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- 1. M. R. Nimbarte, L.P. Raut / International Journal of Engineering Research and Applications(IJERA) ISSN: 2248-9622 www.ijera.comVol. 3, Issue 3, May-Jun 2013, pp. 1230-12351230 | P a g eEfficiency Analysis of Hydraulic Power Steering SystemM. R. Nimbarte,1L.P. Raut21M.Tech. (CAD/CAM) Student, G.H. Raisoni College of Engg. Nagpur(MS), INDIA2Assistant Professor G.H. Raisoni College of Engg. Nagpur(MS), INDIAABSTRACTHydraulic assisted power steeringsystem is a high bandwidth servo with stringentperformance requirements on accuracy,reliability, and cost. Design of such a system canbe best achieved by using a validated and userfriendly computer simulation program.Hydraulic integrated power steering ( HIPS )program has been developed using basic conceptsfrom science and engineering. HIPS provides adesign and test environment for the integratedsteering and suspension system subjected todisturbance forces, which may be induced bypump flow oscillations and tire loads. Two real-world automotive hydraulic steering systems aresimulated with HIPS. The simulation resultsagree closely with the dynamometer test results.The application of HIPS for design optimizationis also demonstrated.I. INTRODUCTION[1] Figure 1 shows a hydraulic powersteering system, which is a high bandwidthnonlinear servo capable of generating a rack force of6,000 N for small trucks and 8,000 N for heavytrucks. About 90% of the sector shaft force comesfrom hydraulic power assist and the remaining 10%comes from driver’s effort. At one-center position ofthe steering wheel, a vane pump, which is driven bythe engine, circulates the fluid in a closed -loophydraulic circuit which includes a reservoir, vanepump with flow control and pressure relief, supplyand return lines, and the liner motion of spool valve(SV). The flow control valve which connects thedischarge and inlet manifolds of the vane pumpregulates the pump flow rate into the supply line at8.5 lpm for trucks, for engine speeds above 900rpm. When the steering wheel is turned by thedriver, SV diverts the supply line fluid flow to eitherside of the power piston for a right or left turn of thevehicle. At the same time, SV passes an equalamount of fluid flow from the other side of thepower piston through the return line to the reservoir.This increases the differential pressure acting on thepower piston and yields the desirable hydraulicpower assist force.[2]An optional speed sensitive steeringcontroller modulates the fluid flow rate drawn bythe SV from the supply line, using signals fromwheel speed sensors and wheel torque sensor. Thedriver’s steering effort is increased by lowering thehydraulic power assist during highway driving,hence improving driver’s feel of the road during cityand highway driving. The steering torque isobtained when the rack force is applied to an off-center joint on the knuckle, which turns about thekingpin at the suspension strut, including a coilspring and damper assembly.[3]The hydraulic power assists level, whichis proportional to the diverted amount of fluid flowrate from the supply line by the SV, is controlled bythe relative angle between the steering wheel andpinion angles. This valve relative angle, which isalso called valve error angle, is equal to the twistangle of the torsion bar. The torsion bar is connectedto the steering wheel intermediate shaft and thepinion at its top and bottom ends respectively.Torsion bar stiffness yields the driver’s steeringeffort, and is designed to let the driver to turn thesteering wheel with ease, and at the same time, togive a memory function for the SV in order toreduce the error angle towards zero, after the end ofthe driver’s steering demand. Since the SV is aservo valve, during on-center operation, it generatesan opposing rack force to power piston motionsubjected to tire shock and vibration loads.Figure 1 Hydraulic assisted power steeringsystem
- 2. M. R. Nimbarte, L.P. Raut / International Journal of Engineering Research and Applications(IJERA) ISSN: 2248-9622 www.ijera.comVol. 3, Issue 3, May-Jun 2013, pp.1231 | P a g e`Figure 2 Spool Valve Operating CharacteristicsFigure 2 shows a simplified schematic ofthe SV and its valve profile curves for two operatingconditions. [3]In valve operating condition 1,occurring during parking, the valve error angle islarge, and valve gain, which is proportional to theslope on the profile curve, is high. As shown on therelated valve profile curve, any oscillatory rackmotion resulting from oscillatory pump flow ratesinduces a high amplitude oscillatory pressure wavein the power cylinder of the steering gear. When thispressure wave excites the structural modes at thesupport frame, steering shudder occurs. Test resultsclearly show that the higher the slope of the valveprofile curves at the operating point, the more theintensity of shudder. In valve operating condition 2occurring during city and highway driving, anyoscillatory rack motion caused by tire shock andvibration loads will be opposed by the hydraulicforce resulting from the servo action of the rotaryspool valve. This causes the formation of oscillatorypressure waves in both power cylinders. But thevalve gain is low, and hence this pressure wave maynot have sufficient intensity to generate shudder.The steering system noise, vibration, and harshness(NVH) related problems occur from dynamicinteraction between the steering gear and suspensionsystems subjected to the disturbances, such asengine torque pulsation, tire shock and vibrationloads. These NVH problems include:1) Steering shudder; resulting from the excitation ofthe fundamental frequency at the frame support by afluid periodic flow force caused by the engine-driven vane pump.2) Steering wheel nibble; resulting from theexcitation of the fundamental frequency of the rackand pinion gear mechanism by a periodic rack forceinduced by brake disk roughness, during braking athighway speeds.3) Mechanical front-end noise; resulting from thedynamic interaction between the rack and steeringgear housing, caused by tire shock and vibrationloads generated by tires riding on the bumps, stones,and pot holes on the road during low speed driving.4) Steering wheel dither; caused by dynamicinteraction between the steering gear and frontsuspension struts subjected to a periodic vertical tireload induced by a tire high spot at a certain highwayspeed.Present solutions for steering system NVHproblems mentioned above are usually achieved byusing tuned hoses, shorter or longer hoses in thehydraulic lines, passive and active dampers, andreducing the SV gain. Since all of the abovesolutions are based on empirical rules, theiradaptation to new car platforms would be timeconsuming and expensive. But, these solutions maycause an increase in power losses, and a reduction inthe steering system bandwidth. The steering systemNVH related problems are solved best, by using avalidated and user-friendly computer simulationprogram. The hydraulic integrated power steering(HIPS) simulation program consists of two modules:Module 1 contains the steering gear model, andModule 2 contains the suspension model.Two real-world examples are studied withHIPS. Example 1 is related to the front-endmechanical noise generated when a truck is drivenat low speed on a bumpy road. Example 2 is relatedto the steering wheel dither generated when a truckwith an unbalanced tire is driven at a certain speedon the highway.II. MODELINGThe most important requirement inmodeling a system is the complete understanding ofthe performance specifications, physical andoperational characteristics of each component in thesystem. The hydraulic rack and pinion powersteering system of an automotive vehicle consists ofhydraulic fluid lines, SV, vane pump including flowcontrol and pressure relief valves, power actuator,inner and outer tie rods, lower and upper controlarms, suspension struts, front and rear rollstabilizers, disturbances such as engine torquepulsation and tire loads. A logical modularization ofthe integrated system into steering gear andsuspension component modules, and establishmentof state variables with initial conditions in eachmodule, is determined to avoid integration blow-up
- 3. M. R. Nimbarte, L.P. Raut / International Journal of Engineering Research and Applications(IJERA) ISSN: 2248-9622 www.ijera.comVol. 3, Issue 3, May-Jun 2013, pp.1232 | P a g eand achieve correct interpretation of simulationresults. This is a required method, since theintegrated model of the real-world system underconsideration consists of nonlinear components withdiscontinuous behavior. Figure 3 displays themodeling architecture of the HIPS simulationprogram.1) Steering Gear Model.a) Closed-loop hydraulic circuit; including vanepump, flow control valve with pressure relief, tunedsupply and return lines, cooler, reservoir, and SV.b) Power actuator; including the steering wheel,torsion bar, pinion gear, rack spring preload, powerpiston, tie rods, knuckle, tires, and housing.c) Driver commands for applying steering angle,steering rate, and engine rpm profiles.d) Disturbances; including pump flow rateoscillations and tire shock and vibration loads,which occur when driving on a rough terrain or withtire high spots on a highway.The steering gear model is obtained by applyingfundamental laws from fluid dynamics, heattransfer, and dynamics. The state variables aredescribed by linear and nonlinear ordinarydifferential equations, with discontinuous behaviorand temperature-dependent parameters.2) Suspension Model.a) The sprung mass frame, with three degrees offreedom; heavy, roll, and pitch.b) Four unsprung masses, each with heavy motion.c) Four suspension struts. Each strut consists of acoil spring and a nonlinear damper. Each strut isconnected between the movable end of the lowercontrol arm and a corner of the frame or the uppercontrol arm. Each control arm is pivoted about arubber bushing connected to the frame. This givesheave motion degree-of-freedom to the tire. Eachstrut is tilted with respect to vertical direction, asdefined by camber and caster angles, along thekingpin axis.d) Front-end and rear-end stabilizers; to maintainroll stability of the vehicle during cornering.Figure 3 HIPS modeling architectureIII. LITERATURE REVIEWHIPS is developed using the CSSL-IVlanguage. HIPS validation is accomplished bymeans of two real-worlds NVH related steeringproblems. Example 1 is related to mechanical front-end noise, which may occur when a car is drivenover a bump at low speed. Example 2 is related tosteering wheel dither, which may occur when truckis driven at high speed on a highway.1) Mechanical Front-End Noise:The following simulation runs have been performedduring the validation study:a) Bump Test.The Truck is driven at a speed of 8 mphand its right tire goes over a bump of 16 mm height.Figure 4 displays right-side tie rod response,denoted by FRTR, in comparison with dynamometertest data. Right side is defined with respect todriver’s right side in forward direction. Figure 5shows left-side tie rod response, denoted by FLTR,in comparison with dynamometer test data. Figure 6and 7 shows the separation gap between rack andpinion gears, denoted by XRK3MM and ZRK3MMin forward and vertical directions respectively. Themaximum amplitude and fundamental frequencyvalues agree with those of dynamometer test data asdisplayed in Figures 6 through 9. Handcomputations done on a 5 dof rigid-body model ofthe steering gear, predict an approximate value of17.8 Hz for the fundamental frequency. Themeasured value of the fundamental frequency is15.4 Hz.b) Parking Test.The steering wheel is rotated by the driverwith a sine wave command signal having anamplitude of 360 degrees and frequency of 0.25 Hz.Figures 8 and 9 show that steering wheel commandangle, denoted by ANGSW, is followed closely bypinion angle, denoted by ANGPN. In addition, thepower piston differential pressure, denoted byDPPP, has correct amplitude and frequency. DPPPwaveform does not show any sign of the steeringshudder.c) Optimization Test.The system parameters representing torsionbar stiffness, rack spring preload, and bushingstiffness at the housing supports of the steering gearhave been changed, one at a time, andcorresponding simulation runs were carried out toobtain responses for XRK3MM and ZRK3MM.The main goal of this study was to reduce theamplitude of above responses, without degrading thesteering feel response, which is represented by theminimum error angle between ANGSW andANGPN responses shown in Figure 8.
- 4. M. R. Nimbarte, L.P. Raut / International Journal of Engineering Research and Applications(IJERA) ISSN: 2248-9622 www.ijera.comVol. 3, Issue 3, May-Jun 2013, pp.1233 | P a g e2) Steering Wheel Dither.The hydraulic rack and pinion powersteering system of a small truck, with 16 inch tiresand independent front and solid-axle rearsuspensions, has been studied using HIPS. Thefollowing results have been obtained:a) Transient Response.A small truck is driven at 80 mph on thehighway. The right front tire, assumed to have asingle high spot, is applying a periodic displacementwith amplitude of 2.85 mm and frequency of 14.2Hz to the suspension strut. The steering wheel is notheld by the driver and is free to rotate. Figure 10shows transient response for the torsion bar torque,denoted by TTB. Figure 11 shows transient responsefor the hydraulic power assist pressure, denoted byDPPP. It can be observed that a single tire high spotduring highway drives at a speed of 80 mph, yieldssteering dither at a beat frequency of 1.5 Hz andcarrier frequency of 14 Hz. This is shown by handcalculations as follows:1. Simplify the steering gear model and determinethe fundamental frequency, computed to be 11 Hzfor the steering wheel, which is free to rotate, and9.25 Hz for the steering wheel, which is held by thedriver. An average fundamental frequency of 10 Hzis assumed.2. Simplify the rigid body motion of the un sprungmass consisting of the tire-wheel assembly, coilspring and damper of the strut. The fundamentalfrequency of the simplified system is computed tobe 9.4 Hz, assuming that the friction is negligible.3. From rigid body dynamics, it is known thatvibrations with beat frequencies may be generated ifthe forcing frequency and the average fundamentalfrequency of system differ slightly. For example, letthe average fundamental frequency of the system be10 Hz. Since the forcing frequency is about 14 Hz,then the carrier and beat frequencies are computedto be 12 Hz and2 Hz respectively.Figure 4 Right Side Tie Rod force responseFigure 5 Left Side Tie Rod force responseFigure 6 Transverse Gear Separation ResponsesFigure 7 Vertical Gear Separation Responses
- 5. M. R. Nimbarte, L.P. Raut / International Journal of Engineering Research and Applications(IJERA) ISSN: 2248-9622 www.ijera.comVol. 3, Issue 3, May-Jun 2013, pp.1234 | P a g eFigure 8 Steering Feel ResponsesFigure 9 Steering power assist pressureResponsesFigure 10 Torsion Bar Torque TransientResponseFigure 11 Hydraulic Power Assist PressureTransient Responseb) Frequency Response.1. This task has been achieved in the followingsteps:Perform transient response for each of the followingcases defined by tire high spot frequency or vehiclespeed:Vehicle Speed Tire High SpotFrequency50 8.860 10.670 12.580 14.290 16.02. For a system variable of interest, such as TTB,compute the Fourier sine and cosine coefficientsover a period, after transients die away and thesteady state conditions are attained, during thetransient response simulation run for each of theabove cases.3. Obtain the root mean square (rms) value of TTBTTB for each tire high spot frequency or vehiclehighway speed. Figure 10 shows the frequencyresponse with a resonance peak at 12.5 Hz.4. This obtains the power spectrum density (psd)value of TTB by squaring its rms value and dividingit by a frequency window for each of the cases inStep 3. The frequency window is selected to betwenty times smaller than the highest forcingfrequency, which is 16 Hz in this case. Figure 10shows the TTB psd frequency spectrum, whichshows a maximum power level at 12.5 Hz. Thiscorresponds to a highway speed of 70 mph.c) Optimization Test.The torsion bar stiffness parameter,denoted by KTB, was changed from 1.37 to 1.13Nm/deg, and a simulation run performed at a vehiclespeed of 80 mph had shown an anticipated reduction
- 6. M. R. Nimbarte, L.P. Raut / International Journal of Engineering Research and Applications(IJERA) ISSN: 2248-9622 www.ijera.comVol. 3, Issue 3, May-Jun 2013, pp.1235 | P a g ein dither torque at the steering wheel. Since such achange will also affect driver’s steering feel, it isnecessary to make additional simulation runs, asshown in Figures 10 and 11, and easily quantify theeffects of such a change.IV. CONCLUSIONSHydraulic power steering system designoptimization can be achieved in a cost-effectivemanner by means of a computer modeling andsimulation program including user-friendliness,modularity, flexibility, and a proven track record inthe solution of NVH related problems.The angular velocity of sector shaft applies thetangential and radial forces on teeth. The vehicletyre applies the resistance to the linear movement ofthe piston.The teeth are subjected to bending andwear stresses due to above forces. So to have propersteering gear ratio both bending and wear strengthof the meshing teeth should be increased. Toincrease the strength the finite element analysis willbe carried out to find the stresses in the teeth andsuggesting the solution.REFERENCES[1] CSSL-IV User Guide and ReferenceManual, Simulation Services, Chatsworth,CA[2] Sahinkaya, Y., A Novel Steering VibrationStabilizer, 1996 SAE International Congressand Exposition, February, Detroit, MI[3] Bosch: Kraftfahrtechnisches Taschenbuch,Publisher: Robert Bosch GmbH,Stuttgart,Vieweg Verlag, 23rd printing, Wiesbaden,Germany, 1999[4] Brander, M.:Modellerstellung undSimulation einer wirkungsgradoptimiertenhydraulischen Fahrzeuglenkung. Diploma-Thesis, Chair of Turbomachinery andFluidpower, Technical University ofDarmstadt, Darmstadt, 2001[5] Dominke, P.: Die Lenkung im Wandel –Von der Mechanik zur MechatronikPresentation at the Fahrzeug- undMotorentechnisches Seminar, Chair ofAutomotive Engineering (fzd), TechnicalUniversity of Darmstadt, Darmstadt,Germany.[6] Ivantysynova, M.: Ways For EfficiencyImprovements Of Modern DisplacementMachines.TheSixth ScandinavianInternational Conference on FluidPower,SICFP99, May 26-28, Tampere,Finland, 1999[7] Lang,A.:AlternativeEnergiesparendeServolenksysteme.Hydraulik undPneumatik in der Fahrzeugtechnik, Hausder Technik, Essen, Germany, Feb. 1995[8] Murrenhoff, H., Wallentowitz,H.:Grundlagen der Fluidtechnik, Teil 1:Hydraulik Verlag Mainz, Aachen, 1stprinting, Germany, 1998[9] Stoll, H.: Fahrwerktechnik: Lenkanlagenund Hilfskraftlenkungen . Publisher: J.Reimpell, Vogel Buchverlag Würzburg, 1stprinting, Würzburg, Germany, 1992[10] ZF integral Hydraulic Ball & Nut powersteering Gear-Short Version: By ZFSteering Gear (India) Ltd.[11] Integral Hydraulic Ball & Nut Powersteering Gears Spare part List: By ZFSteering Gear (India) Ltd.[12] Service Training Handouts: By Rane TRWSteering Systems Ltd.[13] Machine Design: By V B Bhandari[14] Automobile Engineering: By Ttti, Bhopal,Jain, K K, Asthana, R B[15] Trends in Automobile Engineering: By A.S. Rangwala (Author)[16] Handbook of automotive power electronicsand motor drives By: Ali Emadi

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