Semi-active hydro-pneumatic spring and damper system

7/27/2019 Semi-active hydro-pneumatic spring and damper system

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Semi-active hydro-pneumatic spring and damper system

ABSTRACT

The basic concept of land vehicle transportation has not changed much in the last

few decades,although much progress was made in improving and optimising vehicle

design and technology. Spring and damper characteristics determine to a large extent the

ride quality and handling of a vehicle. Since the requirements for good handling and good

ride are conflicting, adjustable suspension elements are developed. In this study a two-

state semi-active hydro-pneumatic spring, in conjuction with a two-state semi-active

hydraulic damper is investigated.

Two types of tests were performed on a prototype spring-damper unit, namely

characterisation tests and single degree of freedom tests. The characterisation tests

included characterising the hydro-pneumatic spring, while for the single degree of

freedom tests, the step response were determined.

Good correlation was obtained between measured and simulated data for thecharacterisation, as well as the single degree of freedom tests. The spring damper model

can be incorporated into a full 3-D vehicle model in order to predict the ride and handling

of a vehicle fitted with such a system.

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INDEX

CONTENTS

SR. NO. NAME PAGE NO.

1

2

3

4

1.1

2.1

2.2

2.3

3.1

3.2

3.3

3.4

3.5

3.6

Introduction

Preamble

Literature and historical overview

Preamble

Need of semi-active hydro-pneumatic suspension

systems

Historical overview

Experimental work

Preamble

Experimental spring and damper unit

Experimental test setup

Hydro-pneumatic spring characterization

Single degree of freedom testing

Ride height adjustment

Heat transfer effects on hydro-pneumatic

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4.1

4.2

4.3

4.4

4.5

suspension systems

Preamble

Determination of spring characteristics

Experimental verification

Modes of heat transfer

Effect of damper heat build-up

Conclusion

References

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LIST OF FIGURES

2.1 Swiss Mowag Piranha

3.1 Experimental semi-active spring and damper unit

3.2 Characterization test setup

3.3 Single degree of freedom test setup

3.4 Floating piston accumulator

3.5 Semi-active spring characteristics (0.01m/s)

3.6 Two state damper characteristics

3.7 Step response for different combinations of spring and damper

3.8 Ride height adjustment

4.1 Experimental spring unit

4.2 Measured and predicted spring characteristics

4.3 Effect of excitation frequency

4.4 Comparision between adiabatic, isothermal and predicted spring characteristics

4.5 Effects of change in gas temperature

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1. INTRODUCTION

1.1 Preamble

The basic concept of land vehicle transportation has not changed much in the last

few decades,although much progress was made in improving and optimising vehicle

design and technology. The quest to always go faster,further and more comfortably,has

lead in recent years to the development of advanced suspension system. An improved

suspension system allows a vehicle to achieve higher speeds over rougher terrain,and

results in better handling,as well as improved ride comfort.

Passive suspension systems(suspension without controllable elements),always

represent a compramise between ride comfort and handling,since a stiff suspension is

required for good handling,while a more compliant suspension is needed for good ride

comfort. Implimenting a controllable suspension (adaptive,slow-active,fully-active) is

therefore an attempt to narrow the gap between the opposing requirements for optimal

ride comfort and handling.

In armoured fighting vehicles and other heavy off-road vehicles it is important to

have a soft suspension (low spring rate) allowing big wheel travel when negotiating

rough terrain. A soft suspension however compromises good handling and straight line

stability at high speed on good roads.

It is known that semi-active dampers can dramatically enhance ride cofort and

handling over rough terrain, still making the choice of spring rate a compromise between

ride comfort and handling. It is now possible to choose between two spring rates- one that

favours good ride comfort and another favouring good handling.

These suspension systems are mainly developed for heavy off-road vehicles like

trucks, military vehicles or passenger vehicles. By introducing a semi-active hydro-

pneumatic spring and damper system two very important suspension elements can be

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changed to suit specific driving conditions. When driving on good roads the hard spring

and damper settings can be used to improve vehicle stability and handling. By being able

to switch between a soft and a hard suspension good off-road capabilities as well as good

handling and stability can be achieved.

This study focuses on a semi-active suspension system,consisting of a two-state

switchable hydraulic damper,as well as a two-state switchable hydro-pneumatic spring.

The different elements of the spring/damper system are characterised to predict the

system performance.

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Fig 1.1 Armoured Personnel Carrier Fitted With Semi-active Damper

2. LITERATURE AND HISTORICAL OVERVIEW

2.1 Preamble

In this chapter, an overview is given of semi-active dampers and hydro-

pneumatic springs.It also focusses on literature concerned with large off-road vehicles,

but in cases where the applicable technology has not yet been demonstrated on heavy

vehicles, reference is made to commercial and passenger vehicles.

2.2 Need Of Semi-active Hydro-pneumatic Suspension Systems

The suspension systems which are used for normal cars do not suit the large off-

road vehicles, since the weight of such vehicles is much more than the normal cars. Also

even if we use such suspension systems, a large stress results on them eventually making

them of no use. So to overcome this difficulty semi-active hydro-pneumatic suspension

systems using spring and damper were invented.

Semi-active dampers were conceptualized in the 1970s and numerous

configurations and control strategies were simulated and tested since then. Semi-active

suspension systems greatly influence the vehicle dynamics (ride comfort and handling).

This is also the reason for developing semi-active suspension systems, namely to improve

ride comfort without compromising handling and stability, by switching between hard

and soft spring/damper characteristics.

2.3 Historical Overview

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Hydro-pneumatic suspensions have been introduced on battle tanks in the 1950s.

The first hydro-pneumatic struts were fitted to a prototype tracked vehicle, as result of

research done by two German companies, Frieseke and Hopfner from Erlangen and

Borgwald from Bermen into the use of compressible fluids in suspension systems

(Hilmes 1982). Since then, several other military vehicles were fitted with hydro-

pneumatic suspensions, but most of them did not go into production due to reliability

problems and short life span of the mechanical components. Initially, confidence in this

type of suspension was low, due to sealing and design problems.

But since the introduction of more reliable sealing techniques, hydro-pneumatic

springs and dampers have become more popular and are occasionally used in passenger

cars and large off-road vehicles. This type of suspension system is popular due to its non-

linear characteristics and versatility.

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Hydro-pneumatic suspensions are not commonly used on commercial vehicles

due to high capital cost involved.

The first production tracked vehicle fitted with a hydro-pneumatic suspension was

the Swiss Strv-103 Main Battle Tank (MBT). This vehicle was fitted with a rigidly

mounted weapon and the height adjustable hydro-pneumatic suspension was used to tilt

the vehicle upward or downward (Hilmes 1982).

Fig 2.1 Swiss Mowag Piranha

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3. EXPERIMENTAL WORK

3.1 Preamble

In this chapter, the experimental work is presented. The experimental work can be

divided into two stages, namely characterization tests and single degree of freedom tests.

The set up, test equipment and characterization procedures for both these stages are

discussed in this chapter. Where deemed necessary, some background information is

supplied, in order to elucidate the characterization process. The test results are presented

in graphical and tabular format.

In the following paragraphs, firstly the test set up is discussed, with reference to

the hardware, software and test equipment. Secondly, the characterization of the spring

and damper are discussed. After that, the single degree of freedom tests are discussed and

finally some closing remarks are made.

3.2 Experimental Spring and Damper Unit

The high and low characteristics for both spring and damper are made possible by

channeling hydraulic fluid with solenoid valves (Fig. 3.1). The spring and damper units

consists of hydraulic strut (1), two gas filled accumulators (2 and 3), a hydraulic damper

(4) and two solenoid valves (5 and 6).

The low spring rate is achieved by compressing a large volume of gas in two

separate chambers (2 and 3). By sealing off one of the chambers (2), a smaller gas

volume (3) is compressed and a higher spring rate is achieved. Spring rates can be

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individually tailored by changing the two gas volumes. For low damping the hydraulic

damper (4) is short circuited by opening a by-pass valve (5). For high damping this valve

is closed and the hydraulic fluid is forced through the damper resulting in a high damping

force.

Fig. 3.1 Experimental semi-active spring and damper unit

3.3 Experimental Test Setup

Two experimental setups were used, one for the component characterisations

(springs, dampers and valves) and another for the single degree of freedom tests. In both

the cases, a 160 kN Schenck hydraulic actuator was used to supply the desired input.

Fig. 3.2 shows schematically the characterization setup.

The top of the strut was fixed to the rigid test frame with a locating pin, while the

bottom mounting was fixed to the hydraulic actuator. Spherical rod ends were used, in

order to eliminate any bending moments on the strut. In this setup, the required relative

strut displacement is generated by vertical actuator motion.

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Fig.3.2 Characterization Test Setup

For the single degree of freedom tests, a separate test frame was build. Fig. 3.3

shows test setup.

A lead mass of approximately 3 tons was used to simulate the sprung mass of a

vehicle, since a static wheel load of between 2.5 and 3 tons are common for military off-

road vehicles. The test frame was equipped with a set of linear bearings, guiding the

sprung mass, which was supported by lower mounting of Schenck actuator. Nitrogen

cylinder is used to fill the accumulators. The accumulators, dampers and valves were

secured on top of the lead mass. The test frame was securely fixed to the test floor, to

ensure that the SDOF setup does not fall over.

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Fig. 3.3 Single Degree Of Freedom Test Setup

3.3 Hydro-pneumatic Spring Characterization

3.3.1 Physical Attributes

There are many different types of hydro-pneumatic springs, but the basic difference liesin the way the gas and oil is separated. Some hydro-pneumatic springs have a rubber

bladder separating the gas and the working fluid, while others have a floating piston.

Both the hydro-pneumatic springs considered here are of floating piston type.

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Fig. 3.4 Floating Piston Accumulator

The static volumes of the two accumulators are 0.3l and 0.7l respectively. When

the valve is open the combined volume of the two accumulators is 1.0l and when the

valve is closed only the 0.3l accumulator is connected with the single acting cylinder.

Nitrogen gas was used as springing medium.

3.3.2 Characterization Procedure

For the hydro-pneumatic spring characterization, two stages of semi-active hydro-

pneumatic spring were characterized by subjecting the strut to a sinusoidal displacement,

of varying frequency. The excitation speed is defined as the piston speed when moving

through the static position. The excitation speed is therefore a function of excitation

frequency and excitation amplitude. The amplitude of signal was approximately 100mm.

Fig 4.5 two spring characteristics for an excitation speed of 0.01m/s. It can be seen that

two very different spring characteristics were achieved.

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Fig. 3.5 Semi-active Spring Characteristics (0.01m/s)

3.4 Hydraulic Damper Characterization

3.4.1 Physical Attributes

The damper pack used has non-linear damping characteristic, which is achieved

by a system of orifices, a sealing washer and Belville springs. This damper was mounted

statically between the strut and the hydro-pneumatic springs (fig 4.1). The damping force

is therefore supplied by resistance to fluid flow through damper pack.

3.4.2 Characterization Procedure

The damper characteristics of the two-state damper were characterized by

subjecting the strut to a sinusoidal displacement input. The actuator force was recorded

when the strut travels through the static position, where the velocity is almost constant

and a maximum. Fig. 4.6 shows the damper characteristics for both on and off states.

It can be seen that off characteristic for velocities above 0.25m/s is approximately half

that of the on state, in the compression direction (negative velocity). The damper force,

in the rebound direction, shows an almost constant damping force at high velocities. This

is because during rebound motion the driving force behind the hydraulic fluid is the

accumulators pressure and in the compression direction, the simple acting strut cylinder.

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Rebound damping can therefore not be made too high, since cavitation may occur at high

velocities.

Fig 3.6 Two State Damper Characteristics

3.5 Single Degree Of Freedom Testing

Several single degree of freedom tests were performed to evaluate the

characteristics and performance of the semi-active spring/damper system. Step response

and ride height adjustment feature were evaluated.

Step Response Input

The tests were done by subjecting the system to a step displacement input of

43mm. The sprung mass displacement, for different spring and damper combinations, are

shown below. It can be seen that for spring on condition an effective natural frequency

of approximately 1.5 Hz is achieved. For the spring off state, the natural frequency is in

the region of 1 Hz. The response for the two cases where the damper is in the off state

indicates that the system is under damped. For the damper on state, the motion is

damped out within one cycle.

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Fig 3.7 Step Response for Different Combinations of Spring and Damper

3.6 Ride Height Adjustment

The ride height of vehicles fitted with hydro-pneumatic suspensions can be easily

adjusted by adding or removing hydraulic fluid from the system. A lower ride height

results in reduced body roll, lower center of gravity, a more stable firing platform and a

lower silhouette. The ride height can also be increased when a greater ground clearance is

needed.

Ride height adjustments are usually achieved by making use of an external power

source, such as an engine driven hydraulic pump. A control system then regulates the

amount of hydraulic fluid in the system with a network of pipes and valves. The Swiss

Mowag Piranah III is an example of a vehicle fitted with such a system. The semi-active

hydro-pneumatic spring/damper system investigated in this study has the added

advantage of being able to adjust the ride height without using an external pump. The ride

height adjustment works as follows:

When valve (6) is closed, the hydraulic fluid in the accumulator (2) is effectively

removed from the system. By opening and closing the valve at the right moment an

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amount of hydraulic fluid can be stored in accumulator (2), thus varying the ride height.

To decrease ride height, valve (6) is kept closed and only opened when the pressure in

accumulator (3) is higher than in accumulator (2). This is done until the pressure in

accumulator (2) reaches a predetermined value corresponding to a specific reduction in

ride height. To increase the ride height, valve (6) is only opened when the pressure in

accumulator (2) is higher than in accumulator (3).

Fig. 3.8 Ride Height Adjustment

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4. HEAT TRANSFER EFFECTS O N

HYDROPNEUMATIC SUSPENSION SYSTEMS

4.1 Preamble

Literature on hydro-pneumatic suspension systems invariably describes the spring

force of the hydro-pneumatic spring by polytropic processes, assuming ideal gas

behaviour. This assumption is a simplification of the real situation as heat transfer effects

between the gas and its surroundings cannot be ignored. The nitrogen gas used as spring

medium cannot be treated as an ideal gas under the pressures and temperatures found in

hydro-pneumatic suspension systems. Most hydro-pneumatic suspension systems

incorporate the spring and damper into one unit for reasons of cost and packaging. This

results in undesirable temperature effects, i.e. variations in ride height and spring rate.

4.2Determination of the spring characteristic

4.2.1Classical approach for determining the spring characteristic

The classical approach for determining the spring characteristic is to assume ideal

gas behaviour and polytropic gas compression processes. Isothermal and adiabatic

characteristics are then calculated. In typical hydro-pneumatic suspension units, the

average gas temperature can vary between -20 and +200 degree Celsius, while gas

pressure varies between 2 and 110 MPa. Under these conditions, especially at pressures

higher than 30 MPa, gas compressibility can result in large errors rendering the ideal gas

assumption invalid.

4.2.2. New approach for determining the spring characteristics using real gas

behaviour and heat transfer models

To circumvent the errors introduced by the ideal gas assumption, the Benedict-

Webb-Rubin equation of state, is used to give a more accurate relationship between gas

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pressure, volume and temperature. Other similar equations of state can be used depending

on the operating range of the suspension unit.

If a gas in a closed container is compressed, the volume decreases while pressure

and temperature increase, storing energy in the process. As the gas is allowed to expand,

this stored energy is released. Energy is however lost during this process because of heat

transfer between the gas and its surroundings. When the gas is compressed, the

temperature and pressure increase and heat is transferred from the gas to the

surroundings. When the gas is expanded, the gas temperature drops and heat is

transferred from the surroundings to the gas.

At very low compression speeds, enough time is available for heat transfer and

the gas temperature stays constant, giving the isothermal characteristic. At high speeds,

very little time is available for heat transfer and the gas temperature varies, resulting in

the adiabatic characteristic. At speeds between these two extremes, the spring

characteristic forms a hysteresis loop representing the energy loss in the cycle. This effect

has been noted in many publications but is usually attributed to friction and fluid losses in

the spring system. Many attempts were made to describe this effect by adopting speed or

frequency dependant polytropic exponents which can be as high as 1.7. This approach is

often used in hydraulic accumulator calculations.

Given below is a differential equation which describes the heat transfer between

the gas and the environment. Eq. (1) is solved using a fourth order Runge-Kutta method.

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It can be seen from the definition of the thermal time constant that it varies during

the cycle as the wall area (Aw) varies due to piston motion. The convection coefficient

(h) also varies due to the change in speed of the gas over the inside surface of the

cylinder. However, a constant value of the time constant fits experimental data very well

and the analysis is fairly insensitive to its value.

4.2.3. Determination of the thermal time constant

The thermal time constant can be determined experimentally, or calculated from

heat transfer models based on empirical data. The average thermal time constant is

determined experimentally by subjecting the spring to a step displacement input while

monitoring gas temperature, gas pressure or spring force. The thermal time constant is

defined as the time needed for the pressure, temperature or force to decrease by 63% of

the difference between the peak and final values. Temperature data is more appropriate to

use but difficult to measure. In the region where the ideal gas assumption is valid, the

same result will be obtained from pressure, force or temperature measurements. Because

the analysis is fairly insensitive to the value of the time constant, pressure measurements

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are used to determine the thermal time constant. Good correlation is achieved between

measured and predicted values for the time constant as indicated in Table 1.

4.3 Experimental verification

4.3.1 Laboratory tests on experimental spring unit

An experimental hydro-pneumatic spring unit was designed and manufactured to

verify predicted results. The unit was designed for a static wheel load of 3000 kg and a

stroke of 250 mm. The basic design of the experimental spring unit is shown in Fig. 1.

The unit was fitted with thermocouples to measure gas and oil temperatures, and pressure

transducers to measure gas and oil pressure.

Schenck hydropuls equipment was used to excite the spring with sinusoidal

displacement inputs of various amplitudes and frequencies. Spring force, displacement,

gas and oil pressures and temperatures were recorded.

Fig. 4.1 Experimental Spring Unit

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4.3.2. Comparison of experimental and predicted results:

The isothermal spring characteristic was determined by compressing the spring

slowly in discrete increments, leaving enough time for the temperature and pressure to

stabilize at each displacement interval. Correlations between measured and predicted

values are excellent.

Fig. 2 shows the measured and predicted spring characteristics for a sinusoidal

excitation with a frequency of 0.1 Hz and amplitude of 80 mm. It can be seen that the use

of Eq. (1) to predict the dynamic spring force is in close correlation with measured data.

Fig. 3 shows the effect of different excitation frequencies on the spring characteristic. It

can be seen that the lower the excitation frequency, the smaller the hysteresis loop

becomes. This means that the energy loss and thus the damping in the cycle becomes

less. Isothermal compression is approximated as there is more time available for heat

transfer between the gas and its surroundings. At higher excitation frequencies, the

hysteresis loop also becomes smaller as adiabatic compression is approximated and less

time for heat transfer is available.

The hysteresis loop means that hydro-pneumatic suspensions have an amount of

inherent damping which is dependent on the excitation frequency (or speed) just like

hydraulic dampers. The amount of damping however decreases as the excitation

frequency increases and adiabatic conditions are reached. This is another advantage of

hydro-pneumatic suspensions which may be worthwhile to utilize.

E.g. trying to manipulate the thermal time constant to get the maximum damping

at the resonant frequency of the suspension system as this damping generates no heat

because the net heat flux is zero.

Fig. 4 shows the adiabatic and isothermal spring characteristics compared to the

predicted dynamic spring characteristic. It is clear that assuming isothermal or adiabatic

behaviour can result in large errors. The presence of inherent damping may seem

insignificant, but becomes an important factor when performing accurate dynamic

simulation of vehicles to improve ride comfort or noise, vibration and harshness levels. It

must also be taken into account when developing semi-active damper systems because

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the inherent damping will limit the minimum value of the off characteristic which usually

needs to be as low as possible.

Fig. 4.2 Measured and Predicted Spring Characteristics

Fig.4.3 Effect of excitation frequency

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4.4. Heat transfer modes

The aim up to now was to predict and verify the nature of the spring characteristic

which only involved heat transfer between the gas and the environment. If a complete

hydro-pneumatic spring system is analyzed, the effects of heat generation in the integral

damper should be included in the analysis.

Heat transfer takes place by means of conduction through the floating piston and cylinder

walls and convection between the gas, cylinder walls and the environment.

4.5 Effects of damper heat build-up

The energy dissipated in the damper can be calculated by integrating the damperforce multiplied by the relative velocity across the damper.

The damper characteristic was measured for static gas pressures of 6, 8 and

10MPa. Changing the gas pressure affects the maximum rebound damper force which

can be achieved because the damper starts to cavitate when the pressure difference across

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the damper is higher than the gas pressure. This will cause the oil to aerate which will

severely affect damper performance.

Fig. 4.6 shows the ambient, oil, gas and wall temperature histories for an

excitation frequency of 0.5 Hz at amplitude of 60 mm. It can be seen that due to heat

generation in the damper, the oil temperature rises until termination of the tests at 130

degree Celsius. The wall temperature follows the oil temperature trend but the gas

temperature rises to only 70 degree Celsius. After the test is terminated, the gas

temperature continues to rise indicating that heat transfer between the oil and gas is very

slow. Conduction through the steel cylinder wall is much greater than the convection

coefficient between the wall and the surroundings, therefore the lumped capacitance

method is used to analyze the heat transfer coefficients. Fitting curves to determine the

thermal time constant between the oil and the surroundings after termination of tests

(when heat input from the damper is zero) yields time constants of between 3000 s at a

temperature difference of 100 degree Celsius and 6500 s at a temperature difference of

less than 20 degree Celsius. The effect of the rise in gas temperature (from 23 to 68

degree Celsius) on the combined spring and damper force can be seen in Fig. 4.5. The

spring rate increases by 58% while the static ride height would increase by 15% (18 mm)

after 75 min of testing.

Vehicle tests were performed over typical terrains at representative speeds. After

45 min of continuous driving, the damper oil temperature stabilized at 85 degree Celsius.

The gas temperature showed no increase at all and the ride height variation was zero. The

test track included sections of very smooth concrete road over which the heat generation

in the damper can be ignored. Thermal time constants of between 150 and 450 s were

measured on the smooth parts of the track. These are an order of magnitude lower than

that measured in the laboratory, indicating that heat generation in the damper is closely

related to terrain roughness and that heat transfer between the spring unit and its

surroundings is very much dependant on airflow over the suspension units.

The specific design of the test unit with good thermal separation between the gas

and damper oil, together with air flow across the unit, results in a hydro-pneumatic spring

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configuration which does not suffer from severe temperature effects as often encountered

on other systems.

Fig. 4.5 Effect of change in gas temperature

Fig. 4.6 Temperature Histories

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CONCLUSION

Following conclusions can be made from the above report:

It is possible to have a suspension with a high and a low spring rate, optimized for

both ride comfort and handling.

Non-linear characteristics are very useful, eliminating the necessity for a bump

stop when the hard spring setting is selected.

A method for easily altering the vehicle ride height without any additional costs

or further complicating the system was illustrated.

This study in short introduced a suspension element which can be altered

continuously to suit the terrain and driver input demands.

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REFERENCES

6.1 Journal Papers

Semi-active hydro-pneumatic spring and damper system : Els. P.S. &

Giliomee C L , 1998 Journal of Terramechanics 35, 109-117.

Heat transfer effects on hydro-pneumatic suspension systems : Els. P.S. & B.

Grobbelar , 1999 Journal of Terramechanics 36, 197-205.

The ride comfort vs. handling compromise for off-road vehicles : Els. P.S.,

N.J.Theron , M.J.Thoresson & P.E.Uys , 2007 Journal of Terramechanics 44, 303-

317.

6.2 Books

Analysis of a four state switchable hydro-pneumatic spring and damper system.

- Christiaan Lambert Giliomee

Department of mechanical and aeronautical engineering

University of Pretoria (2003).

6.3 Web links

www.sciencedirect.com

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www.car_suspension_bible/1-4.com

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Semi-active hydro-pneumatic spring and damper system