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A seat suspension with a rotary magnetorheological damper for heavy duty vehicles

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2016 Smart Mater. Struct. 25 105032

(http://iopscience.iop.org/0964-1726/25/10/105032)

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A seat suspension with a rotarymagnetorheological damper for heavy dutyvehicles

S S Sun1, D H Ning2, J Yang1, H Du2, S W Zhang3 and W H Li1

1 School of Mechanical, Materials and Mechatronic Engineering, University of Wollongong, New SouthWales 2522, Australia2 School of Electrical, Computer & Telecommunications Engineering, University of Wollongong, NewSouth Wales 2522, Australia3Department of Precision Machinery and Precision Instrumentation, University of Science and Technologyof China, Hefei, Anhui province 230026, People’s Republic of China

E-mail: weihuali@uow.edu.au and swzhang@ustc.edu.cn

Received 29 May 2016, revised 21 July 2016Accepted for publication 18 August 2016Published 26 September 2016

AbstractThis paper presents the development of an innovative seat suspension working with a rotarymagnetorheological (MR) fluid damper. Compared with a conventional linear MR damper, thewell-designed rotary MR damper possesses several advantages such as usage reduction ofmagnetorheological fluid, low sealing requirements and lower costs. This research starts with theintroduction of the seat suspension structure and the damper design, followed by the property testof the seat suspension using an MTS machine. The field-dependent property, amplitude-dependent performance, and the frequency-dependent performance of the new seat suspensionare measured and evaluated. This research puts emphasis on the evaluation of the vibrationreduction capability of the rotary MR damper by using both simulation and experimentalmethods. Fuzzy logic is chosen to control the rotary MR damper in real time and two differentinput signals are considered as vibration excitations. The experimental results show that therotary MR damper under fuzzy logic control is effective in reducing the vibrations.

Keywords: seat suspension, rotary MRF damper, vibration control

(Some figures may appear in colour only in the online journal)

1. Introduction

Prolonged periods of time spent in a vibrating machine causesthe hazard known as whole-body vibration or WBV. Vibra-tion transferred from uneven road surfaces, vibrating tools,and vibrating machinery to a vehicle driver’s body sig-nificantly influences drivers’ comfort and safety. Long termexposure to high levels of WBV can also cause physical painin the driver’s neck/shoulders, lower back injuries, spinalinjuries, and other musculoskeletal diseases, not only whiledriving but during their off time as well, and is becoming amajor problem [1]. The WBV problem is much more seriousin the mining industry, construction field and agriculturebecause rough road conditions can cause fierce vibration tovehicles or heavy machines (bulldozers, shearers, etc) and the

suspension of heavy machines and trucks used in the aboveareas must stiffen enough to carry their weight. The hardsuspension, however, makes the vibrations caused by unevenroads and the operating actions of heavy vehicles easilytransmitted to the body and this causes the drivers toexperience an uncomfortable ride and ultimately, suffer healthproblems. Long term exposure to WBV can cause workerspermanent health problems such as numbing of nerves, gutproblems, spinal injuries and damage to a woman’s ability toconceive [2, 3].

The aim of this paper is to investigate ways of protectingthe workers from such unhealthy and unsafe working routinesby arming the heavy machinery and trucks with advanced seatsuspensions. Seat suspension has been adopted in vehicles,particularly used for commercial industry, mining,

Smart Materials and Structures

Smart Mater. Struct. 25 (2016) 105032 (10pp) doi:10.1088/0964-1726/25/10/105032

0964-1726/16/105032+10$33.00 © 2016 IOP Publishing Ltd Printed in the UK1

agriculture, and transport purposes [4, 5], to provide the dri-vers with ride comfort, to reduce the fatigue due to long hoursof driving or exposure to severe working environments, andto improve the safety and health standards. The optimisationand control of seat suspensions for vertical vibration reduc-tion has been an active research topic for decades. To date,passive [6], active [7–9] and semi-active [3, 10, 11] seatsuspensions have been proposed. A passive seat suspension issimple, reliable, and cost-effective but it cannot provide acontrollable force and consequently, performance is inevi-tably limited. While maintaining the geometric and dynamicproperties of a passive suspension structure, an active or semi-active device has been considered for incorporation in modernseat suspension structures to meet the requirements. In part-icular, semi-active seat suspensions offer performance com-parable to that brought by active seats without requiring highpower consumption and expensive hardware. In the pastdecades, magnetorheological fluid (MRF) has been the pre-ferred choice to make semi-active devices realistic [12–14].MRF is very responsive to magnetic field, with an estimatedresponse time of less than 10 ms [15], and requires relativelylow power to operate [16, 17]. Semi-active seat suspensionsfounded on MR dampers have begun to be investigated andhave achieved certain expectations. For example, Choi andHan [11, 18] applied MR dampers to attenuate the verticalseat vibration. Choi and Wereley [19] investigated theapplication of MR dampers to helicopter crew seats toenhance the survivability of a crash. Bai and Wereley [20]extended their work on the application of MR seat suspensionfor mitigating ground vehicle crashes. The current MRdampers have all been linear dampers with inherent draw-backs which need to be addressed: (1) in principle, highpressure inside damper is always needed to avoid the force lagperformance, which require advanced sealing technology forthe avoidance of fluid leakage; (2) With respect to the usageof MRF, the reserve of the linear damper must be totally filledup, which results in higher consumption of the MRF materialand high cost. For the above reasons, there is great possibilityto advance the current seat suspension design by overcomingthe above-mentioned requirements.

In order to reduce the requirements of a conventional seatsuspension with linear dampers, this paper proposes a newseat suspension structure using a rotary MR damper. The seatsuspension with a linear MR damper has been widelyinvestigated but the research on semi-active seat suspensionutilising a rotary MR damper is very rare. Compared withconventional seat suspension with a linear MR damper, whichneeds large amounts of MR fluids to fill the reserve of thelinear MR damper, a rotary damper based seat suspensiononly utilizes a smaller quantity of MRF because the reserve ofthe rotary MR damper is much smaller than that of a linearone. This advantage significantly reduces the cost of the seatsuspension [21]. Another advantage over linear dampers isthe lack of need for sealing the rotary MR damper. Theworking mechanism of the linear MR damper relies on thepressure difference between the two sides of the piston, whichmeans the pressure of at least one side is quite high and thusneeds good sealing. In terms of the rotary MR damper, the

working mechanism is different and the MRF pressure isquite low, thus it is much easier to solve the sealing problem.The advantages of compact structure, anti-sedimentation ofMR fluids and easy assembly of rotary MR dampers shouldalso be noted.

For these reasons, the development of a rotary MRdamper for heavy duty vehicles aims at upgrading the currentseat suspension technology and improving the ride comfort ofdrivers. The specific contents of this paper are introduced asfollows. Following the introduction, the structure and designof the proposed seat suspension will be presented. Theproperty test of the seat suspension will be conducted insection 3 and the test results will be discussed. Sections 4 and5 present the vibration attenuation performance of the seatsuspension numerically and experimentally, respectively. Theconclusion is drawn in section 6.

2. The structure and design of the seat suspension

2.1. The structure of the seat suspension with rotary MRdamper

Figure 1 shows the structure of the proposed seat suspensionsystem which is mainly composed of the scissors structurewith rotary MR damper installed. One beam of the scissorsstructure is fixed with the shaft of the damper and the other isfixed with the outside cylinder. When this seat intends to

Figure 1. The structure of the seat suspension: (a) workingmechanism; (b) photograph of the suspension system.

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Smart Mater. Struct. 25 (2016) 105032 S S Sun et al

vibrate vertically, the relative motion of these two crossedbeams will drive the rotary damper to work. In this way, thevertical linear movement is transformed into rotational motionof the rotary dampers. Subsequently, the damping variation ofthe rotary MR damper will induce the overall damping var-iation of the seat suspension.

2.2. The design of the rotary MR damper based seatsuspension

The design of the seat suspension, including its structure andpeak force, is important for the seat to achieve better ridecomfort. Specifically, the structure of the seat suspensionshould be compact and the maximum force of the MR dampershould be able to emulate the maximum desired control forcethat occurs under worst-case base disturbance. This subsec-tion presents the design and peak force calculation procedurefor the seat suspension in order to provide a design guideline.

Figure 2 shows the structural design of the proposedrotary MR damper. In general, it can be seen that this rotaryMR damper consists mainly of a cylinder, a coil, a shaft, arotor, and the enclosed reservoir filled with MRF. Unlike thelinear MR damper with single shaft, the rotary MR damperdoes not need a spring or gas chamber to provide the initialpressure in the damper. The rotor is made of low-carbon steeland the shaft is made of aluminium. The choice of materialfor different parts is under the consideration of magneticcircuit design. The coil is composed by winding the copperwire around the rotor. When the coil is energized, two sets ofmagnetic circuits are formed as shown in figure 2. Thismagnetic circuit design guarantees that the magnetic fluxpasses through the MRF.

The peak torque is the most important index for therotary MR damper design and directly determines the ridecomfort of the seat suspension. The calculation method isshown as follows [22]:

( )( )

( )

pm pt

p t

=+

- W + -

+ +W

t

+⎡⎣⎢⎢

⎛⎝⎜

⎞⎠⎟

⎤⎦⎥⎥

⎛⎝⎜

⎛⎝⎜

⎞⎠⎟

⎞⎠⎟

TR

n d

R

RR R

R L KR

d

4

31

4

3

2 ,

1

n

n

eq o4

e

i

o

3ye

o3

i3

o2

a ya aO

o

e

a

( )( )

( )

( )

pm pt

p t

=+

- W + -

+ +W

t

+⎡⎣⎢⎢

⎛⎝⎜

⎞⎠⎟

⎤⎦⎥⎥

⎛⎝⎜

⎞⎠⎟

TR

n d

R

RR R

R w KR

d

4

31

4

3

2 ,

2

n

n

oeqo o

4

o

i

o

3o

o3

i3

o2

d o oO

o

o

o

where tT is the torque with magnetic field and tT o is the torquewithout magnetic field. t K nandye, e, e, are yield stress, theconsistency and the behaviour index of the MRF in the end-

face duct. m ,eq governed by m =W -

⎜ ⎟⎛⎝

⎞⎠K

R

d,

n

eq eo

1e

is the

equivalent viscosity of the MRF in the end-face duct.@ -L w wa d c is the effective length of the annular duct.

The meanings of d, D, L, do, Ro, Ri, wc and wd have beengiven in figure 2. The detailed size of the rotary damper isshown in table 1. According to the size, the MRF usage forthe rotary damper can be calculated, which is 13.87 ml. As acomparison, the MRF usage in a representative paper [3]using a linear MR damper for seat suspension is approxi-mately 126.24 ml. The MRF usage in the rotary damperdesigned in this paper is significantly less.

Based on the reference [23], the rheological properties ofthe MRFs can be estimated by the following equation:

( )( ) ( )= + - -¥ ¥- -a aY Y Y Y e e2 , 3oB B2SY SY

where Y represents rheological properties of MRF includingpost-yield viscosity, yield stress, fluid consistency, and flowbehaviour index. The value of the MRF rheological propertiesY has a range of Yo (zero applied field) to ¥Y (saturationvalue). B is the applied magnetic flux density and can becalculated by finite element analysis method. aSY is thesaturation moment index. Based on the curve fitting resultswith experimental results, the values of Y, ¥Y and aSY

corresponding to different rheological properties of MRF(LORD MRF-132DG) can be determined and they are givenin table 2. Subsequently, the peak torque of the rotary MRdamper under 0 A current and 1 A current can be calcu-lated: = =t tT T0.4 N m and 2.7 N mo .

As there is a gear box (ratio is 8) is mounted between therotary damper and the scissors structure, the torque working

Figure 2. The structure of the rotary MR damper.

Table 1. The size of the rotary MR damper.

do 1.5 mm wc 22 mmRo 34 mm wd 12 mmRi 20 mm L 45 mmD 100 mm d 1.5 mm

Table 2. The parameters for the MRF-132DG.

KO 0.22 Pa sn t ¥y 30 000 Pa

¥K 3900 Pa sn asty 2 T−1

ask 5 T−1 nO 0.917tyO 10 Pa ¥n 0.25

asn 32

3

Smart Mater. Struct. 25 (2016) 105032 S S Sun et al

on the scissors structure will be:

( )=¹=

t

t

⎧⎨⎩TT B

T B8 if 0,

8 if 0.4

o

The relationship between the MR damper torque and thevertical damping force working on seat suspension is essentialto guide the seat suspension design. Figure 1 shows thegeometry of the seat suspension and the relationship betweenthe MR damper torque T and vertical damping force fMRF canbe obtained as:

( )= = -f T w T L H2 2 , 5MRF o2 2/ /

where Lo is the length of the beam of scissor structure, H isthe height of the suspension. Lo is a constant value of0.287 m, and H can be measured in real-time.

Thus the overall force generated by the seat suspensioncan be obtained by:

( )= + +F kx f f , 6z r MRF

where k is the stiffness of the seat suspension, x is thedisplacement of the seat suspension, fr=64 N is the frictionforce.

In summary, the above section provides the designprinciple which will offer useful guidelines for seat suspen-sion design with a rotary MR damper. The comparison of thecalculation pick force with the tested peak force is shown infigure 5.

3. The property test of the seat suspension andresults discussion

3.1. Testing method

The dynamic properties of the seat suspension system aretested and evaluated in this section in terms of its field-dependent responses, amplitude, and frequency-dependentperformance. In figure 3, the seat suspension was fixed onto acomputer-controlled MTS machine (Load Frame Model:

370.02, MTS Systems Corporation), which was driven by aservo hydraulic system capable of exerting large axial loadsonto the test specimen. The self-contained linear displacementtransducer and load cells transmit the immediate responses ofthe tested specimen under harmonic excitations to thecomputer through a DAQ board. A DC power with twochannels was ready to energize the magnetic fields of thespecimen whenever needed. Once the testing started, thespecimen moved in accordance with the preprogramed sinu-soidal routine.

3.2. Test results

A series of experimental tests were conducted to evaluate andcharacterize the performance of this seat suspension using theabove described experimental setup. For field-dependenttests, various harmonic inputs under different current levels,which are used to energize the magnetic field working on theMRF, were chosen to load this seat suspension. The proper-ties of variable damping can be observed in figure 4(a). It canbe seen that the force–displacement relationships form dif-ferent parallelograms and the slopes of the parallelogramsincrease with the increasing current. It is also observed infigure 4(a) that the area of the force–displacement loops has alarge increase before saturation occurs when the current levelincreases from 0 to 2 A with a step of 0.5 A. As the variationof damping is usually indicated by the area change of theenclosed force–displacement loops, the results demonstratethat this seat suspension shows controllable dampingvariations.

Figure 4(b) shows the responses of force–displacementwhen the seat suspension was loaded with sinusoidal signalswith different amplitudes (1, 5, 10, and 15 mm) at constantfrequency (1 Hz) and current (0 A). The effects of changingthe amplitudes clearly show that the maximum force gains alarge increase with the increasing amplitude. This set ofresults shows the amplitude-dependent properties of the testedseat suspension. Likewise, the effects of changing frequencieson the performance of the seat suspension are presented infigure 4(c). It can be seen that frequency variations have slightinfluence on the maximum force, the area of the loops, andthe slopes of the parallelograms. In order to verify the cor-rectness of the calculation in section 2.2, the comparison ofthe tested peak force and calculated peak force is presented infigure 5, where the tested peak force is obtained by using thepeak force in figure 4(a) minus the preload force of 1045 N.The fitting result shown in figure 5 demonstrates the calcul-ation method in section 2.2 accurately predict the dynamicperformance of the seat suspension.

The above experimental results sufficiently verify thatthis proposed rotary MR damper has the same requiredproperties as the linear MR dampers do while having someadvantages that linear MR dampers do not possess. Thecompact design enables the rotary MR damper to be installedwithout any modifications to the original seat suspensionstructure and has further advantages in terms of weightreduction, smaller installation space, easy sealing and main-tenance. Additionally, the rotary damper design improves the

Figure 3. Test system for seat suspension.

4

Smart Mater. Struct. 25 (2016) 105032 S S Sun et al

cost efficiency because it requires smaller quantity of MRfluid than linear MR dampers do.

4. Numerical effectiveness evaluation of the rotaryMR seat suspension

4.1. The dynamic model of the seat suspension

Figure 6 shows the semi-active seat suspension model withstiffness k, friction fr and the Bouc–Wen model for the MRdamper. The suspension upper and base displacements are zs

and z ,v respectively. The governing equation of motion for theseat suspension is

( ) ( )+ - + + =mz k z z f f 0, 7s s v r MRF

where m is the total mass of a driver’s body, seat suspensiontop platform and cushion. In this dynamic model, the overallforce working on the seat suspension can be calculated by:

( ) ( )= - + +Fz k z z f f . 8s v r MRF

Based on the phenomenological model for magnetorheologi-cal dampers proposed by Spencer [24], the damping force

Figure 4. The test results of the seat suspension (a) different currents(b) different amplitudes (c) different frequencies.

Figure 5. The comparison between the experimental peak force andcalculated peak force.

Figure 6. The mathematic model of seat suspension.

5

Smart Mater. Struct. 25 (2016) 105032 S S Sun et al

fMRF can be obtained by the following equations:

( )

[ ( ) ]

∣ ∣∣ ∣ ( )∣ ∣ ( )

–

( )

a

g m

a a a a

= + -

=+

+ - +

=- - - - + -= -== + += + +

= + +

-

f c y k x x

yc c

z k x y c x

z x y z z x y z A x yx Z Zy Z Z

I I

c c c I c I

c c c I c I

,

1,

,,

,

,

,

,

9

n n

MRF 1 1 0

o 1o o

1

s v

m v

a b c2

o oa ob oc2

1 1a 1b 1c2

where I is the current applied to the MR damper. Theparameters m g a ax k k c c c A n, , , , , , , , , , , ,o 1 O oa ob oc a b

a c c c, , , andc 1a 1b 1c are used to characterize the MRdamper. The optimized values for the parameters aredetermined by fitting the model to the experimental dataobtained in the experiments. The least-square method incombination with the Trust-region-reflective algorithm avail-able in MATLAB (R2013b) can be used for the parameteridentification. The detailed parameter identification methodhas been given in our previous paper [25]. Figure 7 illustratesthe tracing performance of the modelling results andexperimental results and the identified parameters are shownin the table 3. It can be seen that the predicted data matchesthe experimental data very well.

4.2. Control algorithm

It is commonly recognized that the MR suspension system isdifficult to model accurately because of its nonlinear andcomplicated nature. Many control algorithms have beeninvestigated to control the MR dampers, including slidingmode control, robust control and fuzzy logic control [26–28].As fuzzy logic control does not rely on the mathematicalmodel of the full system, it is chosen to control MR damper inthis study. The parameters of the dynamic seat model, shown

in equation (6), are m=60 kg, k=4100 Nm−1. The inputvariables to the fuzzy controller are seat mass velocity Vs andrelative velocity Vrel while the desired current is its output.The fuzzy logic control rules are based on the sky-hookstrategy detailed as:

* ** *

=

<

⎧⎨⎩fc V V Vc V V V

0,0,MRF

max rel s rel

min rel s rel

where the damping cmax is the on-state damping and cmin isthe off-state damping. The on–off damping force is controlledby the sign of *V V .s rel Based on the sky-hook controlalgorithm, the fuzzy logic control rules can be derived. Thefuzzification of Vrel and Vs used in this research is NB, NS,ZE, PS and PB while the fuzzification of current is NB, NM,NS, ZE, PS, PM and PB. The fuzzy control rules are based onsky-hook control algorithm. For example, when the Vs andVrel are NB, the output is maximum current which iscorresponding to PB or NB in the fuzzy control algorithm.The other rules are designed following the same method. Thefinal control rules are shown in table 4 [28]. It should be notedthat the restriction of MR damper still exists in this study.Specifically, the MR damper should generate an active forcein the same direction with relative motion when * <V V 0s rel

in order to achieve better vibration control performance. TheMR damper, however, can only generate a force in theopposite direction with the relative motion and as a result thecurrent is set as zero when * <V V 0.s rel

4.3. The numerical simulation results

4.3.1. The simulation results under harmonic excitation. Inthis test, the excitation, Zv, is a sweep frequency signal with10 mm amplitude. Its frequency linearly varies from 0.1 to2.5 Hz in 25 s. The simulation results are shown in figure 8.Passive-off means that the rotary MR damper is out of control

Figure 7. The fitting result of the mathematic model.

Table 3. The identified model parameters.

kO 1500 A 2k1 700 n 2xo 0 aa 89 839.1g 6e+5 ab 205 322.1m 3e+6 ac −65 841coa 198.1 c1a 155 679.6cob 635.6 c1b 9.04705E+6coc −221.4 c1c −375 621.5

Table 4. The rules of the fuzzy controller.

Vrel

Vs NB NS ZE PS PB

NB PB PM PS ZE ZENS PM PS ZE ZE ZEZE PM PS ZE NS NMPS ZE ZE ZE NS NMPB ZE ZE NS NM NB

6

Smart Mater. Struct. 25 (2016) 105032 S S Sun et al

because no control current was applied to it in this case.Passive-on refers to the situations where a certain currentsignal was chosen (0.25, 0.5, 1 A) to energize the rotarydamper, and semi-active case means the rotary damper wasunder control of the fuzzy logic. It can be seen that the seatsuspension with off state MR damper performs the worst andthe resonance happens at around 16 s. On the contrary, theresonance peak is lower and lower with increasing current andcan be barely seen for the seat suspension with 1 A current orcontrolled MR damper but it still can be observed that thesemi-active seat suspension performs better than all thepassive seat suspensions.

4.3.2. The simulation results under random excitation. Therandom excitation is generated by the following method.Firstly, the road displacement of the random ground surface[29, 30] can be defined as:

( ) ( ) ( ) r+ =z t Vz t VW , 10r r n

where ρ is the road roughness parameter, V is the vehiclespeed, andWn is white noise with intensity d rV2 2 in which d2

is the covariance of road irregularity. When the randomground profile is being generated, r d= =-0.45 m ,1 2

300 mm ,2 and = -V 20 m s 1 are chosen. Then, a quarter carmodel is excited by this random ground surface and theresponse of the sprung mass will be used as the excitation forthe seat suspension.

The simulation results are shown in figure 9 and it can beseen that the seat suspension with passive-off MR damperperforms the worst and the controlled semi-active seatsuspension performs the best.

5. Experimental effectiveness evaluation of therotary MR seat suspension

5.1. Test system

The vibration isolation experiment was conducted to evaluatethe effectiveness of the rotary MR damper to attenuate thevertical vibration. Two different excitations (constant

frequency harmonic excitation and random road profile) wereemployed to evaluate the vibration reduction effectiveness ofthe proposed MR seat suspension. The experimental testingsystem is shown in figure 10. A normal commercial vehicleseat (GARPEN GSSC7) was fixed to the suspension system

Figure 8. The simulation response of seat under harmonic excitation.Figure 9. The simulation response of seat under random excitation.

Figure 10. The effectiveness evaluation system for the seat vibrationcontrol.

7

Smart Mater. Struct. 25 (2016) 105032 S S Sun et al

which was equipped with the rotary MR damper. The seatsuspension system was driven to vibrate vertically by avibration platform, as shown in figure 10. This vibrationplatform is controlled by NI CompactRio 9076. One dis-placement sensor (Micro Epsilon ILD1302-100) is installedon the bottom base of seat to measure the suspension relativedisplacement while the other displacement sensor (KeyenceLB-11) was used to measure the seat’s absolute displacement.Two accelerometers (ADXL203EB) were used to measureseat acceleration and excitation acceleration, respectively.Those measured signals were feedback to the seat suspensioncontroller through a NI CompactRio 9074. The suspensioncontroller then output a desired current signal. The currentsignal will be amplified by a power amplifier and then used tocontrol the rotary damper.

5.2. Test results

5.2.1. The test results under harmonic excitation. In this test,different harmonic signals with different frequencies wereused to excite the seat suspension. The responses of the seatsuspension under two frequencies, 1.75 and 2 Hz, are given as

representative examples. Figures 11(a) and (b) present theacceleration responses of the seat suspension under 1.75 Hzand 2 Hz, respectively. The performance of the seatsuspension under passive-off (0 A), passive-on (0.25, 0.5,1 A) and semi-active control cases were considered for eachfrequency testing. The comparison results between thesecases are persuasive in terms of the performance of the semi-active rotary MR damper. Based on the testing results shownin figure 11, the semi-active case holds the minimumamplitude of the seat acceleration. The passive-off caseperforms worst when the excitation frequency is 1.75 Hz,however, it performs better than the passive-on cases whenthe excitation frequency is 2 Hz.

In order to further verify the vibration reductioneffectiveness of the semi-active seat suspension, transmissi-bility was measured under different control cases. Thetransmissibility is defined as the ratio of the seat accelerationto the excitation acceleration. The smaller the ratio is, themore effective the suspension is. The experimental results, asshown in figure 12, verified that the semi-active case can

Figure 11. The test results of the MR seat suspension underharmonic excitation.

Figure 12. The transmissibility of the MR seat suspension underdifferent frequencies.

Figure 13. The performance of MR seat suspension under randomexcitation.

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Smart Mater. Struct. 25 (2016) 105032 S S Sun et al

effectively reduce the transmissibility and performs the bestover a large frequency range.

5.2.2. The test results under random excitation. Apart fromthe constant frequency excitation, random excitation was alsoused to test the seat suspension. Figure 13 shows thecomparison of seat acceleration under passive-on cases(0.25, 0.5, 1 A), passive-off and semi-active control case.The acceleration amplitude under semi-active case is largelyreduced compared to all passive-off and passive-on cases. Amore direct comparison is given by calculating the rms underthose three control cases. Both the rms values from thesimulation and experimental results are given in table 5.Clearly, the semi-active control case has the smallest rmsvalue, indicating that the rotary MR damper under fuzzy logiccontrol has the most effective vibration reduction capability.

6. Conclusion

This paper proposes an innovative seat suspension designwith a rotary MR damper, which can reduce not only theamount of MRF needed but also the sealing requirements.The testing results on the MTS machine show that it has thecapability to vary damping in response to any change incurrent. Then an evaluation system is constructed, includingthe semi-active vehicle seat suspension, the control loop andvibration platform providing the excitations to this system.The vibration reduction effectiveness of the seat suspension isevaluated under two kinds of excitations, i.e. constant fre-quency excitation and random excitation. For both of thesituations, the seat acceleration under fuzzy logic controlholds the minimum value, which verified the semi-activerotary MR damper based seat suspension has the mosteffective vibration attenuation performance.

Acknowledgments

This research is supported by ARC Linkage Grants(LP160100132, No. 150100040), Open Research Fund Pro-gram of the State Key Laboratory of Advanced Design andManufacturing for Vehicle Body, Hunan University(31515001) and the University of Wollongong and ChinaScholarship Council joint scholarships. The authors wish togratefully acknowledge the help of Dr Madeleine StrongCincotta in the final language editing of this paper.

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Table 5. The acceleration rms values.

0 0.25 A 0.5 A 1 A Semi-active

Experiment 1.023 1.050 0.803 0.871 0.621Simulation 1.104 1.022 0.861 0.815 0.663

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