Experimental Evaluation of Magneto rheological Damper for Passive on-off State

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1 ISSN Experimental Evaluation of Magneto rheological Damper for Passive on-off State #1 A.P Aher, #2 M.W.Trikand, #3 V.V Jagirdar, #4 R.R Kharde 1 aherabhijeet02@gmail.com 2 Mukund.trikande@gmail.com 3 vinit.jagirdar@gmail.com 4 R_R73300@yahoo.co.in #14 Pravara Rural Engineering College, Loni, India #23 Vehicles Research and Development Establishment, Ahmednagar, India ABSTRACT In this paper, a method of estimating equivalent damping coefficients of magnetorheological (MR) damper is presented. At first, a MR damper characterization is carried out. Performance testing is done for this damper with servo hydraulic structural testing machine and equivalent damping coefficients at different input current levels of MR damper are estimated from work diagrams. Estimated equivalent damping coefficients are used for simulation of quarter car model. Finally, a real quarter car model is set up including MR damper by replacing conventional damper. Simulation results of quarter car model built up in simulink validated well with experimental results and shows good conformity of estimated equivalent damping coefficients.evaluation of MR damper is considered in this study for ride and comfort. Keywords Characterization, Displacement transmissibility, Damper, Quarter car model, Ride and comfort ARTICLE INFO Article History Received :18 th November 2015 Received in revised form : 19 th November 2015 Accepted : 21 st November, 2015 Published online : 22 nd November 2015 I. INTRODUCTION Vibration control of vehicle suspension systems has been a very active subject of research, since it is concerned with ride comfort for drivers and passengers. For a long time, efforts were done to make the suspension system works by optimizing the parameters of the suspension system, but for intrinsic limitation of passive suspension system the improvement is effective only in a certain frequency range. Compared with passive suspensions, active suspensions can improve the performance of the suspension system over a wide range of frequency. Semi-active suspensions were proposed in the early 1970s which can be nearly as effective as fully active suspensions in improving ride quality. When the control system fails, the semi-active suspension can still work in passive condition. Compared with active and passive suspension systems, the semi-active suspension system combines the advantages of both active and passive suspensions; i.e. it provides good performance compared with passive suspensions and is economical, safe and does not require either higher-power actuators or a large power supply. According to G.Z Yao, G.Chen, S.H.Yeo(2002) in the last few years a new development has arisen in the form of dampers taking advantage of the electro rheological or magneto-rheological property of liquids. Both principles are based on the alteration of the damping medium s viscosity depending on an applied electric or magnetic field. [1]. Solepatil.S.B, Awadhani L V. (2014) explains the effects of viscosity on transmissibility of magnetorheological damper. They also focused on MR fluid compositions, magnetizable particles, carrier fluids and additives [2].According to K. Hudda and H. Jamaluddin (2005) accurate modeling is crucial for suspension analysis and design. In many practical applications the damper characteristic exhibits a strong nonlinearity, which must be taken into account in simulation studies in order to obtain realistic results when investigating system performance [3]. A. Giua, M. Melas and X. Song (2004, 2009) discuss the unique characteristic of a magnetorheological damper. The unique characteristic is called damper constraint, where the force-velocity 2015, IERJ All Rights Reserved Page 1

2 relationship of the damper lies only in the first and third quadrants of Cartesian coordinate [4, 5]. K. Senthilkumar (2009) discuss the use of force displacement diagram also called work diagram is measure of energy absorb by damper during bump and rebound of road load. The area inside force-velocity is measure of power consumed by damper which can be delivered in the form of heat [6]. M. T. Braz- César, R. C. Barros explains complete experimental program for characterization of magnetorheological damper [7]. G.Z Yao and F.F Yap (2002) investigated that the MR damper has a very broad changeable damping force range under magnetic field and the damping coefficient increases with the electric current, but decreases with excitation amplitude [8]. S. Kciuk and R. Turczyn, (2010) gives brief description about the constructional feature of MR damper [9]. Wolf R. Krüger and Ondřej Vaculín gives brief discussion over active, semi-active and passive damping techniques for ground vehicles [10].According to W H Liao, C Y Lai,(2002) an equivalent damping coefficient of damper can be determined by equating the energy dissipated in a full working cycle. When the displacement is kept fixed an increase in frequency leads to a decrease in damping coefficient. The amount of reduction is larger for higher voltage or current input. For lower excitation frequencies, a smaller velocity will result as the same excitation amplitude. Therefore, on average, the damping coefficient of the MR damper working at higher frequencies is smaller [11]. Chen Da-wei and GU Hong-bin (2010) proposed the expression of MR damper equivalent damping coefficient with voltage and excitation amplitude [12].Mahmoud El-Kafafy and Samir M. El-Demerdash (2012) investigated the performance of automotive ride comfort using Bouc-Wen type magneto-rheological (MR) fluid damper and studied using a two degree of freedom quarter car model [13]. Abdolvahab Agharkakli and Rijumon K(2012,2013) gives brief overview on mathematical modeling for passive and active suspension systems as well as control strategies for semi-active suspension system for quarter car model [14, 15].P.S. Els, N.J. Theron,(2007) highlights the ISO standards of ride comfort and handling for off road vehicles[16]. In this paper, a MR damper is characterized first, and then equivalent damping coefficients are estimated from work diagrams. Quarter car model simulation results are compared with those of experimental quarter car model results. Through this analysis, MR damper is evaluated for ride and comfort criteria. II.MR DAMPER CHARACTERIZATION To apply the MR damper in vibration control of vehicle suspension system, the property of the damper should be determined first. An experimental test rig is set up to determine the property of the MR damper and to obtain the dynamic data necessary for estimating the equivalent damping coefficients. In this test rig, the MR damper is fixed on a computer-controlled servo-hydraulic actuator. This actuator incorporates a load cell and a displacement sensor to measure the force produced by the MR damper and the displacement of the piston. TABLE I SINUSOIDAL EXCITATION PARAMETERS FOR RD-8041 MR DAMPER Parameter Values Frequencies (Hz) 0.4, 1.2, 2.0,2.8,3.6 Amplitude (mm) 10 Current supplies (Amp) 0.0, 0.1,0.2, 0.3,0.4,0.5 The sinusoidal type of excitations with frequencies from 0.4 Hz to 3.6 Hz and the increment of 0.8 Hz and the fixed 10 mm amplitude of excitation are used. The applied electric current is from 0 to 0.5 A with increment of 0.1 A. The force and displacement responses of the damper are sampled simultaneously by the computer via an A/D converter. MR damper Load cell Hydraulic actuator Fig.1 Test rig for characterization The excitation signal is produced by the computer and sent out to the hydraulic actuator via MCU controller of test machine. Velocity response has been obtained by differentiating the displacement. A. Characterization Results The responses of MR damper at 1.2 Hz excitation under five constant electric currents are shown in Figure below. The effect of magnetic field on the damping force is clearly shown in these figures. With the increasing of the applied electric current, the damping force will increase remarkably. Fig.2 Force Vs Velocity Response (1.2 Hz, 10mm, variable current) 2015, IERJ All Rights Reserved Page 2

3 Fig.3 Force Vs Displacement Response (1.2 Hz, 10mm, variable current) An equivalent damping coefficient is determined by equating the energy dissipated in a full cycle. Area of work diagram can be estimated by using poly-area command in Mat-lab. Let, the energy dissipated by the MR damper in one cycle be = (1) Where ω d is the driving frequency of the sinusoidal excitation is the relative velocity of the damper and F MR is the measured damping force. Assuming a simple harmonic excitation, x (t) = X sin ω d t, where X is the amplitude of the relative motion of the damper, = Therefore can be found as = (2) Fig.4 Equivalent damping coefficients Vs. Frequency The equivalent damping coefficients of the damper against forcing frequency under various electric currents are shown in Fig. 4 It is seen that at low frequency, equivalent damping coefficient will increase markedly. As the frequency increases, the equivalent damping coefficient under high electric current decreases rapidly whereas that without electric current decreases slowly. At high frequency, the effect of current on equivalent damping coefficient is also not so significant. This phenomenon means that the MR damper cannot be treated as a viscous damper under high electric current. The values of equivalent damping coefficient at intermediate frequencies can be estimated by using interpolation method. B. Modeling of system To simulate the performance of vehicle subjected to sinusoidal road surface, the passive quarter car model as shown in Fig.5. Fig.5 Passive quarter car model The equations of motion for this linear model is M S Z S K S Z S Z u C Z S zu M u Zu K S Z S Zu C Z S Zu k z Z g Where Zu Z s Z s z u C. Quarter car model simulation The MATLAB/Simulink model for passive suspension system with MR damper is prepared and the sprung mass displacement and acceleration are for sinusoidal road excitations are obtained for suspension parameters given in TABLE 2 SUSPENSION PARAMETERS OF QUARTER CAR MODEL System Parameter Sprung mass(ms) of quarter car analysis Unsprung mass acceleration Sprung mass acceleration Unsprung Mass (Mu) Suspension Stiffness (Ks) Tire Stiffness (Kt) Sprung mass velocity Unprung mass velocity Damping coefficient(cs) Value 190 Kg 20 Kg N/m 200 KN/m t Variable u (3) D. Experi mental setup (4) 2015, IERJ All Rights Reserved Page 3

4 Frequency (Hz) Simulink Sprung mass disp (mm) 0A 0.1A 0.3A 0.5A Fig.6 Instrumentation scheme for QCM The quarter-car rig was designed as a physical representation of a classic two-degree of freedom (2DOF) system. The quarter car test rig consists of two LVDT mounted on the sprung and unsprung masses to measure displacements and one accelerometer on sprung mass for measuring acceleration..a servo-hydraulic actuator is used to provide necessary input excitation.the proposed sytem is tested with harmonic excitations with 10 mm amplitude. The sensor readings are processed in signal conditoning circuit and fed to National Instruments cdaq chasis 9174 where cdaq 9234 is the input module.computer based virtual instrumentaion control using National instruments LABVIEW is used to test the system.a constant DC current of various magnitude is supplied to MR damper using DC power supply. III EXPERIMENTAL AND SIMULATION RESULTS A. Sprung mass displacements Sprung mass displacements and accelerations are measured by both experimental and simulink at frequencies ranges from 0.4 Hz to 3.6 Hz with incremental of 0.4 Hz at passive off state i.e. 0 amp current and passive on state ie current input with increment of 0.1 amp up to 0.5 amp. Fig.7 Sprung mass displacement at 2.4 Hz, 0 amp Fig.8 (a) Disp. Transmissibility Vs Frequency, 0 amp Fig 8(b) Sprung mass disp. Vs Frequency TABLE 3 EXPERIMENTAL SPRUNG MASS DISPLACEMENT RESULTS Frequency (Hz) Experimental Sprung mass disp (mm) 0A 0.1A 0.3A 0.5A Fig.9 Sprung mass displacement at 2.4 Hz, 0.1 amp TABLE 4 SIMULINK SPRUNG MASS DISPLACEMENT RESULTS 2015, IERJ All Rights Reserved Page 4

5 Fig.10 (a) Disp. Transmissibility Vs Frequency, 0.1 amp Fig. 12(b) Sprung mass disp. Vs Frequency,0.3 amp Fig.10 (b) Sprung mass disp. Vs Frequency, 0.1 amp Fig13. Sprung mass displacement at 2.4 Hz, 0.5 amp Fig11. Sprung mass displacement at 2.4 Hz, 0.3 amp Fig 14 (a) Disp. Transmissibility Vs Frequency, 0.5 amp Fig.12 (a) Disp. Transmissibility Vs Frequency, 0.3 amp Fig.14 (b) Sprung mass disp. Vs Frequency,0.5 amp B. Sprung mass acceleration results Sprung mass acceleration plots are obtained for both experimentally and simulation at each current input ranging from 0 amp to 0.5 amp and frequencies from 2 Hz to 3.6 Hz. 2015, IERJ All Rights Reserved Page 5

6 Then RMS accelerations are obtained from plots and compared. TABLE 4 EXPERIMENTAL SPRUNG MASS RMS ACCELERATION RESULTS Experimental Frequency (Hz) Sprung mass RMS accl (m/s 2 ) 0A 0.1A 0.2A 0.3 A Fig 16 RMS acceleration Vs Frequency at 0 amp TABLE 5 SIMULINK SPRUNG MASS RMS ACCELERATION RESULTS Simulink Frequency (Hz) Sprung mass RMS accl (m/s 2 ) 0A 0.1A 0.2A 0.3 A Fig 17 RMS acceleration Vs Frequency at 0.1 amp Fig.14 Sprung mass acceleration at 2.4 Hz, 0 amp Fig 15 Sprung mass acceleration at 2.4 Hz, 0.2 amp Fig 18 RMS acceleration Vs Frequency at 0.2 amp IV. CONCLUSIONS The equivalent damping coefficients of the damper against forcing frequency under various electric currents shows that at lower frequencies, equivalent damping coefficient will increase markedly. As the frequencies increases, the equivalent damping coefficient under high electric current decreases drastically whereas that without electric current decreases slowly. At higher frequencies, the effect of current on equivalent damping coefficient is also not so significant. This phenomenon means that the MR damper cannot be treated as a viscous damper under high electric current..an equivalent damping coefficients of the MR damper in terms of fixed displacement, input current, and frequency are investigated. From the experimental and simulation results of displacement transmissibility using a quarter car test rig, it can be noted that estimated equivalent damping coefficients of MR damper using work diagrams 2015, IERJ All Rights Reserved Page 6

7 shows good conformity with experimental results above 2 Hz forcing frequency. The experimentally and simulation measured displacement transmissibility for several passive cases are shows that there is shift in the resonant frequency due to the increased coupling of sprung and unsprung masses at higher damper currents. Both simulation and experimental results of sprung mass RMS acceleration shows that, damper input current above 0.2 amp is unacceptable from ride and comfort point of view. ACKNOWLEDGEMENT The authors are highly indebted to Dr. Manmohan Singh, Director, VRDE Ahmednagar, for his invaluable support and encouragement during the execution of this work. The authors express their sincere thanks to Mr. Sujithkumar, Sc C for his guidance and keen interest shown during the entire period of execution of this work. The authors also acknowledge the efforts put by Mr.Swanand Kulkarni, Testing engineer, VRDE during experimental measurements. REFERENCES [1] G.Z. Yao, F.F. Yap, G. Chen, W.H. Li, S.H. Yeo. MR damper and its application for semi- active control of vehicle suspension system Mechatronics 12 (2002) [2] Solepatil.S.B, Awadhani L.V. Effect of Viscosity on the Transmissibility of a Magnetorheological Damper International Journal of Science and Research 2014, ISSN [3] K. Hudha, H. Jamaluddin, Effects of control techniques and damper constraint on the performance of a semi-active magnetorheological damper Int. J. Vehicle Autonomous Systems, Vol. 3, Nos. 2/3/4, 2005 [4] Giua, M. Melas, C. Seatzu, G. Usai, Design of a predictive semi active suspension system, Vehicle System Dynamics, Vol. 41, No. 4, pp , Apr [5] Xubin Song, Cost-Effective Skyhook Control for Semi active Vehicle Suspension Applications The Open Mechanical Engineering Journal, 2009, 3, [6] Senthilkumar, K, Manish M. Analysis of Characteristics of Dampers of Hydrogas Suspension and the Effect of Damping Configuration on the Vibration Dynamics of a Light Tracked Vehicle SAE Paper No [7] M. T. Braz-César, R. C. Barros, Experimental behavior and numerical analysis of MR dampers LISBOA [8] Yao, G. Z., Yap, F. F., Chen, G., Li, W. H. & Yeo, S. H. MR damper and its application for semiactive control of vehicle suspension system Mechatronics, 12 (7), , (2002). [9] S. Kciuk, R. Turczyn, M. Kciuk, Experimental and numerical studies of MR damper with prototype magnetorheological fluid Journal of Achievements in Materials and Manufacturing Engineering 39/1 (2010) [10] Wolf R. Krüger, Ondřej Vaculin Evaluation of Active Damping for Reduction of Noise, Vibration and Motion of Ground Vehicles by Multibody Simulation RTO-MP-AVT-110, October 2004 [11] W H Liao, C Y Lai, Harmonic analysis of a magnetorheological damper for vibration control Smart Materials and. Structures. 11 (2002) [12] Chen Da-wei, GU Hong-bin, Application of magneto-rheological (MR) damper in landing gear shimmy IEEE,2010. [13] Mahmoud El-Kafafy Samir M. El-Demerdash, Automotive Ride Comfort Control Using MR Fluid Damper journal of scientific research and Engineering, 2012, 4, [14] Abdolvahab Agharkakli, Ghobad Shafiei Sabet, Simulation and Analysis of Passive and Active Suspension System Using Quarter Car Model for Different Road Profile International Journal of Engineering Trends and Technology- Volume [15] Rijumon K, Murtaza M A, A comparison between passive & Semi active suspension systems, International Journal of Innovative Research in Science, Engineering and Technology Vol. 2, Issue 6, June 2013 [16] P.S. Els, N.J. Theron, The ride comfort vs. handling compromise for off-road vehicles Journal of Terramechanics 44 (2007) , IERJ All Rights Reserved Page 7