in a singleoptimization procedure (updating w.r.t. the predecessor model) ina replacement scenario. Naturally, once the replacement model hasbeen established, a fast concept optimization is easily achievable.Individual beam properties can be modified, which is not possible(or easily achievable) on the complex-shaped cross-sections of theactual shell mesh.AcknowledgmentThe work presented in this paper has been performed inthe framework of the research project “Analysis Leads Design-Frontloading Digital Functional Performance Engineering”, which issupported by I.W.T. Vlaanderen.References1 K. Volz, Car body design in the concept stage of vehicle development, in:Proceedings of the Second European LS-DYNA Conference, Gothenburg, Sweden,June 1415, 1999.2 H. Van der Auweraer, J. Leuridan, The new paradigm of testing in todays productdevelopment process, in: Proceedings of the ISMA2004, Leuven, Belgium,September 2022, 2004, pp. 11511170.3 H. Shiozaki, Y. Kamada, S. Kurita, S. Goossens, J. Van Herbruggen, V. Cibrario,L. Poppelaars, CAE based vehicle development to reduce development time, in:Proceedings of the JSAE Annual Congress, no. 20, Yokohama, Japan, 2005.4 R. Hadjit, M. Brughmans, H. Shiozaki, Application of fast body optimizationprocedures to shorten car development cycles, in: Proceedings of the JSAEAnnual Congress, no. 18, Yokohama, Japan, 2005.5 B. Torstenfelt, A. Klarbring, Conceptual optimal design of modular car productfamilies using simultaneous size, shape and topology optimization, FiniteElements i
- 温馨提示:
1: 本站所有资源如无特殊说明，都需要本地电脑安装OFFICE2007和PDF阅读器。图纸软件为CAD,CAXA,PROE,UG,SolidWorks等.压缩文件请下载最新的WinRAR软件解压。
2: 本站的文档不包含任何第三方提供的附件图纸等，如果需要附件，请联系上传者。文件的所有权益归上传用户所有。
3.本站RAR压缩包中若带图纸，网页内容里面会有图纸预览，若没有图纸预览就没有图纸。
4. 未经权益所有人同意不得将文件中的内容挪作商业或盈利用途。
5. 人人文库网仅提供信息存储空间，仅对用户上传内容的表现方式做保护处理，对用户上传分享的文档内容本身不做任何修改或编辑，并不能对任何下载内容负责。
6. 下载文件中如有侵权或不适当内容，请与我们联系，我们立即纠正。
7. 本站不保证下载资源的准确性、安全性和完整性, 同时也不承担用户因使用这些下载资源对自己和他人造成任何形式的伤害或损失。
人人文库网所有资源均是用户自行上传分享，仅供网友学习交流，未经上传用户书面授权，请勿作他用。
川公网安备: 51019002004831号