24

Chapter 2

LITERATURE REVIEW

In the presented chapter, a brief review of the past research carried out in the

field of regenerative electromagnetic shock absorber has been discussed in brief, for

identification of the gaps, and for proposal of the new research work listed in the present

study.

2.1 Electromagnetic Regenerative Shock Absorbers

Electromagnetic regenerative shock absorbers convert kinetic energy of

vibrations into useful electricity (Zhang et al., 2013). Depending on their structural

configuration, these systems are broadly classified in three types as,

- Electromagnetic shock absorber with linear generator

- Electromagnetic shock absorber with rotary generator

- Hybrid systems with hydraulic transmission

2.1.1 Electromagnetic shock absorber with linear generator

This type of systems replace viscous damper, as present in the conventional

automotive suspensions with a linear generator. The linear generator moves with

relative motion between sprung and un-sprung mass of the vehicle directly and does not

involve any additional transmission elements.

Linear electrical devices are used to generate electricity, where the prime mover

moves through reciprocating motion within few centimeters (Ping et al., 2006, Kou et

al., 2008, Szabo et al., 2007). However, since last two decades researchers are

investigating its application as a pertinent alternative to fluid damper in conventional

shock absorbers. It is shown that when used in an automobile shock absorber, linear

generator can harvest considerable amount of electric power for the suspension

velocities; those are normally encountered on the road (Zuo et al., 2010; Zuo and

Zhang, 2013). Vibration energy scavengers consisting of electromagnetic transducers

with rare earth magnets ensure high energy output (Beeby et al., 2008).

Permanent magnet linear devices are preferred in regenerative electromagnetic

shock absorber to provide damping and regeneration. Moreover, these can also be used

25

to apply active force in the suspension system. Ability to manipulate their damping

force with low cost electronics control gives it advantage for implementing semi-active

suspension, in comparison to the conventional fluid solution (Mirzaei et al., 2001).

Liu and Lin (2013) investigated feasibility of developing a

suspension(regeneration) unit consisting of tubular permanent magnet linear generator.

Authors have designed an electromagnetic device which can generate maximum power

of 12 W and 400 N damping force, when fitted in a car shock absorber. Gupta et al.

(2011) conclude that large magnetic field in the air gap of linear generator ensures

efficient energy recovery. Authors have validated analytical model of the prototype

linear generator consisting of rare earth permanent magnets. During experimentation,

the prototype could harvest maximum energy of 0.28- 0.0029 W for excitation

frequency of 10Hz to 100Hz. Goldner et al., (2001) validated analytical model of a

linear electromagnetic harvester consisting of rare earth magnets. It has been concluded

that the device can convert 20-70% of the vibration energy into useful electricity, which

is otherwise lost in conventional fluid dampers.

Oprea et al. (2012) explained analytical framework for using linear generator as

a damper in vehicle shock absorber. Finite element simulations are used to determine

the generator dimensions for maximum flux density in the air gap. Authors have

observed that damping force is not exactly proportional to the excitation velocity and

few harmonics are present in the force. Design and analytical simulation of linear

generator consisting of rare earth magnets has been discussed by Zuo et al. (2010).

Authors have validated voltage waveform results of the theoretical model with

experimentation on a reduced scale prototype. It has been estimated that real size

version of the device will harvest 16-64 W of power for normal vertical suspension

velocities. Zhen and Wei (2010) have analyzed structure of linear generator for a

vehicle with mass of 1.3 ton. Theoretical calculations revealed that the disclosed device

will give maximum damping force 1145 N for suspension vertical velocity of 0.17 m/s.

However, the work does not demonstrate model of the linear generator fabricated to

verify the calculations.

Goldner and Zerigian (2005) superimposed magnetic field by two concentric

columns of magnets to achieve higher flux density in the generator air gap. Authors

have used analytical calculations and estimated maximum theoretical efficiency of the

device to be 44%. However bulky structure of the device with mass of 70 kg makes it

impractical for any commercial application.

26

Energy harvested and damping force in the linear generator increases with

higher magnetic flux density in the air gap (Zuo et al., 2010; Goldner and Zerigian,

2005). Therefore, in the previous studies much effort has been made to improve the air

gap flux density by optimizing the generator configuration with FE analysis (Mikolanda

et al., 2009; Inoel et al., 2005; Parel et al., 2010; Yilmaz and Krein, 2008; Zhao et al.,

2005). Also numerical or analytical simulation has been used to select the winding

parameters for efficient energy recovery within the desired velocity range (Baker et al.,

2004; Petkovska and Cvetkovski, 2013).

Ebrahimi et al., (2008) presented equations to compute overall dimensions of the

generator for maximum damping force. Simulations performed with the numerical

model estimated maximum damping coefficient to be 9319.5 N-s/m. In his doctoral

thesis Zador (2008), presented tubular slotted linear synchronous generator

configuration for integration with McPherson strut suspension. Finite element analysis

and analytical modeling approach has been used to optimize damping force of the

device. Simulation results show that with use of rare earth magnets, maximum force of

4000 N can be achieved for the coil relative velocity of 0.7 m/sec. Ohashi and

Matsuzuka (2005) evaluated a prototype linear synchronous generator with neodymium

magnets. It was observed that amplitude of the generated voltage is inversely

proportional to square of the excitation frequency.

Energy harvested by the regenerative shock absorber can be utilized for some

useful application like vehicle battery charging. One possible arrangement for

controlling the harvested energy, consists of rectifier diode bridges and a shunt

resistance (Paz, 2004). Moreover, this arrangement ensures continuous damping by the

electromagnetic device.

Linear generator can be used in self powered vibration control systems

(Khoshnoud et al. 2013; Tang and Zuo, 2010; Kang et al., 2003). These systems use

linear generator as the harvesting and actuating element. Depending on the input

excitation velocity, the controller adjusts the system for regenerative, drive or brake

mode. In regenerative mode, the linear generator acts as harvester to convert kinetic

energy into electricity, which may be stored in a battery. During drive mode, the

generator functions as an actuator and utilizes the recovered energy for vibration

attenuation by applying active force in the system. During brake mode, voltage across

the generator is lesser than the battery threshold voltage and the system provides

necessary equivalent damping. Simulation tests on these systems indicate that although

27

their performance is inferior to fully active systems, they are better than that of passive

systems (Nakano et al., 2003).

Suda et al., (1998) used two linear DC motors to build a self powered active

vibration control system. As illustrated in Figure 2.1, it comprises of an actuator and a

harvester, with the harvester utilizing the actuator energy for vibration attenuation by

applying active force. The control algorithm ensures than the system works as passive

damper at lower suspension velocities and switches to active mode at the higher vertical

velocities. Active vibration control system with energy generation capability, proposed

by Orkisz (2011) consists of two linear DC motors that can act as harvester or actuator.

This system uses a battery or capacitor to store the harvested electrical energy.

Figure 2.1 : Active vibration control system (After Suda et al., 1998)

Ebrahimi et al., (2011) incorporated passive eddy current damping and

electromagnetic harvester in a single unit, as shown in Figure 2.2. Authors claim that

the proposed solution integrates superior performance of the active system along with

reliability of the passive damper, in a single pack. Numerical model of the system was

validated with experimentation performed on single degree of freedom quarter car

model. Finally, FE method along with numerical simulations is used to derive the full

size damper configuration, which would achieve damping coefficient of 1570 N-s/m.

28

Moreover, it was calculated that the system will also generate about 70% of the power,

required to sustain a fully active electromagnetic damper.

Figure 2.2 : Hybrid electromagnetic/eddy current damper

(After Ebrahimi et al., 2010)

Energy flow management for an electromagnetic suspension devices have been

discussed by Stribrsky et al. (2007) and Wang et al. (2012). The discussed systems use

generator force and suspension velocity signal for implementing the system control.

When the generator force is in same direction as that of the suspension velocity, energy

from the harvester is used to apply active force in the system. Otherwise, it is

accumulated in a battery for future use. Bose company used linear generator in vehicle

suspension, which is used as the generator/actuator, as shown in Figure 2.3. However,

rather than energy generation, this system consumes about one third of power required

to operate the car air conditioner (Zhang et al., 2013). The development and integration

of energy harvesting system for a MR damper is discussed by Lafarge et al. (2012), as

shown in Figure 2.4. Electrical energy harvested by the generator is used to operate the

MR damper, with both generating and MR damper systems integrated as a single unit.

Numerical model of the quarter car system has been implemented in Matlab Simscape.

It has been concluded that the harvester produces enough power to ensure autonomous

working of the MR damper.

29

Figure 2.3 : Electromagnetic shock absorber developed by Bose

(After Zhang et al., 2013)

Figure 2.4 : Self powered MR damper (After Lafarge et al., 2012)

Vehicle shock absorber patented by Klausner and Yankowski (1975) consisted

of two electromagnets. First electromagnet has fixed polarity whereas the other operates

with active control. It has been concluded that implementation of passive and active

control methodology results in better vibration attenuation.

2.1.2 Electromagnetic shock absorber with rotary generator

A rotary generator can be used in an energy harvesting shock absorber. Ball

screw or rack-pinion arrangement with bevel gears, which convert reciprocating motion

into rotary, can be used to drive the generator (Zhang et al., 2013). Number of helical

30

gears may also be incorporated to increase rotational speed of the generator. In

comparison to direct drive linear generators, rotary harvesters operate with amplified

velocity, which results in greater power output (Gupta et al., 2007). This type of

arrangement is favored in self-powered active vibration controllers. Ball screw-nut and

bevel gear arrangements used in regenerative suspensions are illustrated in Figure 2.5 (a

and b).

Figure 2.5 : Regenerative shock absorbers (After Zhang et al., 2013)

(a) Ball screw arrangement (b) Bevel gears arrangement

Murty (1988) constructed an electromagnetic suspension system, comprising of

three phase rotary electric alternator, with ball screw and nut arrangement. The device

incorporated variable electric load to change the damping intensity. Harold (1971) used

long spiral lead screw for converting reciprocating motion of the shock absorber into

rotary motion, which was later used to drive an electrical generator. Output energy has

been stored in a battery. Authors have controlled ride stiffness by changing resistance,

connected in series with the battery. In case of ball-screw arrangement, power

transmission section (ball-screw and helical gears) delay rotation of the generator from

that of input excitations. It has been concluded that, this delay gives some time to adjust

31

the inertia of generator rotor, resulting in lesser acceleration transmissibility (Kondo et

al., 2008).

Application of rotary electromagnetic damper in active suspension system is

demonstrated by Kawamoto et al., (2007). An electric motor is used to apply active

force with use of ball screw arrangement. Amati et al., (2011) illustrated design and

modeling of electromagnetic damper comprising of rotary DC motor and ball screw

arrangement. The system has electric terminals of the motor shunted by a variable

resistor, to control the damping characteristics. Authors have observed that at higher

frequencies, electromagnetic damper acts as a mechanical spring. With damping

coefficient-to-weight ratio at 2000, good compromise has been achieved between mass

and the damper force. Integration of full scale version of the system in McPherson

suspension has been also discussed in the paper. Design, modeling and testing of

retrofit regenerative shock absorber consisting of bevel gears and rack-pinion, has been

presented by Li et al., (2013). Special care has been taken while fabricating the

prototype by preloading the rack and use of teflon ring, to minimize the gear backlash

and frictional force. Since failure of the mechanical gears will result in zero damping

force, thorough evaluation of contact fatigue and teeth strength has been recommended.

Authors have observed that up to 162 W of power is dissipated in each car shock

absorber, whereas the prototype device could generate peak power of 68 W and average

power of 19 W. In the theoretical modeling, rotary generator is assumed to be a

torsional damper and inertia of the system is considered together with the gear

transmission as equivalent mass. Damper asymmetry has been achieved with

combination of rectifier diodes and electrical resistances in the disclosed device. A

smaller gear transmission ratio has been recommended to achieve higher efficiency and

good compactness.

Choi et al. (2009) have used rack and pinion mechanism to amplify relative

displacement in the shock absorber. An electric generator is driven by the pinion and the

output energy is used to sustain semi-active suspension system employing

electrorheological fluid. It was observed that voltage generated using the amplification

mechanism is about 1000 times the voltage generated without amplification.

Possible use of the electromagnetic device includes their application in active

control systems. Such devices harvest the vibration energy at higher suspension

velocities, which is later used to apply active force. An electro-chemical solution for

this methodology includes use of an electromagnetic generator/actuator with capacitor

32

and accumulator (Zheng et al., 2008). The generator/actuator assembly consists of

brushless DC motor with lead screw arrangement. The disclosed device consists of an

electronic capacitor, which stores the vibration energy immediately after the vertical

excitation. The control algorithm operates with a sensor used to measure suspension

deflection and use the capacitor energy either for applying active force or the energy is

stored into an electro-chemical accumulator (battery). When needed the accumulator

energy can be used for active control or to run other electrical equipments in the vehicle.

Since this system used original vehicle battery as the accumulator, no additional weight

is imposed for implementing the regenerative suspension.

Coil configuration is an important consideration in design of electromagnetic

harvesters. Voltage and useful power output from the electromagnetic device depends

on the coil copper wire diameter and number of turns. Referring to Figure 2.6, number

of turns in the coil is governed by copper wire diameters and density with which the

coils are wound. The copper wires do not fill the coil volume completely and percentage

of copper wires filled in the coil is given by the term called coil filling factor (Cekov

and Bossche, 2005).

Coil filling factor = wi re c

co i l

A N

A (2.1)

where Nc : number of copper wire turns in each phase

Acoil : overall cross section area of the coil

Awire : cross section area of the copper wire

Coil filling factor should be higher for better power output from the

electromagnetic devices. The coil filling factor depends on tightness of winding,

accuracy of the winding machine and shape of the coil. Filling factor for the round

conductors is in the range of 0.4-0.6 (Priya and Inman, 2009).

Figure 2.6: Copper wires in the armature coil

33

High density rare earth magnets are recommended for electromagnetic

harvesters used in vehicle shock absorbers (Furlani, 2001). These magnets are 7-8 times

stronger than conventional ferrite magnets and ensure higher energy recovery with a

much compact size (Jensen and Mackintosh, 1991; Jiang and Lang, 2007). However,

since these possesses very high magnetic coercive force (up to 10800 KOe), special

precautions needs to be followed during their assembly to avoid human injury and

damage to the magnets.

Rare earth magnets are composed of lanthanide group of elements. The two

lanthanide materials preferred for magnets are Neodymium (Nd) and Samarium (Sm).

Most common commercially available magnet materials are Neodymium-Iron-Boron

(Nd-Fe-B) and Samarium Cobalt (Sm-Co). These magnets are available in bonded and

sintered form, with sintered form having higher field strength than the other type.

Sintered NdFeB magnets are generally coated to avoid corrosion (Jensen and

Mackintosh, 1991).

2.1.3 Hybrid systems with hydraulic transmission

Regenerative shock absorbers consisting of ball-screw and rack-pinion

arrangement with planetary gear box include large number of elements in motion

transmission, which limits performance of these systems. Also frequent reversing

because of the shock absorber motion, loses significant power due to inertia loss of the

rotating components. On the contrary, hybrid hydraulic electromagnetic systems use

pressurized fluid for velocity amplification. These systems have improved efficiency

due to absence of inertia effect, which increases the generator speed without time delay

and does not involve energy loss due to frequent reversing.

Regenerative hybrid electromagnetic damper, called ‘GenShock’ developed by

Levant Power Corporation, is shown in Figure 2.7. It consists of rectifier pipes,

hydraulic motor and an electrical generator (Web1). As piston reciprocates inside the

cylinder due to external stimulus, high pressure fluid flows through rectifier pipes to the

hydraulic motor, converting linear motion to rotary. The hydraulic motor is connected

to the rotary electrical generator. Rectifier pipes ensure consistent direction of rotation

for the hydraulic motor, and eliminate loss of energy due to frequent reversing. This

arrangement gives higher power output in comparison to the rotary electromagnetic

harvesters with mechanical gears.

34

Figure 2.7 : Hydraulic transmission in hybrid regenerative damper

(After Web1)

Hydraulic Electromagnetic Shock Absorber (HESA) proposed by Fang et al.

(2013), shown in Figure 2.8 consists of check valves, accumulators, hydraulic motor

and an electric generator. During its working, high pressure fluid flow from the damper

cylinder to accumulator through the check valves, due to linear movement of the piston.

Difference in pressure settings of the check valves ensures desired damper asymmetry.

Pressurized fluid, as it leaves the accumulator with reduced pressure variations, drives

the hydraulic motor, which in turn rotates the electrical generator. It has been reported

that, use of hydraulic accumulators ensures lesser frequency fluctuations in the

harvested electrical energy.

Figure 2.8 : Hydraulic electromagnetic shock absorber

(After Fang et al., 2013)

35

Zhang et al. (2013) have revealed that reliability is a key factor in developing an

electromagnetic suspension. It has been concluded that both vibration isolation

performance and harvesting efficiency of regenerative shock absorber are affected due

to inertia of its components. Primary purpose of a suspension system is to provide better

vibration control. However, when using regenerative electromagnetic shock absorber,

there is always some conflict between harvesting efficiency and shock absorber

performance (Zhang et al., 2012).

2.2 Fluid damper modeling

Fluid dampers dissipate vibration energy by throttling viscous fluid through

restricted orifices in the piston. Accurate modeling of the damper is essential for

suspension design and analysis. Moreover, it is also aimed at reducing time and cost in

developing new products (Lee, 1997).

Jiuhong et al., (2008) analyzed fluid flow between two parallel plates with

approximate numerical analysis. Based on the analysis, theoretical model has been

derived to calculate performance parameters of the damper. Mollica and Youcef-Toumi

(1997) developed dynamic model of a monotube shock absorber. Authors have included

laminar flow, cavitation and fluid inertia effects to identify interaction between

mechanical and fluid elements of the damper. Accuracy of the analytical model has

been verified within the stimulated frequency range of 0.5-3.3 Hz. Further, Bond Graph

method has been used to derive simple analytical model of the monotube damper. Non-

linear parametric model of the monotube damper, proposed by Liberati et al., (2004)

was able to compute the damping force with high accuracy. Authors have interfaced the

numerical model to ADAMS virtual prototyping.

Zhou et al., (2008) introduced governing differential equations for deformation

of the shim valves, which are solved with suitable boundary conditions to develop

analytical model of the fluid damper. Simulation results from the analytical model are

then used to establish characteristic model by piecewise linear function. Shim valves as

shown in Figure 2.9, are used in automobile dampers to control piston valve fluid flow

area (Kulkarni et al., 2013a; Kulkarni et al., 2013b). Deformation of the shim valves has

been used to calculate oil flow rate through the valves and damping force. Further, this

analytical model has been validated with FE fluid flow analysis.

36

Figure 2.9 : Shim valves in fluid damper (After Dixon 2007)

Eyres et al. (2005) outlined several methods for modeling of fluid damper

comprising of spring loaded relief valve.

Referring Figure 2.9, damper force in compression stroke is calculated as,

1 2( ) ( ) p frictionF t P P A F= − +

(2.1)

where, P1 : compression chamber pressure

P2 : rebound chamber pressure

Ffriction : friction force

Ap : piston area

Pressure difference between compression and rebound chamber is due to viscous

frictional losses because of the flow through orifice and at the exit of orifice. Numerical

model has been used to evaluate the laminar and turbulent frictional losses for

computing the damping force. Dynamics of the spring loaded bypass relief valve has

also been considered in the model (Eyres et al., 2005).

Analytical approach uses computation of viscous and turbulent pressure drops at

the piston valve for finding the damping force (Zhou et al., 2009). Numerical

simulations with NASTRAN® software have been used to derive the damper valve

configurations for desired performance by following fluid structure modeling approach

(Czop et al., 2012). Authors have recommended use of few simplification assumptions

for reducing computation time. Two dimensional unsteady Euler’s equation has been

37

used to construct numerical model of hydraulic shock absorber (Koren et al., 1995). The

damper performance has been improved by reducing turbulence effect in the fluid flow

with simulations on dynamic model in Ansys CFX software (Czop et al., 2011).

2.3 Stroke Dependent Damper

Stroke sensitive damper construction changes force intensity depending on the

shock absorber displacement. It can be observed from force-displacement curve

illustrated in Figure 2.10 that, stroke sensitive dampers give lesser force intensity for

displacement near its mean position. This ensures better ride comfort at lower excitation

velocities. However at higher suspension velocities, larger shock absorber displacement

ensures elevated force, so that tire does not lose contact with the ground.

Figure 2.10 : Force displacement curve for stroke sensitive dampers

(After Lee and Moon, 2006)

Lee and Moon (2006) developed analytical model of Displacement Sensitive

Shock Absorber (DSSA), which offers two modes of damping (soft and hard),

depending on position of the piston inside the damper cylinder. Construction of DSSA

is shown in Figure 2.11, which has additional flow passage in the cylinder wall, near the

equilibrium position of the piston.

Additional flow passage near equilibrium position of the piston in DSSA, shown

in Figure 2.11 ensures lower force level for small amplitude road disturbances.

However, as the piston moves away from its mean position, annular clearance between

piston and cylinder is reduced to ensure higher force level. Analytical model of the

38

damper is constructed with flow continuity equations, which has been validated with a

prototype on quarter car test rig.

Figure 2.11 : Displacement sensitive fluid damper

(After Lee and Moon, 2006)

Stroke dependent damper model proposed by Etman et al., (1997) consist of dry

friction and fluid damper. The device is investigated for acceleration transmissibility

and damping force characteristics with numerical simulation on two degree of freedom

quarter car model. It was observed that stroke dependent damping reduces heavier

discomfort, at larger suspension velocities, resulting due to incidental road disturbances.

Authors have suggested that stroke dependent damping characteristics can also be

achieved with provision of extra bypass channels and valves in the fluid damper.

Ellifson et al. (2013) has designed a stroke dependent damper with the cylinder divided

into primary and secondary volume. For the shock absorber displacement near its mean

position, primary volume controls damping, which gives lower force intensity. On the

other hand, higher force level is achieved for piston displacement away from its

equilibrium position, where secondary volume governs damping intensity. During this

part of the piston stroke, the device is designed to deliver higher force level. Cox (2014)

further added an adjustable metering valve to control damping force of the secondary

volume.

39

2.4 Mechanical Motion Amplification

Mechanical motion amplification has been identified as effective solution in

increasing sensitivity of transducers and vibration attenuation systems. Flexure based

displacement amplification mechanisms are used to improve harvesting efficiency of

piezoelectric devices (Buchwald et al., 1992; Kim et al. 2011; Moler et al., 2012). A

five bar symmetric compliant motion mechanism was developed for increasing

sensitivity of a piezoelectric transducer (Ouyang et al., 2005). The device provided

amplification ratio of 24.4 with much compact size. Authors have recommended use of

the disclosed mechanism for other applications in velocity amplification. Zeimpekins et

al., (2012) replaced conventional cantilever beam arrangement in a Micro-Electro-

Mechanical Systems (MEMS) transducer with a micro-machined mechanism. For the

identical natural frequency, the proposed mechanism could give higher amplification

(up to 40) in comparison to the conventional arrangement. Hadas et al., (2009) varied

stiffness of an electromagnetic energy harvester to increase its sensitivity. It was

concluded that increasing sensitivity of the device results in higher energy output or

reduction in size and weight.

Displacement in the vibrating structure can be magnified to ensure effective

implementation of the damping devices. Taylor and Constantinou (2003) designed

vibration attenuation system, consisting of toggle brace mechanism and fluid dampers.

Toggle mechanism magnified relative displacement in the structure and increased

effective stroke of the fluid dampers. It has been highlighted that besides being simple

in design, economical and reliable, the disclosed solution allows for incorporation of

small diameter and larger stroke dampers. Huang (2004) investigated design parameters

of the toggle brace mechanism, to achieve the desired force intensity. Huang and

McNamara (2009) evaluated motion amplification in a scissor-jack type amplification

device. Simulations performed on the mathematical model indicate that, the viscous

damper operating with amplified motion shows significant improvement in damping

efficiency.

Vehicle suspension proposed by Hendrowati et al. (2012) consists of

piezoelectric generator and force amplifying mechanism, shown in Figure 2.12. Authors

have investigated the device for acceleration transmissibility and generator voltage, with

simulations on quarter car model. Numerical analysis revealed that the disclosed

arrangement will increase the voltage output by 175%.

40

Figure 2.12 : Force amplification mechanism (After Hendrowati et al., 2012)

Restriction of movement for the inertial mass, within available suspension stroke

limits electrical power generated by the linear electromagnetic devices. To overcome

this limitation, Moss et al. (2014) proposed hybrid rotary-translational motion

amplification device, which uses mechanical advantage to achieve significant

improvement in power density.

2.5 Experimental Studies

Validation of theoretical study on vibration analysis is accomplished by

introducing some forcing function into the system. Experimentation is performed to

examine response of the device under test for defined vibration input to reveal important

characteristics like motion and force transmissibility. Hota and Vakharia (2014)

explained performance parameters for evaluation of vibrating shaft. Authors have used

single degree of freedom lumped mass formulation to analyze peak transmissibility

under different modes. Shah et al. (2014) analyzed vibration signal to study dynamic

behavior of the machinery using NX software. Williams et al., (2005) validated

numerical study of adaptively tuned vibration absorber system with two degree of

freedom analytical model. The disclosed device could achieve significant reduction in

steady state vibrations with approximately 15 % variation in the system natural

frequency.

Vibration evaluation of automotive shock absorbers on test rig investigates its

dynamic performance over the preferred frequency range. During experimentation on

the shock absorber, force is applied with an electrical or hydraulic actuator at one end,

whereas the other end is fixed or supporting sprung mass. Scotch yoke mechanism is

also used in the shock absorber test rigs to convert rotary motion of the prime mover

into reciprocating. It is preferred over reciprocating engine mechanism and cam-

41

follower system to achieve true sinusoidal motion (Dixon, 2007). Moreover, the scotch

yoke mechanism has lower height and gives less vibration. Accelerometers are used to

measure movement at both the ends and load cells are used to record the damping force.

The purpose of this testing is to validate vibration attenuation and road holding ability

of the shock absorber. Test rig proposed by Peng et al., (2014) shown in Figure 2.13,

consists of the damper driven by upper reciprocating mass through a hydraulic actuator.

Lower end of the damper is connected to the frame and held stationary. Upper mass

moves through sinusoidal excitations of 10 mm amplitude at 1 Hz frequency. Load cell

has been used to record the damping force. Later, hysteresis loop has been plotted to

study energy dissipating characteristics for different damper settings.

Figure 2.13 : Shock absorber test set up (After Peng et al., 2014)

Ping (2006) and Lambert (2004) evaluated motion and force transmissibility of

the shock absorber with experimentations on an electro-dynamic shaker, shown in

Figure 2.14. Lower end of the shock absorber was mounted on the shaker table, whereas

the upper end was supporting a mass block. Accelerometers were mounted on the

shaker table and mass block with data acquisition system for recording the sensor

signal. Authors have investigated displacement transmissibility and acceleration

transmissibility of the prototype to evaluate actual performance or working

characteristics of the shock absorber.

42

Figure 2.14 : Test set up with an electro-dynamic shaker (After Ping, 2006)

Talbott and Starkey (2002) performed experimental evaluation of shock

absorber with test bench, shown in Figure 2.15. The arrangement comprises of

hydraulic linear actuator, load cell, displacement transducer and data acquisition system.

Force-displacement and force-velocity curves were evaluated at various excitation

frequencies to check accuracy of the shock absorber numerical model.

Figure 2.15 : Quarter car test rig (After Talbott and Starkey, 2002)

Theoretical analysis of wire gauze fluid damper has been performed with

Runge- Kutta method (Ping et al., 2006). Authors observed that the wire gauze

construction effectively attenuates vibrations. Moreover, it also has strength to sustain

violent impact. Two degree of freedom numerical model evaluated nonlinear stiffness

and damping of the wire gauze, throttle fluid force, flow inertia, friction and spring

43

force to compute performance parameters of the shock absorber. The theoretical model

analyzed coupling between fluid and dry friction damping for accurate simulation

results. Lumped mass formulation has been used to build numerical model of the shock

absorber with an additional bypass valve (Czop et al., 2011). Authors evaluated

throttling losses in the piston and bypass valve to find the fluid damping force.

Numerical model of the damper is then validated by testing performed on servo

hydraulic tester for acceleration transmissibility, displacement transmissibility and force

measurement. Geluk (2005) evaluated effect of dry friction on performance of fluid

shock absorber with an analytical model. Single degree of freedom test set up has been

used to measure damping force and acceleration, within 1.5-3 Hz frequency range.

Reineh (2012) validated dynamic model of a racing car shock absorber with single

degree of freedom testing on a dynamometer for displacement, force and pressure

measurement.

Guglielmino et al. (2008) performed the experiments to measure frequency

response of sprung mass acceleration with harmonic vertical excitations applied to the

tire. The experiment results are reported in Figure 2.16. It can be observed from Figure

2.16 that the acceleration increases up to certain frequency and then reduces slightly.

Thereafter, the acceleration further increase with frequency. In another work, Demic et

al. (2002) presented comfort analysis curve shown in Figure 2.17.

Figure 2.16: Acceleration frequency response to sinusoidal

displacement input to the wheel (After Guglielmino et al., 2008)

44

Figure 2.17 : Comfort curve in vertical direction (After Demic et al., 2002)

It is observed from Figures 2.16 and 2.17 that acceleration increases up to

certain value of frequency and then slope of the curve reduces. Thereafter, the

acceleration increases with frequency. A typical acceleration Frequency Response

Function curve (FRF) has peaks and valleys. The valleys in the FRF correspond to anti-

resonance frequencies (Lien and Yao, 2000). It is observed that acceleration amplitude

reduce at the anti-resonance frequencies (Harsha, 2015).

2.6 Mathematical Modeling and Simulation

Mathematical model is simplified analytical or numerical representation, which

is used to study characteristics of the system (Bender, 1977; Nguyen, 2010). The model

has defined input or independent variables and it seeks to compute output or dependent

variables with some simplifying assumption. Equations that range from simple algebraic

to higher order ODEs are used in formulating the model.

Numerical modelling is used for prediction and analysis of mechanical systems

(Desai & Kaware, 2012; Rao, et al. 2012; Liang et al., 2009). Numerical simulations

have been performed on a vibration control system comprising of MR damper and

shape memory alloy (Aravindhan and Gupta, 2006, 2010). Authors evaluated single

degree of freedom FE model in Matlab to calculate the damper force. It was reported

that higher current in MR damper coil gives considerable reduction in motion

transmissibility. Shelke and Venkatachalam (2011a, b) extended analytical simulations

to optimize copper loss in an electromagnetic bearing. All types of losses have been

45

analyzed including the copper loss to find the air gap width for optimum performance of

the system.

Prabhakar et al., (2001) extended FE method to model a bearing system passing

through its critical speed. Time response of the system has been revealed by defining

the coupling as a frictionless joint with stiffness and damping. A set of steady state and

dynamic equations have been used for theoretical formulation of lubricated ball

bearings (Sarangi et al., 2004). Authors have used finite difference method for

calculating load capacity, damping coefficient, stiffness and pressure distribution in the

bearing. Singh and Gupta (1994) presented FE formulation of a composite beam to

analyze bending mode vibrations. Performance characteristics of non-recessed hybrid

journal bearing have been computed with numerical modeling using FE formulation

(Awasthi et al., 2007). A general procedure for design of piezoelectric smart structure

for use in active vibration control has been developed by Chhabra et al., (2011).

Numerical modeling of the vibrating system is used to find optimum location of the

piezoelectric mass on the vibrating body. Arora et al., (2011) applied FE numerical

analysis in design of mechanical components. Bhushan et al. (2002) performed dynamic

analysis to study stability of a pressure dam bearing.

A nonlinear dynamic model of the fluid shock absorber is constructed with

analysis of internal fluid dynamics (Yang et al., 2007). The mathematical model

calculated total force by evaluating damping contribution by throttling of the fluid, flow

inertia, structural damping and frictional force. Later the model has been used to

evaluate effect of key parameters like oil viscosity, damping area and excitation

amplitude on the shock absorber performance. Talbott and Starkey (2002) evaluated

piston valve fluid flow and shim stack stiffness characteristics to derive numerical

model of the fluid damper. Newton’s method has been used to solve number of non-

linear governing equations with Matlab computer code. Li et al., (2009) performed

electromagnetic analysis of a biomechanical energy harvester. The analytical model has

been used in deriving configurations to achieve significant energy recovery from human

motions.

Waters et al., (2009) adopted single degree of freedom numerical model to

evaluate an automobile shock absorber. Authors have checked potential benefits of

changing damping level for reducing peak acceleration transmitted to the sprung mass.

For impulsive force, significant reduction in acceleration was achieved with lower

damping level. Analytical model of the variable shock absorber with an adjustable

46

bypass valve has been implemented in Matlab Simulink (Park et al., 2005). Solenoid

electromagnetic valve, spool plunger and fluid flow sub model have been integrated in

the Matlab model to determine pressure differential across the piston and damping

force. The derived model has been verified with experimentation on a prototype at 0.3

m/s piston speed. Clancy (1996) derived mathematical model for a current controlled,

permanent magnet driven electromagnetic shock absorber. Author have used single

degree of freedom quarter car model for evaluating the shock absorber. Matlab program

has been used to estimate the actuator current for desired control approach. A shock

absorber has been represented with MODELICA software by Hou et al., (2011).

Governing equations have been solved analytically to calculate flow through the piston

valve, pressure differential and damping force. Physical performance of a racing car

shock absorber has been investigated with one dimensional AMESim numerical

simulation tool (Reineh, 2012). Model of the shock absorber has been developed by

integrating mechanical, hydraulic and electronic subsystems to reveal effect of each

subsystem. AMESim has custom blocks for hydraulic elements like orifice and check

valves. Pressure drop across the fluid elements has been evaluated to calculate damping

force of the shock absorber. Effectiveness of an external adjusting valve to control

damping characteristics has also been investigated with the theoretical model.

Palm (2013) analyzed dynamics of a vehicle suspension with physical modeling

in Matlab Simscape. Authors have demonstrated quarter car model to study effects of

change in suspension parameters on dynamics of the system. Kazemi and Jooshani

(2012) evaluated quarter car model of a car suspension with Matlab Simscape model.

Incorporation of Stewart robot in the suspension has been investigated for comfort and

road holding ability. Tandel et al. (2013) presented multi body dynamic analysis of a

double wishbone suspension. Physical modeling of the system has been performed in

Matlab Simscape. Matlab Simscape is used to solve dynamic vibration analysis problem

involving a spring, mass and damper (Estandiari, 2014).

Lepikhin et al. (2014) used Matlab Simscape to model two linked robot

manipulator. Authors have presented kinematic and dynamic analysis to derive control

law for the system. Six degree of freedom robot manipulator has been investigated for

kinematic and dynamic behaviour (Naidin et al., 2011). The manipulator model

comprises of mechanical components, revolute joints and hydraulic elements. Simscape

has been used for kinematic and dynamic analysis of gear box and drive train systems

with rotary components (Dhupia et al., 2013; Enocksson, 2011).

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2.7 Equation Solving

Quarter car lumped parametric model involves linear second order differential

equations. Exact solution of this model is not possible, since variable being solved is

implicit in these equations. Instead, these equations can be solved with numerical

algorithms like Runge-Kutta method. Complex numerical model of the fluid damper

developed by Eyres et al., (2005) uses fourth order Runge-Kutta method. ODE also can

be easily solved with commercial software like Matlab, Modelica or Maple. Kulkarni et

al., (2012) implemented shock absorber model in Radioss block explicit solver.

Arbitrary Lagrangian- Eulerian formulation has been used to evaluate fluid flow,

pressure differential and valve displacement. It was possible for the authors to use

output from the derived model as initial guidelines in selecting the valve configuration.

Sung and Hsiang (2008) used a commercial non linear analysis program SAP2000N for

simulation of energy dissipation in the displacement dependent fluid damper. The fluid

damper comprised of hydraulic jack, check valve, relief valve and throttle valve.

Physical model of the shock absorber was defined in SAP2000N with connections

between spring, viscous dashpot and friction elements. Florin et al., (2013) analyzed

comfort and handling performance of a vehicle shock absorber with quarter car model

in Matlab Simulink. The two degree of freedom model was evaluated with step input to

calculate tire deflection and car body displacement. Authors have observed that the

simulation results are in close agreement to that of the output from state space and

transfer function numerical models. More advanced software like Matlab Simscape and

AMESim have mechanical and hydraulic design libraries that reduces modeling time

and gives greater flexibility in the design. Polach and Hajzman (2010) investigated

multibody model of semiactive damper with Alaska simulation tool.

Awasthi et al., (2006) applied Newton-Raphson method for FE formulation to

solve Reynolds’s equation, governing flow of the lubricant. Due to scarcity of published

data, numerical model of the system has been developed using Matlab® 12. Simulations

on this model are then used to derive bearing specifications for desired performance

characteristics.

Numerical integration methods are widely used for the range of applications

including automotive suspension analysis. Numerical analysis involves estimation for

vibration analysis parameters with certain time step. Selection of the time step is very

48

crucial for accuracy of the solution. The critical value of the time step depends on

maximum frequency of interest and is given as (Sinha, 2014; Roa, 2011);

max

1

10cr

tf

∆ = (2.2)

where fmax : maximum frequency of interest

Numerical simulation time step should be less than or equal to the critical time

step. It is recommended to choose the analysis time step from a well defined interval,

which should not be too small or large. Maximum frequency of interest for the

suspension analysis is about 10 Hz (Dixon, 2007; Engelhardt, 1999). Therefore

according to Equation 2.2, simulation time step for vehicle suspension analysis is taken

in terms of mili-seconds (Collette and Preumont, 2010).

Engelhardt (1999) performed dynamic analysis for motion and acceleration

transmissibility of a vehicle suspension. Maximum frequency of interest in the analysis

is 10 Hz and the recommended minimum time step is 0.01 s. Predictive control model

proposed by Cseko et al. (2011) incorporated semi-active suspension control

methodology with simulation time step of 0.010 s. Giorgetti et al. (2006) have used time

step of 0.010 s for quarter car numerical analysis of semi-active suspension system.

High speed train bogey suspension analysis is performed with time step of 0.010 s

(Herrero, 2013).

2.8 Gaps identified

In the last sections, complete review on construction, design, analysis and

testing of regenerative electromagnetic shock absorber have been presented. Also

theoretical modeling of automotive shock absorbers, fluid dampers and stroke

dependent damper have been discussed. Based on the literature surveyed, following

conclusions can be drawn.

1. With use of linear generator as the harvesting/dissipative element in vehicle

shock absorber, frequency of the generated voltage waveform is different from the

excitation frequency (Zuo et al., 2010). This leads electromagnetic damping force

frequency to differ from the excitation frequency, which gives ON-OFF effect on the

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electromagnetic damping force (Oprea et al., 2012). However many authors have not

investigated its effect on comfort and handling of the vehicle.

2. Damping force of electromagnetic devices significantly depends on the electrical

load (Amati et al., 2011). Harvested energy from a passive regenerative shock absorber

with linear generator can be used for some useful application (e.g. Automobile battery

charging). However, in this case damping coefficient of the shock absorber will

significantly change with variation in the electrical load resistance, resulting in

inconsistent damping performance.

3. Possibility of zero damping force in case of armature winding or supporting

hardware failure is one of the major drawbacks of existing systems. These aspect

necessitates use of fluid damper as additional dissipative element in a regenerative

shock absorber. Design damping coefficient of fluid damper that can be used along with

the linear generator has not been discussed by the researchers.

4. Existing electromagnetic shock absorbers have limited feasibility for application

in a vehicle suspension because of higher mass and bulky structure (Goldner and

Zerigian, 2005; Fang et al., 2013). Attempts are being made to build lighter energy

harvesting shock absorbers with use of rotary generator, which includes use of ball

screw arrangement or rack and pinion to convert linear motion in to rotary motion (Li et

al., 2013; Amati et al., 2011). For increasing transmission ratio number of helical and

bevel gears are also included in the system. Balls screw harvester gives poor

performance at higher frequencies (Zuo and Zhang, 2013). Presence of backlash and

possibility of cracks seriously affects reliability and durability of regenerative shock

absorbers with mechanical gears. Also use of number of transmission elements along

with gears affects dynamics of the system, increases cost and limits harvesting

efficiency.

5. Much research has been carried out on development of self sustaining active

suspension system. However, these systems are complex and involve costly sensors and

supporting electronics hardware. Moreover, most of the harvested energy is consumed

in applying active force by the actuator.

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6. For a fluid damper, flow clearance area and cracking pressure of the valves

decide damping coefficient, which is fairly linear (Eyres et al., 2005). However, effect

of armature coil parameters and electrical load on mechanical non-linearity of the

electromagnetic damping coefficient still needs to be studied.

7. Primary purpose of vehicle shock absorber is better vehicle comfort and

handling. However, in using regenerative electromagnetic suspension, there is conflict

between suspension performance and energy harvested (Zhang et al., 2013). In this

regard, some justification is required for loss of vibration control to the amount of

electrical energy harvested by the regenerative electromagnetic shock absorber.

2.9 Objectives of Research Work

Objectives and scope of the present study have been defined based on gaps

observed in the literature review. Development of the regenerative electromagnetic

shock absorber will be undertaken with consideration of presently available resources.

Objectives of the present work are to develop a regenerative electromagnetic shock

absorber, which will harvest significant amount of energy for normal running conditions

of the vehicle. Moreover, the proposed system will work without power driven actuators

or sensors. Study deal with design, analysis and performance evaluation of the proposed

system.

The approved objectives of the present research work are as follows:

1. To study theoretical aspects of electromagnetic damping.

2. To study theoretical aspects of displacement sensitive fluid damping.

3. To build shock absorber that will economically harvest the energy lost in

vehicle suspension without compromising safety and comfort criteria.

4. To validate theoretical study of electromagnetic and fluid damping.

5. To perform theoretical analysis of the proposed shock absorber using

quarter car model.

6. To validate the above theoretical analysis by testing a suitably scaled

prototype of the shock absorber on test rig.

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2.10 Scope of Research Work

Scope of the research work is outlined as:

1. As discussed in the literature review, a regenerative shock absorber combining

electromagnetic and fluid damping has not been studied. Hence, the present work deals

with design and development of new hybrid electromagnetic hydraulic shock absorber.

Moreover, velocity amplification has also been incorporated in the system, which

significantly increases power output from the presented device.

2. Linear generator has been used as the harvesting element in proposed shock

absorbers. A dynamic model of the linear generator is developed using FE analysis and

numerical simulation. Essential electro-mechanical characteristics such as voltage,

current and coil braking force have been evaluated. Further, experimentation has been

performed on a prototype consisting of rare earth magnets to validate the theoretical

model.

3. Harvested electrical energy of the generator will be used to charge a battery. It is

reported that, braking force on the generator coils will be effective only when the coil

voltage is greater that the battery threshold. Therefore, the linear generator will be

connected to the battery through a control circuit, which will consist of electrical

switches and a resistance, to ensure continuous braking force.

4. Next design and analysis of EMHSA-1 is discussed, which consists of a fluid

damper and electromagnetic energy harvester with rare earth magnets. Important

conclusions have been drawn based on numerical analysis supported with experimental

validation.

5. Focus of the presented research work is to develop an energy efficient

suspension/generation unit, which can replace conventional fluid shock absorbers

without increased weight penalty on the vehicles. For which, design and analysis of

EMHSA-2 is presented. Innovative feature of this device is use of mechanical links for

motion amplification and incorporation of displacement sensitive fluid damping. The

presented solution (EMHSA-2) has been designed to deliver better energy dissipation

characteristics and improved fail-safe nature. While designing the energy harvesting

shock absorber, due consideration has been given to the weight, whilst accounting

damping factor and energy harvested.

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6. Computer code has been developed that evaluates energy harvested and other

dynamic parameters of the shock absorbers. Essential performance characteristics which

have been evaluate includes motion transmissibility, acceleration transmissibility, rms

tire deflection and damper hysteresis curve.

7. Experimental test set up has been developed to measure and validate

performance characteristics of the proposed system. Accelerometer, Fast Fourier

Transform (FFT) analyzer, Load cell, Linear-Variable Differential Transformer (LVDT)

and Analog-to-Digital Convertor (ADC) has been used in the experimental set up. More

emphasis is given on comfort and vehicle handling evaluation, alongwith the power

harvested.

8. Numerical simulation results have been compared with experimental findings.

Also the investigations on numerical analysis of full scale version have been carefully

compiled and illustrated in the form of graphs. Many important conclusions have been

drawn based on the present investigation results.