Upcoming SlideShare
×

# Analysis of Automobile Suspension

2,302 views

Published on

test prod 8-25

1 Like
Statistics
Notes
• Full Name
Comment goes here.

Are you sure you want to Yes No
• Be the first to comment

Views
Total views
2,302
On SlideShare
0
From Embeds
0
Number of Embeds
3
Actions
Shares
0
142
0
Likes
1
Embeds 0
No embeds

No notes for slide

### Analysis of Automobile Suspension

1. 1. Analysis of an Automobile Suspension by Derek Maxim Hieu Nguyen Ryan Parent Eric Twiest School of Engineering Grand Valley State University EGR 350 – Vibrations Section A Instructor: Dr. Ali Mohammazadeh August 4, 2006
2. 2. Introduction Modeling the suspension of an automobile is of interest for many automotive andvibrations engineers. Of importance for these engineers are the ride quality of the vehicletraversing over broken roads and control of body motion. When traveling over rough terrain, thevehicle exhibits bounce (up and down), pitch (rotation about the center of gravity along thevehicles length) and roll (rotation about the center of gravity along the vehicles width) motions.For this project, the bounce and pitch motion of the car over rough roads are of interest and willbe analyzed in this report.Assumptions For the analysis, it will be assumed that the vehicle is a rigid body with a suspension thatwill be modeled as a two-degree-of-freedom (DOF) system. The setup of the suspension willconsist of equivalent springs in which the stiffness of the tire and the spring are combined, andequivalent dampers that account for the shock absorber and the damping of the tire.Theory Figure 1 shows the two DOF system schematic that was used to determine the equationsof motion of the vehicle. Figure 1: Spring-mass-damper model of the vehicle
3. 3. To determine the equations of motion, Lagranges equations, also known as the energymethod, were utilized. Equation (1) shows the general form of Lagranges equations (1)where L is the sum of the kinetic and potential energies, or (2)where T is the kinetic energy, U is the potential energy of the system. The terms qi and Qi fromEq. (1) represents a degree of freedom and the non-conservative work for each DOF (subscript idenoting the first and second degrees of freedom); represents the derivative of qi. To derive the equations of motion using Lagrange, the degrees of freedom i needs to bedefined. This is shown in Eqns. (3) and (4). (3) (4)Next, the kinetic energy of the system is shown in Eq. (5), (5)where M is the mass of the body, is the bounce velocity of the body about its center ofgravity, J is the polar moment of inertia, and is the angular acceleration of the body. The potential energy of the system is shown in Eq. (6) (6)where k1 and k2 are the equivalent spring rates of the front and rear suspension, xCG is thedisplacement of the bodys center of gravity, l1 and l2 are the distances from the center of gravity
4. 4. to the front suspension and rear suspensions, and y1 and y2 are the input functions of the road forthe front and rear of the system. Combining Eqs. (5) & (6) produces the energy equation, Eq. (7) (7) The equations for non-conservative work for both degrees of freedom are shown in Eqs.(8) & ( 9) (8) (9)where Q1 and Q2 are non-conservative work for q1 and q2, c1 and c2 are the damping coefficientsof the system and and are the time derivatives of the road input function. Finally, taking the derivatives of the q terms and combining all of the equations into theform of Eq. (1), the equations of motion for the system are (10) (11) The parameters of the system are as follows: k1 = k2 = 30000 N/m, c1 = c2 = 3000 N*s/m,M = 2000 kg, J = 2500 kg*m2, l1 = 1 m, and l2 = 1.5 m. Substituting these values and expandingEqs. (10) & (11) yields Eqs. (12) & (13) (12) (13) The car is traveling at 13.88 m/s over road that is assumed to be sinusoidal in cross-
5. 5. section with an amplitude of 10 millimeters (0.01 meters) and having a wavelength of 5 meters.With this information, the input functions y1 and y2 are defined in Eqs. (14) & (15) (14) (15)Where, t is the time traveled and π is the time shift that accounts for the time that it takes for therear suspension to negotiate the "bump" that the front suspension had negotiated.ResultsSIMULINK The system was simulated using MATLABs SIMULINK program. Figure 2 shows theschematic that was used for analysis. Figure 2: SIMULINK model of the two-degree-of-freedom systemThe schematic shown in Figure 2 was used to determine the natural frequencies ω1 and ω2 of thesystem. Using MATLAB, the modes of vibration, which are due to the system possessing twodifferent natural frequencies, were calculated to determine ω1 and ω2 in SIMULINK. FromMATLAB, the first and second modes of vibration were 0.477 and -0.596 (see MATLAB
6. 6. results). Figures 3 and 4 show the plots produced by SIMULINK, which contains the naturalfrequencies, and verify the MATLAB results. From Figures 3 and 4, the natural frequencieswere determined from the "Power Spectral Density" plots (middle graphs) and were ω1 = 5.1 rad/s and ω2 = 6.5 rad/s.Figure 3: SIMULINK plot results for the first mode of vibration showing the bounce (left graph) and pitch response (right graph); Power Spectral Density graph used to determine natural frequency ω1 Figure 4: SIMULINK plot results for the second mode of vibration showing bounce (left plot) and pitch (right plot) response where natural frequency ω2 can be determined from the Power Spectral Density graph In addition, SIMULINK was used to model the response of the system to the roadconditions. Once road conditions were modeled, the SIMULINK model was modified using a
7. 7. slider gain to reduce the pitch motion of the vehicle. Figures 5 and 6 show the response of thesystem under the given car parameters and Figures 7 and 8 show the response when the gains onthe dampers in the system were modified to achieve the most desirable results. ComparingFigures 5 and 6 to Figures 7 and 8 the figures, it was easy to see that by increasing damping gainby a factor of 10, pitching motion decreases from 5x10-4 meters to less than 1x10-4 meter.Bounce motion also decreases from 3x10-3 meters to 1x10-3 meters. Figure 5: Bounce (left) and pitch motion (right) plot results for the unadjusted modeling of the system under original conditions Figure 6: SIMULINK model used to determine the response of the system
8. 8. Figure 7: Bounce (left) and pitch motion (right) response plot results for the system with higher viscosity (increased gain) dampersFigure 8: SIMULINK model with slider gain block included to reduce the pitching motion of the systemMATLAB MATLAB, a mathematical processing software, was used to compare and verify themodel analyzed in SIMULINK. The program was also used to compare the responses of thesystem using a function known as "lsim" and modal analysis. Attached at the end of this reportare the codes used to run lsim and the modal analysis. Before the analysis of the system was performed using the lsim function, the modes and
9. 9. natural frequencies of the system were determined. Figure 9 shows the plot of the modesproduced in MATLAB. From Figure 9, mode 1 is seen to have an oscillation of lower amplitudethan mode 2, which has an oscillation of higher amplitude. Using modal analysis, thedisplacement degrees of freedom due to natural frequencies ω1 and ω2 were u1 = [-0.0197,0.0094] meters and u2 = [0.0105, 0.0176] meters. From these results, it can be concluded thatmode 2 has a greater effect on the system than mode 1.Figure 9: Plot of the modes of the system; mode 1 is shown to have an oscillation with a smaller frequency than mode 2 To use the lsim function in MATLAB. To convert the equations into transfer functions,the equations themselves must undergo a Laplace transformation. The generic equation for thetransfer function is shown in Eq. (16), whereas the specific transfer functions of the system, afterundergoing the Laplace transformation, are shown in Eqs. (17)-(20) (see Appendix A forderivation of these equations). (16) (17) (18) (19)
10. 10. (20)With these transfer functions entered into MATLAB, the frequency response plots of the bounceand pitching motion were created and are shown in Figures 10 and 11.Figure 10: Bounce motion plot resulting from the analysis of the system using the lsim function Figure 11: Pitching motion plot of the system resulting for the use of the lsim functionComparing Figures 10 and 11 to Figure 5 (SIMULINK plot of the system), it can be seen thatboth models show similar behavior to the road input, with small differences in amplitude. Theresponse of the front and rear suspensions to the road using lsim analysis are shown in Figures12-13 and Figures 14-15. Figure 12 shows the front suspension response to bounce, Figure 13
11. 11. shows the pitching response of the same suspension, Figure 14 shows the rear suspensionresponse to bounce, and Figure 15 shows the pitching motion response. Figure 12: Front suspension response to bounce using the lsim function Figure 13: Pitching response of the front suspension to the road input Figure 14: Bounce response of the rear suspension to road input
12. 12. Figure 15: Pitching response of the rear suspension to road input Modal analysis was performed using MATLAB to compare the response of the system tothe lsim analysis and the matrices needed to perform the analysis can be seen in Appendix A.However, it was not completed at the time of writing, so it cannot be proved in this report thatthe response from the use of the lsim function is similar to the response resulting from modalanalysis. It is expected that the results would be similar, assuming that the matrices included inthis report from modal analysis were correct and the parameters and input functions weretransformed correctly.Conclusions Using MATLAB to model the suspension system (albeit simplified two-degree-offreedom compared to a system that can be modeled with as much as ten degrees of freedom), itwas found that the suspension with front and rear spring rates of 30,000 Newton per meter, frontand rear dampers of a rate of 3,000 Newton-second per meter for a 2,000-kg vehicle quells theexcitation produced by the road in approximately 1.5 seconds. The second mode of vibrationwas found to contribute the bounce and pitch motion of the vehicle more than the first mode ofvibration. The response of the system using modal analysis was also performed to verify theresponse of the system. Though the eigenvalues and eigenvectors were determined usingMATLAB, unfortunately, the response of the system from the analysis was incomplete at thetime of writing. SIMULINK was also used to model the suspension system and it was found to be withinagreement with the MATLAB model. Using the slider gain to increase or decrease the dampingrate on the SIMULINK model, it was found that by increasing the damping gain (and therefore
13. 13. damping rate), the bounce and pitch motions of the vehicle decreased by a factor ofapproximately 5 and 3, respectively.