A universal suspension test rig for electrohydraulic active and passive automotive suspension system Alexandria Engineering Journal (2017) xxx, xxx–xxx HO ST E D BY Alexandria University Alexandria En[.]
Alexandria Engineering Journal (2017) xxx, xxx–xxx H O S T E D BY Alexandria University Alexandria Engineering Journal www.elsevier.com/locate/aej www.sciencedirect.com ORIGINAL ARTICLE A universal suspension test rig for electrohydraulic active and passive automotive suspension system Mahmoud Omar *, M.M El-kassaby, Walid Abdelghaffar Mechanical Department, Faculty of Engineering, Alexandria University, Egypt Received 29 October 2016; revised 24 December 2016; accepted 16 January 2017 KEYWORDS Active suspension; Passive suspension; Servo; Hydraulic; Control; PID Abstract A fully active electro-hydraulic and passive automotive quarter car suspensions with their experimental test-rigs are designed and implemented Investigation of the active performance compared against the passive is performed experimentally and numerically utilizing SIMULINK’s Simscape library Both systems are modeled as single-degree-of-freedom in order to simplify the validation process Economic considerations were considered during the rig’s implementation The rig consists of two identical platforms fixed side by side allowing testing two independent suspensions simultaneously Position sensors for sprung and unsprung masses on both platforms are installed The road input is introduced by a cam and a roller follower mechanism driven by 1.12 kW single phase induction motor with speed reduction assembly The active hydraulic cylinder was the most viable choice due to its high power-to-weight ratio The active control is of the proportional-inte gral-differential (PID) type Though this technique is quite simple and not new, yet the emphasis of this paper is the engineering, design and implementation of the experimental setup and controller A successful validation process is performed Ride comfort significantly improved with active suspension, as shown by the results; 24.8% sprung mass vibration attenuation is achieved The details of the developed system with the analytical and experimental results are presented Ó 2017 Faculty of Engineering, Alexandria University Production and hosting by Elsevier B.V This is an open access article under the CC BY-NC-ND license (http://creativecommons.org/licenses/by-nc-nd/4.0/) Introduction The three available classifications of the suspension system (Fig 1) are passive, semi-active and active suspension systems, and this classification depends on the ability of the system to absorb, add or extract energy The passive suspension (Fig 1a) is the most commonly used due to its simplicity, robustness and low price It has limited performance because * Corresponding author E-mail address: mahmoudomar91@gmail.com (M Omar) Peer review under responsibility of Faculty of Engineering, Alexandria University its components can only store or dissipate energy and can never create energy which cannot satisfy both the comfort and handling requirements under varying road conditions Most passive suspension systems employ spring with hydraulic or pneumatic shock absorber The damping force created by shock absorbers is based on converting vibration energy into heat, then dissipating it to surroundings This leads to change in oil viscosity which influences the damping characteristics [1] Traditional automotive suspension designs have been a compromise between three conflicting criteria which are road handling, load carrying, and passenger comfort Good ride comfort requires a soft suspension but it will be sensitive to changes in applied loads Good handling requires a suspension http://dx.doi.org/10.1016/j.aej.2017.01.024 1110-0168 Ó 2017 Faculty of Engineering, Alexandria University Production and hosting by Elsevier B.V This is an open access article under the CC BY-NC-ND license (http://creativecommons.org/licenses/by-nc-nd/4.0/) Please cite this article in press as: M Omar et al., A universal suspension test rig for electrohydraulic active and passive automotive suspension system, Alexandria Eng J (2017), http://dx.doi.org/10.1016/j.aej.2017.01.024 M Omar et al Nomenclature ER DOF HILS LPT MR NG6 PDF PID PWM QC electrorheological degree of freedom hardware in the loop linear position transducer magnetorheological setting neither stiff nor soft The conventional passive suspension involves spring and damper with constant coefficients [2] Due to these conflicting demands, suspension design should compromise between these two problems as shown in Fig [3] Nowadays picking a car became such a prolonged and tiring process Cars were previously chosen according to their size and power but now as people spend a considerably long time in their cars, comfort became one of the major aspects of choosing one’s vehicle Hence, all of car manufacturers are competing in providing the utmost level of comfort by modifying their suspension systems to cope up with the road bumps and potholes Although the excitations arising from road roughness primarily affect the vehicle ride comfort, it is the input over which vehicle design engineers and vehicle drivers have the least amount of control There are three different models for potholes which are smooth, non-smooth and statistical potholes [4] Thus automotive manufacturers started to explore alternatives for the passive suspension to eliminate the above mentioned compromise and that is when the principle of active and semi-active suspensions started to be increasingly employed in high-end luxury cars as they improve comfort and stability despite their high price and power consumption [5] The semi-active (Fig 1b) was first introduced by Karnopp and Crosby in the early 1970s, [6] based on the well-known skyhook control The damping coefficient is varied by variety of methods but still the suspension system can only dissipate the road forces and can’t add additional force to the system With the right control system, the passive suspension’s compromise can be reduced resulting in a smart system making cars comfortable regardless of the road they are driven on Choi et al [7] and Yao et al [8] discussed the design and con- Figure nominal size valve pseudo derivative feedback proportional integral derivative pulse width modulation quarter car trol of the Magnetorheological (MR) dampers via several techniques while utilizing Hardware-in-the-loop-Simulation (HILS) methodology Another type of semi-active suspension utilized the Electro-rheological (ER) damper system Choi et al [9] performed field test to evaluate performance characteristics of a semi-active ER suspension system associated with skyhook controller They demonstrated that ride comfort and steering stability of the vehicle were improved Cao et al [4] showed that semi-active systems have advantages over active systems, including low power requirements, simplicity, ease of implementation and low-cost Active suspension systems (Fig 1c) employ a controllable actuator between the sprung and unsprung masses This actuator is able to both add and dissipate energy to and from the system The early studies on active suspensions performed by Hrovat [10] included numerous approaches such as modal analysis, eigenvalue assignment, model order reduction, nonlinear programming, multi-criteria optimization, and optimal control Classic control methods have also been considered, Vehicle Handling Ride Comfort Low Damping Figure > High Damping Damping compromise for passive dampers Suspension system classifications Please cite this article in press as: M Omar et al., A universal suspension test rig for electrohydraulic active and passive automotive suspension system, Alexandria Eng J (2017), http://dx.doi.org/10.1016/j.aej.2017.01.024 A universal suspension test rig for electrohydraulic active and passive automotive suspension system such as root locus, Bode diagrams, and Nichols plots Before applying any of those control techniques well-defined, linear model for the system is a necessity To design the controller, linearization of the system is a must The main obstacle for commercialization of such systems is the significant power requirement In order to reduce the cost associated with the required power, practical active suspension designs generally function as a low-bandwidth system that requires kW of peak power for road vehicle applications Chandekar et al [11] investigated the servo controlled hydraulic suspension The hydraulic pressure is supplied by a radial piston pump Body movement and vehicle level are monitored in real time by the controller which operates the hydraulic servos, mounted beside each wheel Bello et al [12] and Venkateswarulu et al [13] designed a PID control for a DOF, nonlinear, half vehicle active suspension system model A comparison was made between nonlinear passive and the developed active system which showed a better active performance The constructed model ignored nonlinearities in the hydraulic actuator as their effect was minimal and was created numerically using Matlab/Simulink Both introduced a sinusoidal road input disturbance and the developed system by Bello performed a 52.29% and 57.47% reduction in front and rear suspension deflection respectively It was concluded that regardless of the power consumption of the active system, it has better performance Bello et al [14] constructed a state space model for DOF quarter car using full statefeedback controller numerically via Simulink For step input of 0.1 m, the sprung mass acceleration and displacement of the active system has been reduced by 80% and 11% respectively compared to the passive system which shows an Figure 3 improvement in the ride comfort, also, the rattle space usage was reduced by 92.5% compared to the passive suspension system The settling time in all cases was about s Fayyad [15], Kumar et al [16] and Elattar et al [17] designed a PID controller for a QC model to improve the ride comfort and road holding ability Fayyad showed numerically that for the step input of 80 mm, the sprung mass displacement has been reduced by 25% while Kumar found that ride comfort is improved by 78.03% and suspension travel has been reduced by 71.05% with active system compared to passive one both experimentally and numerically Elattar et al compared between PID and PDF controllers and showed that although both showed improved performance, PDF has more potential Experimental setup A detailed layout of the test rig is shown in Fig with all components specified During the design stage, a reasonable factor of safety was employed while designing the tables and the camshaft to ensure durability of the test rig and its ability to withstand large hydraulic and inertial forces safely The test rig shown in Fig consists of, two testing platforms fixed side by side Each of them consists of 0.5 m steel square table 0.6 m high (1 & 10), two journal bearings (14) to hold the camshaft, a bottom base plate (6), steel guide bars (8), a spring and damper shock absorber assembly with its fixation points (21), a sprung masse divided into two plates (11,12 & 2,3), an unsprung mass (4 & 5) and upper plate to hold the assembly together (7) Both platforms are almost identical and Detailed schematic with components Please cite this article in press as: M Omar et al., A universal suspension test rig for electrohydraulic active and passive automotive suspension system, Alexandria Eng J (2017), http://dx.doi.org/10.1016/j.aej.2017.01.024 M Omar et al assembled together side by side in a way allowing one camshaft (15) to drive both platforms to ensure that the road input is identical for both suspension systems under test The assembly is lifted on wheels in case that it is needed to be transported and also equipped with round base plates for ground fixation to secure the assembly in position while operating As seen from Fig the left platform is for the passive suspension and the right one is equipped with universal fixation points to accommodate a spring and damper assembly, an active hydraulic cylinder or even both of them side by side Both sprung and unsprung mass plates are equipped with a fixation point for a linear position transducer of type (OPKON LPT200) [18] for the closed loop feedback control system The camshaft is supported by four journal bearings for uniform weight distribution, and driven by 1.12 kW induction motor (19) through a 1:7 reduction gear box and 1:5 reduction chain (17) to reduce the rpm of the motor from 1450 rpm to 41.4 rpm as a camshaft rotation speed, and this speed corresponds to a vehicle’s linear velocity of 1.56 km/h which is slow but was chosen to initially test the control system’s response A schematic of the test rig is shown in Fig indicating important sections in the design Fig shows the mechanism for road disturbance input, which is developed on the basis of vertical movement of the suspension system produced by a cam (3) and follower (4) mechanism driven by an induction motor (7) through a gearbox (2) and chains (5) for speed reduction and torque augmentation The desired input is sinusoidal which is achieved by using a circular cam 190 mm in diameter and its driving shaft is offset from its center a distance OQ = 50 mm (Fig 5) The Figure cam profile (X) can be calculated from Eq (1), where X is in mm [19] X ẳ OQ1 cos hị ¼ 50ð1 cos hÞ ð1Þ where OQ is the offset distance between the cam and rotating centers The cam had 100 mm peak to peak stroke (calculated at b = 180°) and it generated a waveform modeled as rðtÞ ¼ 100 sin xt where x ¼ 2pf and f ¼ 0:7Hz There was no need for a return spring on the follower because the heavy weight of the system, about 80 kg, above each follower is enough to always ensure surface contact between the cam and the follower The test rig allows configurations for 1DOF (Fig 6(a)) and 2DOF (Fig 6(b)) systems by removing or installing the springs representing the tire stiffness 2.1 Hydraulic power unit A hydraulic power unit is used to drive the active suspension system A schematic for it with the active suspension system is designed by Automation Studio Program as shown in Fig It consists of a 60-liter hydraulic oil reservoir (1) and a pressure compensated axial piston pump of 14 cm3/rev (2) mounted on top of the reservoir and driven by a 2.23 kW single phase induction motor (3) running at 1450 rpm The maximum supply pressure of the pump is MPa A medium pressure filter (4) is of filtrations 10 lm filters the oil supplied by the pump The pump delivery line includes a check valve Schematic layout of the test rig Please cite this article in press as: M Omar et al., A universal suspension test rig for electrohydraulic active and passive automotive suspension system, Alexandria Eng J (2017), http://dx.doi.org/10.1016/j.aej.2017.01.024 A universal suspension test rig for electrohydraulic active and passive automotive suspension system Figure 5 Cam profile and cam drive Tire Springs Rigid Columns (b) (a) Figure (a) 1DOF, (b) 2DOF (5), an accumulator (6) of volume 0.6 liter and a pressure gauge (7) The pressure line feeds through two lines, the first for the active suspension’s cylinder servo valve and the second for the road input disturbance cylinder servo valve Both lines are connected to a specially designed servo valve adapter block according to the DIN standard porting pattern to adapt a size Rexroth single stage servo valve and a size MOOG single stage servo valve 2.2 Electronic control panel The electric and electronic control panel is developed especially for the test rig including the following major components: An Arduino Mega 2560 used for processing and control of the digital and analogue I/O signals via serial communications, a universal multioutput power supply with input 220 VAC and outputs +12, 12, +5, 5, +3.3 VDC, a Rexroth and a MOOG NG6 Driver Card with integrated Signal Generator and On-Board Relay Module, circuit breakers used as protection for the main supply and the two induction motors, control relays and cooling fan The human interface is through start and stop push buttons on the front panel, two potentiometers for manual control of the two servo valves and a main switch 2.3 Testing parameters Both numerical and experimental results were performed with the same parameters which are stated in Table All parameters were calculated experimentally to be applied to the Simulink models Springs were calibrated by a hydraulic cylinder equipped with one position transducer and two pressure sensors for the piston and rod pressures, and the force was calculated from Eq (2) F ẳ Ppiston APiston ị ðProd ðAPiston Arod ÞÞ ð2Þ The distance and force were recorded for the used suspension (Fig 8) and tire helical springs (Fig 9) and the average linear trend line slope represented the stiffness For the twin-tube type dampers (Fig 11) the distance from the sensor was differentiated to get the velocity and then it was plotted against the force (Fig 10) and the slope represented the damping coefficient [20] The pump [21], Rexroth [22] and Moog [23] valves parameters were obtained from their manufacturer’s datasheets Please cite this article in press as: M Omar et al., A universal suspension test rig for electrohydraulic active and passive automotive suspension system, Alexandria Eng J (2017), http://dx.doi.org/10.1016/j.aej.2017.01.024 M Omar et al Hydraulic oil Reservoir (40 l) Variable Displacement Pump (21.17 LPM) Electrical Motor (2.27 kW) Medium Pressure Filter Check Valve Table Testing parameters Suspension parameters Sprung mass Unsprung mass Suspension spring stiffness Damping coefficient Tire stiffness Active cylinder parameters Piston diameter Piston side effective area Apiston Rod diameter Rod side effective area Arod Stroke Pump parameters Displacement Rated speed Nominal pressure Setting pressure Rexroth size servo valve Rated flow at 70 bar Input signal Supply Moog size servo valve Rated flow at 70 bar Input signal Supply Accumulator Pressure Gauge 4/3 Servo Valve Double Acting Cylinder Hydraulic circuit diagram of the hydraulic power supply unit Force (N) Figure (6) (7) (8) (9) 51 kg 14 kg 14,000 N/m 5000 N s/m 30,000 N/m 38 mm 0.001134114 m2 16 mm 0.000933053 m2 240 mm 14.8 cm3/rev 1450 rpm 70 bar 60 bar 12 l/min ±10 mA ±12 VDC 2500 2000 1500 1000 500 y = 14029x - 345.28 0.05 0.1 0.15 0.2 Displacement (m) Figure Suspension spring calibration 1500 Force (N) (1) (2) (3) (4) (5) y = 30012x - 321.51 1000 500 0 0.01 0.02 0.03 0.04 0.05 0.06 Displacement (m) Figure Tire spring calibration Numerical results 24 l/min ±40 mA ±12 VDC 3.1 DOF passive suspension A Simulink model has been created to simulate a DOF passive suspension system using the Simscape library within the Please cite this article in press as: M Omar et al., A universal suspension test rig for electrohydraulic active and passive automotive suspension system, Alexandria Eng J (2017), http://dx.doi.org/10.1016/j.aej.2017.01.024 A universal suspension test rig for electrohydraulic active and passive automotive suspension system Force (N) 1500 1000 y = 5008.1x + 1036.3 500 -0.1 -0.05 0.05 0.1 Velocity (m/s) Figure 10 Damper calibration Simulink as seen in Fig 12 The road input here is either a step input created by a fast ramp in order to limit the derivative initial overshoot or the actual cam profile of a peak amplitude of 100 mm and a frequency of 0.7 Hz and the output is viewed on the scope and recorded in the Matlab workspace for the validation process 3.2 1DOF active suspension To improve the performance, the active suspension system had to be incorporated to absorb the shock A Simulink model was created in the Simscape environment to model the active suspension system incorporating a double acting cylinder, as the Figure 11 main actuator, installed parallel to the passive suspension and driven by a servo valve as seen in Fig 13 To make the active suspension effective a control system must be incorporated to control the output signal to the servo valve according to the reference input, which in our case a reference sprung mass height from the ground to be tracked regardless of the disturbed input road signal One of the robust control systems which is utilized here is the PID controller Using the auto tuning feature of the Simulink, the system was linearized and the correct PID gains were obtained to enhance the system performance The servo valve’s parameters were optimized so that the model perfectly simulates the actual valve’s performance according to its datasheet Both the numerical active and passive sprung mass displacements against the road input were recorded (Fig 14) for the half sine wave input For a half sine wave input of 100 mm the passive suspension system reached a maximum of 101 mm Also when the input dropped to zero the suspension went down for only 0.872 mm which is almost zero The overall amplitude achieved by the passive suspension is equal to 101.87 mm which neither attenuated nor augmented the road input, because both the damping coefficient and spring stiffness are relatively high compared to the low sprung mass value of 50 kg Section view of the Twin-tube shock absorber [20] Figure 12 Simscape 1DOF passive modeling Please cite this article in press as: M Omar et al., A universal suspension test rig for electrohydraulic active and passive automotive suspension system, Alexandria Eng J (2017), http://dx.doi.org/10.1016/j.aej.2017.01.024 M Omar et al Figure 13 Simscape DOF active suspension model 120 Road input Sprung mass displacement (Acve) Sprung mass displacement (Passive) Displacement (mm) 100 80 60 40 20 0 0.5 1.5 2.5 -20 -40 Time(sec) Figure 14 Numerical active, passive and road input vs time The Active suspension had a maximum peak value of 23.6 mm, which is a 76.4% reduction in amplitude resulting in improved ride comfort However, after the peak value of the input road signal has passed the system took a dive of -36.7 mm which is much higher than the dive level of the passive system The system has an overall displacement of 60 mm which means that 40% of the bump’s input and 41% improvement from the passive suspension behavior were successfully achieved The obvious drawback of the suspension system is the dive that the sprung mass endures when the bump ends, which is justified because it is related to the hydraulic characteristics Please cite this article in press as: M Omar et al., A universal suspension test rig for electrohydraulic active and passive automotive suspension system, Alexandria Eng J (2017), http://dx.doi.org/10.1016/j.aej.2017.01.024 A universal suspension test rig for electrohydraulic active and passive automotive suspension system of the system When the bump is introduced to the system the suspension response is to retract the piston via the servo valve so pressurized oil flows into the rod side of the cylinder However, when the peak tip of the bump passes and the sprung mass starts falling, the system’s response is to extend the cylinder thus supplying the pressurized oil to the piston side of the cylinder Taking into consideration that the maximum flow of the pump Qmax is constant then the relation between extension and retraction speeds can be obtained from Eqs (3)-(5) [24] Qmax ¼ Qextension ẳ vextension Apiston ẳ 0:001134114 vextension 3ị Qmax ¼ Qretraction ¼ vretraction ðApiston Arod Þ ð4Þ ¼ 0:000933053 vretraction Figure 15 vextension ¼ 0:8227 vretraction ð5Þ which means that the cylinder needs more time to extend than to retract and while the time for the rising and falling ramps of the bump is the same then the active system dives but settles and reaches steady state after s from passing the bump Experimental results The two used position sensors are of the linear resistor type OPKON LPT200 Position sensor has been mounted on the sprung mass and the second one on the unsprung mass plate The position sensor only outputs the displacement but utilizing the Simulink resources, the signal is differentiated once for velocity and twice for acceleration if needed Experimental data acquisition for passive suspension system Figure 16 Simulink program for experimental active suspension Please cite this article in press as: M Omar et al., A universal suspension test rig for electrohydraulic active and passive automotive suspension system, Alexandria Eng J (2017), http://dx.doi.org/10.1016/j.aej.2017.01.024 10 M Omar et al 120 Road input Sprung mass displacement (Acve) Sprung mass displacement (Passive) 100 Displacement (mm) 80 60 40 20 0 0.5 1.5 2.5 -20 -40 -60 Time (sec) Figure 17 Experimental active, passive and road input vs time 150 100 PWM 50 0 0.5 1.5 2.5 -50 -100 -150 Time (sec) Figure 18 Applied PID output signal to the servo valve from the PWM pin from the Arduino 120 Sprung mass displacement Exp (Passive) Sprung mass displacement Num (Passive) 100 Sprung mass displacement Exp (Acve) Displacement (mm) 80 Sprung mass displacement Num (Acve) 60 40 20 0 0.5 1.5 2.5 -20 -40 -60 Figure 19 Time (sec) Comparison between experimental and numerical results 4.1 DOF Quarter car passive suspension system The camshaft motor is started, and the output signals from the sensors are transferred to the Simulink via the analogue input pins of the Arduino Mega 2560 through serial communication The transferred data are processed by the Simulink interface as the Arduino board serves the function of input and output interface and data acquisition card as seen in Fig 15 Please cite this article in press as: M Omar et al., A universal suspension test rig for electrohydraulic active and passive automotive suspension system, Alexandria Eng J (2017), http://dx.doi.org/10.1016/j.aej.2017.01.024 A universal suspension test rig for electrohydraulic active and passive automotive suspension system Table All obtained results Theoretical Passive Active 11 Experimental Overshoot (mm) Dive (mm) Overall Displacement (mm) Overshoot (mm) Dive (mm) Overall Displacement (mm) 101 23.6 0.872 36.7 101.87 60 95.7 33.6 0.12 41.6 95.58 75.2 4.2 DOF Quarter car electro-hydraulic active suspension system Another Simulink program was created with the aid of the Arduino integrated library (Fig 16) to read the position signal for both sprung and unsprung mass displacements and introduce a closed loop PID controller to try to keep the sprung mass’s position constant regardless of the road disturbance The control system uses the sprung mass displacement reading from the position sensor as a feedback signal to track According to the deviation from the reference position set point, the PID controller outputs a digital signal to the servo valve driver card through the PWM (Pulse width modulation) pin in the Arduino Mega 2560 board This digital signal from to 255 is equivalent to an analogue output signal from to V that is converted by the valve’s electronic driver card to ± 40 mA needed to drive the MOOG servo valve The output data obtained from the programs for the cam profile and the sprung mass displacement (Fig 17) were plotted against time It is seen that for 100 mm peak half sine road input the passive sprung mass displacement had a maximum value of 95.7 mm, while the minimum steady state value was 1.2 mm, with an overall amplitude value of 95.58 mm Those results indicate that the suspension system slightly attenuated the road input but a steady state error was produced; this could be justified as a result of the ignored friction forces between the guide bars and the sprung mass sliding on them or due to internal friction of the shock absorber assembly Those results are slightly different than the Simscape modeling; however, they are very close thus indicating that a good model validation has been performed Tuning process for the active PID controller was performed until the stable parameter combination was reached and applied It is seen that with the actual active suspension system and for the same 100 mm half sine input, the peak overshoot for the sprung mass is 33.6 mm, while the maximum recorded value of dive was -41.6 mm, thus resulting in an overall amplitude of 75.2 mm which means that 24.8% of the bump’s input and 21.3% of the passive suspension behavior were successfully attenuated The servo valve output signal generated by the PID controller through the PWM output pin from the Arduino was recorded as seen in Fig 18 There is a slight difference between the experimental and simulation results as seen in Fig 19 because the friction between the guide bars and both sprung and unsprung masses wasn’t taken into consideration during the mathematical modeling phase in order to simplify the process All the obtained results (Table 2) are tabulated to be compared in order to perform a comparison between the different experimental and theoretical setups’ performance It’s obvious that the active performance is superior to the passive in the overshoot and the peak to peak values both experimentally and numerically Although the dive in the active is worse than the passive justification has been made for that issue A total improvement of 40% was achieved theoretically; however experimentally the improvement was only 24.8%; this is due to the design variation between the actual system and the mathematical modeling The settling time for the active suspension was 2.4 s in the numerical work but only 1.4 s in the experimental results Conclusion An electro-hydraulic active suspension PID control system is designed and implemented both numerically in Simulink and experimentally on the specially designed suspension test rig A successful validation process has been performed to define all the system’s parameters and to compare the developed model against the experimental setup The obtained results show that the active suspension has a very good potential of reducing the road input disturbance of up to 24.8% for 100 mm half sine bump Future work Faster speeds should be considered to evaluate the system’s performance to input road disturbance under variable conditions The need for a larger size hydraulic pump to introduce higher oil flow rates into the system which would improve the response speed Also the regenerative hydraulic connection for the active hydraulic cylinder would be tested as it will improve the cylinder’s extension speed thus reducing the large dive that occurs with the active system One more thing would be using another form for the disturbance road input resembled in a hydraulic cylinder driven by another servo valve to achieve a more controllable custom amplitude and frequency road input References [1] Y Zhang, Xinjie Zhang, Min Zhan, Konghui Guo, F Zhao, Zongwei Liuc, Study on a novel hydraulic pumping regenerative suspension for vehicles, J Franklin Inst 352 (2) (2015) 485–499 [2] S.A Patil, S.G Joshi, Experimental analysis of DOF quartercar passive and hydraulic active suspension systems for ride comfort, Syst Sci Contr Eng.: An Open Access J (1) (2014) 621–631 [3] D.E Simon, Experimental evaluation of semiactive magnetorheological primary suspensions for heavy truck applications, Blacksburg, Virginia: 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Engineering, McGraw-Hill, New York, NY, USA, 2009 Please cite this article in press as: M Omar et al., A universal suspension test rig for electrohydraulic active and passive automotive suspension system, Alexandria Eng J (2017), http://dx.doi.org/10.1016/j.aej.2017.01.024 ... manufacturer’s datasheets Please cite this article in press as: M Omar et al., A universal suspension test rig for electrohydraulic active and passive automotive suspension system, Alexandria... and output interface and data acquisition card as seen in Fig 15 Please cite this article in press as: M Omar et al., A universal suspension test rig for electrohydraulic active and passive automotive. .. Simscape library within the Please cite this article in press as: M Omar et al., A universal suspension test rig for electrohydraulic active and passive automotive suspension system, Alexandria