IJREAS VOLUME 6, ISSUE 4 (April, 2016) (ISSN ) International Journal of Research in Engineering and Applied Sciences (IMPACT FACTOR 6.

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1 THICKNESS OPTIMIZATION OF CAR FRAME FOR STRENGTH Manoj Kumar Singh 1 R.Hussain Vali 2 P.Yagnasri 3 Associate Professor Assistant Professor Assistant Professor ABSTRACT Department of mechanical engineering,, Gurunanak institutions technical campus, Hyderabad, India. The objective of the present work is to model a FEM car frame and solve for strength analysis using RADIOSS solver, and to determine the torsion stiffness and bending stiffness of a stainless steel FEM car frame, isotropic type by using torsion analysis and bending analysis. Also to improve the stiffness of the car frame by changing the geometrical dimension and structural properties.two static load-cases are defined to compare the Original and the Modified model by evaluating the stiffness of the full vehicle under torsion and bending in accordance with the standards used by automotive OEM companies. A dynamic comparison is done between two models, based on global frequencies and modal shapes of the full vehicle. Design improvements are made to improve the load carrying capacity of the structure. Thickness has been modified to study as a part of the improvement. Keywords: FEM car frame, torsion analysis, bending analysis, RADIOSS. INTRODUCTION Weight reduction has become a primary concern in automotive industry. In fact, safety standards and emission regulations impose conflicting performance targets that need to be satisfied at the same time. While the respect of the safety standards pushes the automotive design process towards heavy weight solutions, environmental issues and handling call for a resolute vehicle weight reduction. Topology optimization is a non-traditional optimization technique, particularly suitable for solving structural mechanics problems at an early design stage using finite elements analysis. The aim of the study is to maximize the torsion, bending and sway modal frequencies. Deals with the topology optimization of an automotive suspension aiming at the minimization of its compliance for a given volume under three different static loading conditions. 62

2 CONSTRUCTION OF BASE MODEL The input model (i.e., Base Model) of Ferrari is designed in CATIA V5R19. Figure 1: CATIA BODY IN WHITE DRAWING FIGURE 2: ALL DIMENSIONS ARE IN MM 63

3 Figure 3: VIEWS OF THE BIW MODEL CAR INPUT DATA TO THE MODEL FOR ANALYSIS: Parameter Value Meshing Size 20 mm Material Stainless Steel Material Type Isotropic Thickness of car frame sheet 5 mm Range of Frequencies Hz Table:1 Stainless steel Properties: Young s Modulus (E) = 2.1 x 10 5 N/mm 2 Poisson Ratio (µ) = 0.3 Density (ρ) = 7.9 x 103 Kg/m3 ANALYSIS METHOLOGY: Torsion Stiffness Analysis: Fig 4: Force of 3270 N is applied at front side spring tower in different direction which shows the twisting of car frame. Fig 5: Force of 3270 N is applied at front side spring tower in same direction to check strength of car. Bending Stiffness Analysis: Force of 750 Newton s is applied at both the ends A-Pillar and B-Pillar for checking the bending stiffness which is shown in below figure. Rear end is fixed and at front side near spring towers we are fixing in three directions like X, Y and Z translation is fixed. 64

4 Normal Mode Analysis Figure 6: Force 750 N is applied at both side of A-Pillar and B-Pillar The equilibrium equation for a structure performing free vibration appears as the eigenvalue problem: where K is the stiffness matrix of the structure and M is the mass matrix. Damping is neglected. The solution of the eigenvalue problem yields n eigenvalues, where n is the number of degrees of freedom. The vector x is the eigenvector corresponding to the eigenvalue. The eigenvalue problem is solved using a matrix method called the Lanczos method. Not all eigenvalues are required -- only a small number of the lowest eigenvalues are normally calculated. The natural frequency f i follows directly from the eigenvalue. In order to run a normal modes analysis, an EIGRL bulk data entry needs to be given because it defines the number of modes to be extracted. In EIGRL card we need to specify the frequency ranges. We are giving 10 to 1000 Hz for seeing the mode shapes of CAR car frame. We are not constraining anything just simply running free-free analysis and checking the natural mode shapes. Topology optimization Topology optimization is used to the optimum material distribution within the domain given by a finite elements mesh. The aim of the project is to maximize the torsion, bending, and modal frequencies. Virtual load cases The chosen load cases for this study are somewhat standard within the car industries. Some are taken from a real scenario and other are used to measure a general stiffness. Since the natural way for the forces to travel into the body structure are through the wheels and wheel suspensions, load cases associated with ride and handling will therefore most often be applied on the spring towers and the spring suspensions attachment points. Boundary Conditions All load cases are static linear, which means that no consideration has been taken to dynamic load cases, e.g. crash impacts, NVH (Noise, vibration and harshness) etc. At the rear end, the end points are fixed in all degrees of freedom. This can be seen as the white line in Figure below. This boundary condition is applied for all the load cases. 65

5 FIGURE 7: Boundary Conditions Of The Model The mass evaluation is made by calculating the weight of the car in the original design with original material choice and the case where all car body material is stainless steel. The material & weight is shown in below. Figure 8: Original Finite Element Model And Material Data Figure 9: Hypermesh Model Car Component CALCULATION METHOLOGY: As per above figure-9 total mass of frame is equal to 4.7 E-01(Altair hypermesh units). If we convert ton to KG then the value will be 470 kg. Static comparison: To validate the proposed approach, static and dynamic indicators of the full vehicle performance are considered. These indicators are evaluated for both the original car model and the New (ie Modified) model. To assess the static behavior, the torsional and bending stiffness of the car were calculated. The body is clamped at the rear suspensions, while static vertical forces are applied at the front suspensions (points A and B). 66

6 Figure 10: Static Load Cases Of car Vehicle Figure 11: Static Load- Cases Car Vehicle Figure 12: Original FE model of the car The stiffness properties of the car were estimated for both the models by performing a static FE analysis (RADIOSS) with both models. The bending stiffness of the original vehicle model is accurately predicted by the model with the replaced with new Modified Model, while a significant discrepancy between the original and the modified models is obtained for torsion. In the latter case, the stiffness of the full vehicle body is over estimated by 10.15%, which suggests that the definition of correction factors is indeed required. The above figure shows the original finite element model of the car component, the applying force is 3270N. BASE MODEL TORSION STIFFNESS ANALYSIS Two vertical forces, one on each spring tower, are acting in opposite direction. The forces act on rigid components which are corresponding to the upper connection of the wheel suspension. The rigid components are then bolted to the spring towers. Figure 13: Force Appling The Car Vehicle Figure 14: Opposite Force Appling To The Car Vehicle The Figure above shows Front portion torsion load on spring towers to find torsion stiffness of the car component applying the force is 3270N the back clamp is fixed and the force applied to the spring towers Base Model Torsion Stiffness Analysis Results Base Model Torsion stiffness analysis is done using RADIOSS software. Figure 15: Base Model Torsion Stiffness Analysis The maximum displacement 11mm is the base model torsion stiffness analysis results. Figure 16: Base Model Torsion Stiffness Analysis The maximum stress 339MPA for the base model torsion stiffness analysis Figure 17: Base Model Torsion Stiffness Analysis The figure shows the maximum strain 0.012MPA is for the base model torsion stiffness analysis 67

7 BASE MODEL BENDING STIFFNESS ANALYSIS Car is constrained in 123 A and D Pillar Body mount. Force (750 N) in -Z is applied on Floor near L.H & R.H rocker at mid of the A-pillar and D-pillar mounts. (Total load is 1500 N) Figure 18: Base Bending Stiffness Analysis constrained at front and rear. Base Model Bending Stiffness Analysis Results (Analysis done by RADIOSS software) Figure 19: Base Model Bending Stiffness Analysis Results The maximum Displacement= 2.82mm of base model bending stiffness analysis results Figure 20: Base Model Bending Stiffness Analysis Results The maximum stress 54MPA for Base Model Bending Stiffness analysis results Figure 21: Base Model Bending Stiffness Analysis is done The maximum strain is 0.229MPA for Base Model Bending Stiffness Analysis results Based on the results of the previous static and dynamic analyses, it can be concluded that the stiffness parameters of car is safe without failing the structure. BASE MODEL NORMAL MODAL ANALYSIS: Figure 22: Base Car Model Normal Modal Analysis As shown in above figure, V1 = starting range of frequency and V2= end value of frequency range. These values we mentioned in Altair Hypermesh Software. 68

8 Frequency The above simulations are carried out to evaluate the strength of vehicle structure, by using different analysis as per Automobile OEMS manufacturer. Results as per above different load cases are show below. Base Model Normal Modal Analysis Results Dynamic comparison In order to compare the simplified and the original model in terms of dynamic behavior, the Frequencies and mode shapes of the car are estimated through an FE modal analysis (Radioss-V11) in the low-frequency range of Hz.When a normal mode analysis in Free free conditions is performed, the first six modes are rigid body modes. From mode seven onwards, the natural modes are found. The original model of the vehicle body has 10 non-rigid modes in the frequency range of interest. Only five out of them are global modes of the car, since the remaining modes involve a local deformation of the structure. The Eigen frequency values of the five global modes are used as dynamic indicators to evaluate the correlation between the original and the conceptual car model. In Table 2, the global frequencies of base model are listed. MODE SHAPES: S.No Modes(n) Frequency(Hz) S.No Modes(n) Frequency(Hz) Table 2: Base model mode shapes Frequency Vs Mode for Base Model Mode Chart 1: Frequency graph for Base Model 69

9 Figure 23: SEVENTH Mode Shape Figure 24: EIGHT Mode Shape Figure 25: NINTH Mode Shape Figure 26: TENTH Mode Shape CALCULATIONS FOR BASE RUN ANALYSIS: Calculation of the Global Torsion Stiffness The global twist angle of the chassis is a function of the vertical displacement between the two involved coil-over mounting points and their distance. K = 2.8 kn-m/deg For the bending stiffness analysis the car was constrained at the rear and front wheel locations. The load of 1.5 kn was applied at the centre of the wheel base(i.e., 0.75 kn is to be applied at the center of the wheelbase on respective side members as shown in the figure below. Figure 27 Calculation of the bending Stiffness Bending stiffness =(Total Applied load) /(Deflection at point of application of load) = 1500 / 0.35 = 4.2 kn/mm A finite element analysis was performed on the model using RADIOSS. Maximum normal stresses from the finite element analysis were compared with the acceptance criteria to verify the model. The results of Finite element analysis were observed to be lesser than that of the acceptance criteria. Acceptance Criteria Calculated Value of Base Model First flexible mode natural frequency ie 7th Mode > 24 Hz Hz Bending Stiffness > 3 kn/mm 4.2 kn/mm Torsion Stiffness > 4 kn-m/deg 2.8 kn-m/deg Total mass of frame 470 Kg Table 3: Acceptance Criteria and FEA Results MODIFICATION OF BASE MODEL: In order to meet the acceptance criteria few modifications were done to the actual base model to improve strength as well as reducing in vibration. For the modifications and analysis, the existing car frame were added with stiffness. Initially the thickness of the model, where the maximum deflection occurs in torsion analysis was increased to certain value with acceptable 70

10 limit. And one more cross beam was added at the centre of the wheel base to add stiffness to the model. Again analysis is carried out on new modified model to achieve the acceptance criteria of torsion stiffness value. Thickness is increased from 5mm to 8mm at critical location and decreased from 5mm to 4mm at strong zone. Natural frequency should also match the given acceptance criteria, so by increasing the material where ever it is critical at that location by adding thickness to the car we can achieve the frequency range to the given value. Modified CAR in Frontal Area: Modified region shown in above figure, thickness is increased in that location. Figure 28: Modified thickness to achieve torsion stiffness value 71

11 Normal Mode Analysis for Modified Body in White Model: Frequencies are mentioned below in table: Figure 29 S.No Modes(n) Frequency(Hz) Table 4: Frequency table for Modified Model The above table shows the frequency response from new modified model and also that mode shapes and its frequency (HZ) mentioned. The first 6 modes are rigid modes and less effective. After 7 th modes the frequency should be calculated and the maximum frequency came from 10 th mode ie Hz. In the Base model the maximum frequency was only ( 10 th modes ) so the frequency improve in Modified Model by 2.3 Hz. 72

12 Frequency Frequency Vs Mode for Modified Model Mode. Chart 2: Frequency Range Graph for Modified model Hence for the given criteria the frequency has been achieved, model is in safe condition and having more stiffness. Modified Car Torsion Stiffness Analysis results Figure 30: Modified torsion analysis results in Z direction displacement = 9.49mm Figure 31: Modified torsion analysis results Stress=190 MPA Figure 32: Modified torsion analysis results Strain=7.20e-04 After modification, the total mass calculated for modified model using HyperMesh M= 4.6 E- 01(Altair hypermesh units). If we convert ton to KG then the value will be 460 kg. RESULTS AND DISCUSSION: Static Comparison: First flexible mode natural frequency ie 7th Mode Acceptance Criteria Calculated Value for Base Model Calculated Value for Modified Model > 24 Hz Hz Hz Bending Stiffness > 3 kn/mm 4.2 kn/mm 4.2 kn/mm Torsion Stiffness > 4 kn-m/deg 2.8 kn-m/deg 4.25 kn-m/deg Total mass of frame 470 Kg 460 Kg Table 5: Static comparison between the Base and the Modified FE model 73

13 Frequency Dynamic Comparison: In order to compare the Base Model with the Modified model in terms of dynamic behavior, the frequencies and mode shapes of the car are estimated through an FE modal analysis (RADIOSS) in the high-frequency range of Hz. Modes(n) Base Model Frequency(Hz) Modified Model Frequency(Hz) Variations occurs Table 6: Dynamic comparison between the Base and the Modified Model in terms of global frequencies and modal shapes Comparision of Base & Modified Model Frequency Mode Chart 3: Comparision of Base Model and Modified Model Frequency 74

14 Chart 4: Base model Vs Modified model stress comparison for Torsion analysis Chart 5: Base model Vs Modified model stress curve for bending analysis In above chart, the global frequencies of both models are listed. CONCLUSION We can conclude that dynamic comparision shows that the modified model has more stiffness than full structure. It can be seen that natural frequencies are shifted upwards by an amount in the range , with the decrement of frame mass from 470kg to 460kg.The modified model correlates very well with the desired model, both in terms of static and dynamic performance. 75

15 REFERENCES [1] H. Van der Auweraer, J. Leuridan, The new paradigm of testing in todays product development process, in: Proceedings of the ISMA2004, Leuven, Belgium, September 20 22, 2004, pp [2] Wesley Linton, Analysis of Torsional Stiffness and Design Improvement Study of a Kit Car Prototype, Cranfield University, Setember [3] B. Torstenfelt, A. Klarbring, Conceptual optimal design of modular car product families using simultaneous size, shape and topology optimization, Finite Elements in Analysis and Design 43 (2007) [4] L. Long, Design-oriented translators for automotive joints, Ph.D. Thesis, Polytechnic Institute and State University, Blacksburg, Virginia, September [5] J. Pirada, S. Huang, S. Corn, C. Stawicki, X. Bohineust, Improvement of dynamic models in car industry, in: Proceedings of the 15th International Modal Analysis Conference, 1997, pp [6] G. Prater Jr., A. Shahhosseini, E. Juo, P. Mehta, V. Furman, Finite element concept models for vehicle [1] architecture assessment and optimization, in: SAE 2005 World Congress Proceedings, no , Detroit, MI, USA, [7] M. Bendsoe, N. Olhoff, O. Sigmund, Topology Optimization Theory, Methods, and Applications, second ed., Springer, Berlin, Germany, [8] L. Wang, P. Basua, J. Leiva, Automobile body reinforcement by finite element optimization, Finite Elements in Analysis and Design 40 (8) (2004) [9] here in this we can see all the material information. [10] Altair OptiStruct Manual, HyperWorks