Final Report1 2

7/28/2019 Final Report1 2

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1. INTRODUCTION

The goal of our project is to design a suspension system for formula type car. Formula

type car is a single seat off-road vehicle that is built to compete in the international SAE

student formula supra series. In order to make the car competitive the vehicle must be able to

handle a variety of terrains, such as off-road trials, jumps, rocks and hills. The primary system

that determines performance, while traveling off road, is the suspension system. The

suspension must isolate the driver from the terrain as well as maintain control of the car. The

major design constraint placed on the All Terrain Vehicle is the weight. Every car in the

S.A.E. Baja competition is provided with a 10 hp Briggs and Stratton engine. The cars are

power limited so every component must be as lightweight as possible. This criterion also

extends to the suspension design and forces detailed analysis to ensure mass is kept to aminimum.

The fundamental requirement for this project is to evaluate available suspension

configurations and choose the design that best fits the given constraints and performance

requirements. Once the design solution has been chosen, various parameters must be selected.

In order to adopt a completely new suspension design, damping and spring rates must be

chosen. The primary method to select these parameters was the quarter car model. The

structural design of the suspension was based on loads cases established by other All Terrain

Vehicles. The suspension capability and its ride handling characteristics largely depends on

the suspension kinematic linkages and the suspension kinematic parameters like the camber,

castor, toe, castor trail, king pin axis inclination and the scrub radius. The desired kinematic

parameters after much analysis was fixed and then achieved through iterative process in Lotus

Suspension Analyzer. The hard points for the kinematic linkages were also taken from the

Lotus Suspension Analyzer and then 3-D part design was made in the Pro-E based on the

Design for Manufacturing and Assembly (DFMA) principles. In order to reduce the unsprung

mass of the vehicle it was decided and designed to use Aluminum knuckles and hubs. The

The overall performance of the rear suspension was then evaluated by driving the vehicle over

a large variety of challenging terrains.


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2. LITERATURE SURVEY

Literature survey is the documentation of a comprehensive review of the published and

unpublished work from secondary source data in the areas of specific interest to the

researcher. The following literature surveys were made in this project.

2.1 SUSPENSION SYSTEM

Suspension is the term given to the system of springs, shock

absorbers and linkages that connects a vehicle to its wheels and allows relative motion

between the two. Suspension systems serve a dual purpose contributing to the vehicle's

road-holding/handling and braking for good active safety and driving pleasure, and keeping

vehicle occupants comfortable and reasonably well isolated from road noise, bumps, and

vibrations, etc. These goals are generally at odds, so the tuning of suspensions involves

finding the right compromise. It is important for the suspension to keep the road wheel in

contact with the road surface as much as possible, because all the road or ground forces acting

on the vehicle do so through the contact patches of the tires. The suspension also protects the

vehicle itself and any cargo or luggage from damage and wear.

2.2 TYPES OF SUSPENSION SYSTEMSSuspensions generally fall into either of two

Groups-solid axles(Dependent Type).

Independent type Suspensions System.

Each group can be functionally quite different, and independent suspension systems

have many advantages when compared to the solid-axles (i.e., dependent suspension systems).

2.2.1 SOLID AXLE SUSPENSION SYSTEMSIn solid axle suspension systems, wheels are mounted at the ends of a rigid beam so that

any movement of one wheel is transmitted to the opposite wheel causing them to steer and

camber together.


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Solid drive axles are used on the rear of many cars and most trucks and on the front of

many four-wheel-drive trucks. Solid beam (non-driven) axles are commonly used on the front

of heavy trucks where high load-carrying capacity is required.

Solid axles have the advantage that wheel camber is not affected by body roll. Thus

there is little wheel camber in cornering, except for that which arises from slightly greater

compression of the tires on the outside of the turn. In addition, wheel alignment is readily

maintained, minimizing tire wear.

The major disadvantage of solid steerable axles is their susceptibility to tramp-shimmy

steering vibrations. The most common solid axles are Hotchkiss, Four link and De Dion.

2.2.2 INDEPENDENT SUSPENSION SYSTEMSIn contrast to solid axles, independent suspensions shown in the figure 2.1 allow each

wheel to move vertically without affecting the opposite wheel. Nearly all passenger cars and

light trucks use independent front suspensions, because of the advantages in providing room

for the engine and the better resistance to steering vibrations.

The independent suspension also has the advantage that it provides inherently higher

roll stiffness relative to the vertical spring rate. Further advantages include easy control of the

roll centre by choice of the geometry of the control arms, larger suspension deflections, and

greater roll stiffness for a given suspension vertical rate.

Figure 2.1 Independent Suspension Systems


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Over the years, many types of independent front suspension have been tried such as

Macpherson, Trailing arm, Swing axle, Multi link and Double wishbone suspension. Many of

them have been discarded for a variety of reasons, with only two basic concepts, the double

wishbone and the Macpherson strut, finding widespread success in many varied forms.

2.3 FUNCTIONS OF SUSPENSION SYSTEM

Road isolation - to absorb or isolate road shock from the passenger compartment.

Directional stability- to maintain the vehicle in a directed path.

Returnability - to return the front wheels to straight ahead after turning.

Tracking - the path taken by the front and rear wheels. Cornering - the ability of the vehicle to travel a curved path.

Maintain correct vehicle ride height.

Reduce the effect of shock forces.

Maintain correct wheel alignment.

Support vehicle weight.

Keep the tyres in contact with the road.

Control the vehicle's direction of travel.

2.4 GEOMETRY PARAMETERS INVOLVED IN SUSPENSION

The suspension geometry on the whole greatly determines the performance of the

suspension system. It determines the steering effort needed, the straight-line drivability, the

traction force generated, cornering tendency, etc,

2.4.1 WHEEL CAMBER-CAMBER ANGLE

Camber angle is regarded as the inclination of the wheel plane to the vertical as shown

in figure 2.2. Negative camber inclines the top of the tire toward the centerline of the vehicle

and positive camber inclines the top of the tire away from the centerline.


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Figure 2.2 Camber angle

A small amount of negative camber of up to 1.5 degrees it is recommended in order to

induce camber thrust. However, changes in camber should be kept at minimum during chassis

roll in order to reduce the loss of camber thrust and the change in wheel track load distribution

during cornering. Camber angle alters the handling qualities of a particular suspension design;

in particular, negative camber improves grip when cornering. This is because it places

the tire at a better angle to the road, transmitting the forces through the vertical plane of the

tire rather than through a shear force across it. Another reason for negative camber is that a

rubber tire tends to roll on itself while cornering. Negative camber can also be caused byexcessive weight on the front wheels. The inside edge of the contact patch would begin to lift

off of the ground if the tire had zero camber, reducing the area of the contact patch. This effect

is compensated for by applying negative camber, maximizing the contact patch area.

On the other hand, for maximum straight-line acceleration, the greatest traction will be

attained when the camber angle is zero and the tread is flat on the road. Proper management of

camber angle is a major factor in suspension design, and must incorporate not only idealized

geometric models, but also real-life behavior of the components; flex, distortion, elasticity, etc

Rate of camber change-The rate of camber change is the change of camber angle per

unit vertical displacement of the wheel center relative to the sprung mass.


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2.4.2 WHEEL CASTER-CASTER ANGLE

Caster angle is the angle in side elevation between the steering axis and the vertical as

shown in figure 2.3. It is considered positive when the steering axis is inclined rearward (in the

upright direction) and negative when the steering axis is inclined forward.

Figure 2.3 Caster angle

Positive caster induces a self correcting force that provides straight line stability, but

increases steering effort. Caster ranges from approximately 2 to 7 degrees in racing vehicles .

Positive caster tends to straighten the wheel when the vehicle is traveling forward, and

thus is used to enhance straight-line stability. The mechanism that causes this tendency is

clearly illustrated by the caster front wheels of a shopping cart (above). The steering axis of a

shopping cart wheel is set forward of where the wheel contacts the ground. As the cart is

pushed forward, the steering axis pulls the wheel along, and since the wheel drags along the

ground, it falls directly in line behind the steering axis. The force that causes the wheel to

follow the steering axis is proportional to the distance between the steering axis and the wheel-

to-ground contact patch-the greater the distance, the greater the force. This distance is referred

to as "trail."

Due to many design considerations, it is desirable to have the steering axis of a car's

wheel right at the wheel hub. If the steering axis were to be set vertical with this layout, the


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axis would be coincident with the tire contact patch. The trail would be zero, and no caster

would be generated

It is possible to create caster by tilting the steering axis in the positive direction. Withsuch an arrangement, the steering axis intersects the ground at a point in front of the tire

contact patch, and thus the same effect as seen in the shopping cart casters is achieved.

The tilted steering axis has another important effect on suspension geometry. Since the

wheel rotates about a tilted axis, the wheel gains camber as it is turned. This effect is best

visualized by imagining the unrealistically extreme case where the steering axis would be

horizontal-as the steering wheel is turned, the road wheel would simply change camber rather

than direction. This effect causes the outside wheel in a turn to gain negative camber, while

the inside wheel gains positive camber. These camber changes are generally favorable for

cornering, although it is possible to overdo it.Three to five degrees of positive caster is the

typical range of settings, with lower angles being used on heavier vehicles to keep the steering

effort reasonable.

Rate of caster change-The rate of caster change is regarded as the change in caster

angle per unit vertical displacement of the wheel centre relative to the sprung mass.

2.4.3 KINGPIN INCLINATION

The angle in front elevation between the steering axis and the vertical is regarded as

kingpin inclination. It is also known as steering axis inclination (SAI).It is used to reduce the

distance measured at the ground between steering axis and tyres centre of pressure in order to

reduce the torque about the steering axis during forward motion. A right kingpin inclination

will reduce the steering effort and will provide the driver with a good road feel.

2.4.3.1 KINGPIN OFFSET

Kingpin offset measured at the ground is the horizontal distance in front elevation

between the point where the steering axis intersects the ground and the centre of tyre contact

as shown in figure 2.4. Kingpin offset it is also known as scrub radius. It is positive when the


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centre of tyre contact is outboard of the steering axis intersection point on the ground. Kingpin

offset is usually measured at static conditions (zero degree camber).

Figure 2.4 Kingpin inclination

The kingpin offset at the wheel centre is the horizontal distance in front elevation from

the wheel centre to the steering axis.

2.4.4 WHEEL TOE-STATIC TOE ANGLE

Static toe angle as shown in figure 2.5 is measured in degrees and is the angle between

a longitudinal axis of the vehicle and the line of intersection of the wheel plane and the road

surface. The wheel is toed-in if the forward position of the wheel is turned toward a central

longitudinal axis of the vehicle, and towed-out if turned away.

2.4.4.1 STATIC TOEStatic toe-in or toe-out of a pair of wheels is measured in millimeters and represents

the difference in the transverse distance between the wheel planes taken at the extreme rear

and front points of the tyre treads. When the distance at the rear is greater, the wheel is toed-

in by this amount; and where smaller,the wheels are toed-out as illustrated in Figure 20.


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Figure 2.5 Toe in/out

It is necessary to set the static toe such way to prevent the tyres to become toeout

during maximum bump and roll in order to prevent the outboard tyre to steer the vehicle to the

outside of the turn when cornering. Toe-in produces a constant lateral force inward toward the

vehicle centerline during forward motion that will enhance the straight line stability.

2.5 SUSPENSION KINEMATIC PARAMETERS

The vehicles road handling ability, roll resistance depends on the instantaneous centre

location and the roll centre location and its relative motion. The oversteer can also be

promoted by having the rear roll centre comparatively lower to that of front roll centre. The

driving stability of the vehicle also depends on the kinematic parameters

2.5.1 INSTANTANEOUS CENTRE

As shown in figure 2.6, the instantaneous centre of a suspension is the point through

which an individual wheel rotates and is also referred to as the swing centre or virtual half-

shaft.. It is also the point through which the respective tyre force acts on the sprung mass.

The instantaneous centre of a four bar link independent suspension is located at the


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intersection of the lower and upper link extensions as seen in Figure 2.6.When analysing

suspension kinematics, both left and right suspensions must be analysed together.

Figure 2.6 Instantaneous centre/ roll centre

2.5.2 ROLL CENTRE

The point in the transverse vertical plane through any pair of wheel centres at which

lateral forces may be applied to the sprung mass without producing suspension roll.

The roll center of a vehicle is the notional point at which the cornering forces in the

suspension are reacted to the vehicle body.

There are two definitions of roll center. The most commonly used is the geometric (or

kinematic) roll center; the Society of Automotive Engineers uses a force-based definition.

The location of the geometric roll center is solely dictated by the suspension geometry,

and can be found using principles of the instant center of rotation. The SAE's definition of the

force based roll center is, "The point in the transverse vertical plane through any pair of wheel

centers at which lateral forces may be applied to the sprung mass without producing

suspension roll".


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The lateral location of the roll center is typically at the center-line of the vehicle when

the suspension on the left and right sides of the car are mirror images of each other.

The significance of the roll center can only be appreciated when the vehicle's center of

mass is also considered. If there is a difference between the position of the center of mass and

the roll center a moment arm is created. When the vehicle experiences angular

acceleration due to cornering, the size of the moment arm, combined with the stiffness of the

springs and anti-roll bars (anti-sway bars in some parts of the world), dictates how much the

vehicle will roll. This has other effects too, such as dynamic load transfer.

The geometric roll center of the vehicle can be found by following basic geometrical

procedures when the vehicle is static. However, when the vehicle rolls the roll centers migrate.

It is this movement of roll centers that vehicle dynamics seek to control and in most cases

limit. The rapid movement of roll centers when the system experiences small displacements

can lead to stability problems with the vehicle. The roll center height has been shown to affect

behavior at the initiation of turns such as nimbleness and initial roll control.

2.5.3 ROLL CENTRE LOCATION

The roll centre is located at the intersection of the lines formed by the tyre contact

patches and their respective instantaneous centres. The tyre contact patches are assumed at the

intersection of wheel centre line with the ground, while in reality these two points may not

coincide. The location of the roll centre is usually different for front and rear suspension. The

vertical location or height of the roll centre determines the resulting two moment arms formed

between the roll centre and both the C.G. and the ground plane. These two moment arms

determine the vehicles sensitivity to lateral acceleration by producing rollover moments and

jacking forces.

2.6 TERMINOLOGIES USED IN SUSPENSION SYSTEM

Compression: downward travel of the suspension. Actions that move the endpoints ofthe shock closer together.


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Emulsion shock: shock without an IFP (Internal Floating Piston) separating the oil andnitrogen.

Frame clearance: distance between the frame and other moving parts, like the shock.

Negative travel: distance the suspension or shock extends from the static ride height.Also referred to as free sag.

Preload: initial force on the spring. Preload is used to adjust rider sag.

Ride height: with the rider on the bike, the basic stance of the bike. Usually measuredfrom the ground to some point on the bike frame.

Rebound: force required to extend the shock or suspension. Can also refer to theextending action of the suspension.

Rider sag: amount the shock compresses with the rider sitting on the ATV in a normalriding position.

Free sag: amount that the ATV sits into travel. Usually measured from the ground toa point on the frame, or as shock stroke, and without a rider on the ATV.

Spring rate: force required to compress a spring one inch. Measured in lb/in. orKg/mm.

Travel: total amount the shock compresses, as measured from eye-to-eye.

Wheel travel: distance the wheel moves when the suspension is cycled through its fulltravel riding.

Bottoming: vehicle has bottomed-out when the suspension reaches the limit of itstravel and stops further downward motion.

Bucking: kicking motion on a rider after a bump or jump landing.

Chatter: small bumps similar to braking bumps prior to a corner or berm. Often refersto the harshness felt when riding over small, closely spaced bumps.

Fading: slow loss of shock damping usually due to heat.

Packing: when the shock does not return quickly enough to adequately absorb the nextbump in a repetitive bump sequence.

Squat: when the rear of the vehicle sits down either due to weight transfer ordriveline forces.


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3. TARGET DESIGN LEVEL SPECIFICATIONS

It is the first step in the design methodology of suspension system. The design criteria/

target specific level targets help one to set the initial values and proceed with it. The targets

like ground clearance, ride height, material selection, suspension travel and the sprung mass to

be achieved are discussed here

3.1 GROUND CLEARANCE

Ground clearance plays a major role in the design of the suspension system. An ATV

is all set to encounter huge rocks, high bumps, and huge logs. Low ground clearance will

make it extremely difficult for the ATV to traverse high bumps. Optimum ground clearance is

very much needed for an all terrain vehicle as it has to traverse high bumps. But increasedground clearance will have some adverse effects on vehicle handling as well. The increase in

ground clearance of the vehicle will increase the height of centre of gravity of the vehicle. Any

increase in the height of centre of gravity of the vehicle will leave the vehicle susceptible to

vehicle roll. The table 3.1 values give the ground clearance set for the all terrain vehicle

suspension project.

Ground Clearance at Front 13 Ground Clearance at Rear 11

Table 3.1 Ground Clearance

3.2 TO ACHIEVE FLAT RIDE CONDITION

Flat ride is the condition, at which the maximum comfort will be experienced by the

vehicle occupants. At this condition the front will vibrate at a frequency much lesser rate than

that of rear frequency which makes the vehicle to cancel out the pitch motion as quick as

possible.

3.3 RIDE FREQUENCY

The first step in choosing spring stiffness is to choose your desired ride frequencies,

front and rear. A ride frequency is the undamped natural frequency of the body in ride. Higher


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the frequency, stiffer the ride. So, this parameter can be viewed as normalized ride stiffness.

Based on the application, there are ballpark numbers to consider for different types of vehicle

as given in table 3.2.

Ride Frequency Type Of Vehicle

0.5 - 1.5 Hz Passenger cars

1.5 - 2.0 Hz Sedan racecars and moderate downforce

formula cars

3.0 - 5.0+ Hz High downforce racecars

1.00-2.0 Hz All Terrain Vehicle Cars

Table 3.2 Ride Frequency

Lower frequencies produce a softer suspension with more mechanical grip; however

the response will be slower in transient (what drivers report as lack of support). Higher

frequencies create less suspension travel for a given track, allowing lower ride heights, and in

turn, lowering the center of gravity.

Ride frequencies front are rear are generally not the same. With the front ride

frequency higher than the rear, the first period is the most dominant on the car when looking at

frequency phase, due to effects of damping evident from the figure 3.1.

Figure 3.1 Higher front frequency


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The out of phase motion between front and rear vertical motion, caused by the time

delay between when the front wheel and rear wheel hit the bump, is accentuated by the

frequency difference. A result of the phase difference is pitching of the body.

3.2.2 FLAT RIDE CONDITION

To reduce the pitch induced, which is an undesirable ride handling characteristics.

The rear needs to have a higher natural frequency to catch up with the front as shown in

figure 3.2. This notion is called producing a flat ride, meaning that the induced body pitch

from road bumps is minimized.

Figure 3.2 Higher Rear Ride Frequency

For a given wheelbase and speed, a frequency split front to rear can be calculated to

minimize pitching of the body due to road bumps. A common split is 10 20% front to rear.

Racecars in general run higher damping ratios, and have a much smaller concern for comfort,

leading to some racecars using higher front ride frequencies.

Low Frequency = Less suspension stiffness

High Frequency = High suspension stiffness

The higher damping ratios will reduce the amount of oscillation resultant from road

bumps, in return reducing the need for a flat ride. A higher front ride frequency in a racecar


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allows faster transient response at corner entry, less ride height variation on the front (the

aerodynamics are usually more pitch sensitive on the front of the car) and allows for better

rear wheel traction (for rear wheel drive cars) on corner exit. The ride frequency split should

be chosen based on which is more important on the car you are racing, the track surface, the

speed, pitch sensitivity, etc.

Figure 3.3 Rear ride frequency 10% higher than front

Front Frequency 1 Hz

Rear Frequency 1.35Hz

Table 3.3 Vehicle Frequency

The figure 3.3 shows the condition of the FLAT RIDE where the rear cancels out the

front pitching motion by having higher frequency at the rear than the front and the table 3.3

shows the desired frequency for the ATV.

3.3 CAMBER GAIN IN BUMP

For an All Terrain Vehicle Race car, camber gain is an important aspect that plays a

major role in providing the required traction force to the tire while the vehicle encounters the


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bump and corners. Traction force generates the required grip of the tire to the road and

prevents the wheel slip.

Mostly positive camber gain is preferred in passenger vehicles as it reduces the

steering effort required for cornering but reduces the traction force. But this reduction in

traction force doesnt impact the grip as the asphalt layer of the road surface will generate

sufficient friction between the tire and the road. Positive camber will lead to excessive tire

wear at the outer surface of the tire.

Negative camber gain will ensure maximum contact patch exists and thus leads to

increased frictional force between the tire and the road surface this providing the much

required traction force.

3.4 TO PROMOTE OVERSTEER

Oversteer is what occurs when a car turns (steers) by more than (over) the amount

commanded by the driver as shown in figure 3.4. Oversteer occurs when the traction at the

rear tires during the cornering becomes less thus leading to wheel slip.

Figure 3.4 Oversteer


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3.4.1 THE REASONS TO PROMOTE OVERSTEER

The off-Road car was designed with open spool drive, therefore it is essential to

promote oversteer to make the turning the corners a lot easier.

An Off-Road racing track has a lot of corners to get on with where minutest of the

time unit advantage plays an important decisive role. Oversteer will help the

vehicle to take the turn with good speed but with a little steering effort.

Accelerating early as the car passes the apex of a corner allows it to gain extra

speed down the following straight. The driver who accelerates sooner and/or harder

has a large advantage.

3.4.2 WAYS TO PROMOTE OVERSTEER:

The following activities will promote oversteer

Distribution of more weight to the vehicle rear

Lesser rear track width than that of front track width

Greater Suspension stiffness at the rear

3.5 ZERO BUMP STEER

In the design of the suspension geometry, the bump steer and roll steer effect should be

taken into account. Bump steer results from the combination of wheel toe in or toe out with

wheel vertical travel and has adverse effects on the ride characteristics. If a bump steer exists

the vehicle will be pulled towards the side where there is maximum traction and leads to the

awkward movement of the vehicle.

The roll steer is produced by the combination of toe in/out and body roll. The

suspension and steering links should be placed such way to minimise the distortion of the

steering geometry with suspension movement. For this reason, to minimise the bump steer, the

steering tie rod is placed parallel with the upper wishbone and to minimise roll steer the upper

wishbone inboard pivot points should be in the same vertical plane with tie rod inboard pivot

point.


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3.6 LOWEST COMBINED MASS POSSIBLE

The combined mass of the suspension linkages, shock absorbers, uprights, hubs and

tires (i.e., unsprung mass) must be minimal to obtain maximum performance of the

vehicle.

The unsprung mass of the vehicle must account only for 12-14% of the sprung mass of

the vehicle.

3.6.1 SPRUNG MASS AND UNSPRUNG MASS

In a vehicle with a suspension, as can be seen in the figure 3.5 sprung mass is the

portion of the vehicle's total mass that is supported above the suspension, including in most

applications approximately half of the weight of the suspension itself. The sprung weight

typically includes the body, frame, the internal components, passengers, and cargo.

The unsprung weight is the mass of the suspension, wheels, and other components

directly connected to them, rather than supported by the suspension. Unsprung weight includes

the mass of components such as the wheel axles, wheel bearings, wheel hubs, tires, and a

portion of the weight of driveshafts, springs, shock absorbers, and suspension links. Even if

the vehicle's brakes are mounted outboard (i.e., within the wheel), their weight is still

considered part of the unsprung weight. The larger the ratio of sprung weight to unsprung

weight, more is the vehicle performance and the less is the uncomfortness caused to the

vehicle occupants.

Figure 3.5 Sprung and Unsprung mass


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4.1 DESIGN METHODOLOGY

Vehicle specific level targets

Suspension type selection

Stiffness & shock selection

Suspension geometry iteration in Lotus Suspension Analyzer Design & Finite Element Analysis in ANSYS

Knuckle hard point generation

Knuckle design & analysis

4.2 SUSPENSION TYPE SELECTION:

It is the important stage in the design of suspension system has appropriate suspension

system must be selected. Independent type of suspension was selected has it has many

advantages such as

It leads to low unsprung mass.

It increases ride quality and ride handling characteristics of the vehicle.

The deflection of one wheel does not affect the other wheel

4.2.1 FRONT SUSPENSION SYSTEM-THE DOUBLE SLA WISHBONE

SUSPENSION SYSTEM

Double SLA wishbone system consists of unequal and unparallel upper and lower

control arms. Double SLA suspension system has several advantages over the other type of

suspension systems like Macpherson Strut. It provides the maximum and the easiest tuning of

suspension kinematics.

Unlike Macpherson Strut it generates negative camber gain while on bump and

cornering. The lower arm which lies closer to the wheel will create a higher moment than the

upper arm that lies at a comparative far distance from the wheel gives rise to the much needed

negative camber gain during cornering.


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Earlier two parallel A-arms of equal length was neglected as it led to wheels to lean

outboard in turns. The design also caused excessive tire scrubbing because of the large

variation in tread-width as the wheel moved off the neural position.

Figure 4.1 Front suspension Linkages

The unequal length, non-parallel A-arm system as shown in the figure 4.1 allows the

designer to place the reaction point of the wheel at virtually any point in space. The actual

position of that point (virtual reaction point) is controlled simply by moving the inboard

connection of the upper and lower A-arms up or down, or closer together or farther apart. For

example, moving the inboard connection points farther apart moves the reaction point farther

way until it reaches infinity when the arms are parallel. If the inboard connection points are

moved still farther apart, the reaction point then flips to the other side and assumes a position

A line projected from the bottom of the wheel to the virtual reaction point establishes thevehicle roll center at the point of intersection with the vertical centerline of the vehicle. The

height of the roll center is therefore controlled by varying the inboard connection points of the

upper and lower A-arms as needed to vary the height of the virtual reaction point


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Figure 4.2 Front suspension System

Anti-dive is another feature that is easily designed into the double A-arm suspension

(figure 4.2). Vehicles with a soft ride tend to dive when braking. This is due to the weight

transfer toward the front of the vehicle. The tendency to dive on braking can be partially

alleviated by tilting the upper A-arm as shown in the drawing in Figure 4.3

Figure 4.3 Anti-Dive Design

4.2.1.1 ADVANTAGES OF SLA WISHBONE SYSTEM

Easy to tune the suspension kinematics such as camber, toe in and toe out.

It provides good ride quality and road holding characteristics


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Provides negative camber gain during cornering and while encountering bump

The double A-arm uses rigid control arms to mount the knuckle to the chassis.

These arms prevent deflections during cornering which ensures that steering

and wheel alignment remains consistent.

4.2.2 MULTI LINK TRAILING ARM REAR SUSPENSION SYSTEM

Multi link trailing arm suspension system consists of control arm called trailing arm

that connects the chassis and the knuckles, and two camber links that gets attached to the

trailing arm and the chassis as shown in figure 4.4, provides the necessary camber gain.

Figure 4.4 Rear Trailing Arm

Of all the suspension system, the trailing arm system has several distinct advantages

like easy packaging and servicing of the engine bay. No other system withstands the axial


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force superior than trailing arm suspension system. Though it provides better articulation, it

does not facilitate Camber Gain, which is required to increase the traction force between the

tire and the road. The addition of camber links to the pure trailing arm system gets rid of zero

camber adjustment drawbacks. It has good load withstanding capacity than any other

suspension system. The trailing arm withstands all the axial forces that act on the suspension

system, while the lateral force that gets generated while cornering is withstood by the camber

links. There is no bending that acts on the trailing arm since the shock absorbers are not

mounted on the trailing arm, instead they are mounted on the knuckles.

4.2.2.1 ADVANTAGES OF MULTI LINK TRAILING ARM SUSPENSION

SYSTEM Easy packaging and servicing of the engine bay:

Multi link trailing arm suspension system facilitated easy packaging and servicing

of the engine bay. The shock absorbers and the control arms need not be removed when

the engine has to be serviced, as it is the case of other suspension system.

High load withstanding capacity:

The multi link trailing arm suspension system withstands most of the forces that

acts on the suspension system. The trailing arm withstands the axial forces (the

longitudinal forces) that act on the suspension system while the camber links withstands

the lateral forces due to cornering.

Generates negative camber gain:

One of the disadvantages of the pure trailing arm is that it does not generate any

camber which is very much essential to generate the required traction force while

cornering and bumping. The addition of camber links to the trailing arm generates the

required negative camber.

Good vertical wheel travel:

It generates good vertical wheel travel than any other suspension system.


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4.3 STIFFNESS CALCULATION AND SHOCK ABSORBER

SELECTION

The stiffness of the spring is an important parameter in determining the roll resistance

of the vehicle. It also plays a role in promoting oversteer.

4.3.1 MOTION RATIO

The motion ratio also has a large effect on the stiffness, seen by the vertical

displacement of the wheel. The motion ratio is defined as the ratio of shock travel to wheel

travel (shock travel/ wheel travel). The shock travel is a fixed value, so by adjusting the

motion ratio the wheel travel of the rear suspension is determined.

One compromise that was made, as a result of using the multi-link rear suspension, is

the implementation of the motion ratio. When using the multi-link design the shock must be

mounted to the upright. This means that the shock must be laid over to create the necessary

motion ratio. When the shock is mounted at a large angle the motion ratio will change with

wheel travel significantly. This creates a stiffening effect on the wheel rate, so as the

suspension compresses it will have an increasing stiffness due to the variation in motion ratio.

In off-road suspension this can be an advantage because it will help eliminate excessive

acceleration of the driver, due to hard bottom outs. The wheel rate also effects how the car

behaves when turning. The non-linear wheel rate is a concern for the cars roll stiffness but the

springs operate in roll in the linear range of the wheel rate curve.

4.3.2 STIFFNESS CALCULATION

Stiffness of the spring refers to the amount of force that has to be applied to deflect the

spring by one unit. Stiffness of the suspension system plays a major role in the vehicle ride


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behavior. The comfortness and the ride handling characteristics majorly depend on the

suspension stiffness. The suspension travel rate also depends on the spring stiffness.

The suspension stiffness is directly proportional to the frequency and the motion ratio.

The motion ratio refers to the ratio of the wheel travel to the spring travel. The desired motion

ratio can be achieved by the location of the mounting of the shock absorbers on the suspension

control arms with respect to the distance from the wheel axis.

Stiffness Frequency and Motion Ratio

Where f= Frequency

For an All Terrain Vehicle, Motion Ratio must be greater than one for optimal suspension

performance.

4.3.2.1 FRONT SUSPENSION STIFFNESS

Where = front suspension frequency = 1Hz

= 190kg

Required wheel travel at front= 11=279.4 mm

Required shocks travel at front = 5.8=147.32 mm

Substituting all the values in the stiffness equation, we get


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4.3.2.2 REAR SPRING STIFFNESS

Where = front suspension frequency = 1.35 Hz

= 190kg

Required wheel travel at front= 11=279.4 mm

Required shocks travel at front = 5.8=147.32 mm

Substituting all the values in the stiffness equation, we get

4.3.2.3 ROLL GRADIENTThe normalized roll stiffness number is the roll gradient, expressed in degrees of body

roll perg of lateral acceleration. A lower roll gradient produces less body roll per degree of

body roll, resulting in a stiffer vehicle in roll. Typical values are listed below:

0.20.7 deg/g for stiff higher downforce cars

1.01.8 deg/g for low downforce sedans

A stiffer roll gradient will produce a car that is faster responding in transient conditions, but at

the expense of mechanical grip over bumps in a corner. Once a roll gradient has been chosen,

the roll gradient of the springs should be calculated, the anti-roll bar stiffness is used to

increase the roll gradient to the chosen value. The roll gradient is usually not shared equally by

Front spring stiffness,Kf = 26193 N/m

Rear spring stiffness,Kr = 37973.30 N/m


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the front and rear. The roll gradient distribution is called as the Total Lateral Load Transfer

Distribution (LLTD). The Lateral Load Transfer is expressed as the percentage of the roll

gradient taken by the front suspension of the car. Roll gradients are degrees of body roll per g

of lateral acceleration. Roll rates are Newton-meters of torque per degree of body roll or ARB

twist.

FRONT ROLL RATE

Where = front track width

= Left front wheel rate

= Right front wheel rate

Figure 4.5 Track width and wheel base


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WHEEL RATE AT FRONT

Wheel rate =

Front suspension stiffness= 26193 N/m

Motion Ratio at front=1.89

Substituting the values in the wheel rate equation, we get

Front track width= 54 = 1.3716 m

KLF = KRF = 7332.66 N/m

Substituting the above values in front roll rate equation, we get

WHEEL RATE AT REAR

Rear suspension stiffness= 37973.30N/m

KLF = KRF = 7332.66 N/m

= 120.3 (Nm/deg roll)


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Motion Ratio at front=1.89

Substituting the values in the wheel rate equation, we get

Rear track width= 50 = 1.2716 m

KLR= KRR= 14071N/m

Substituting the above values in front roll rate equation, we get

4.3.2.4 ROLL GRADIENT CALCULATION

Where W= Vehicle weight (N)

H = Centre of gravity to roll axis distance

Substituting the values of

W = 330*9.81 =3237.3 N

H = 6.5* .0254 m =0.1651 m

= 120.3 Nm/deg roll

= 206.94 Nm / deg roll

KLR= KRR= 14071 N/m

= 206.91 (Nm/deg roll)


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Ddf (-ve sign indicates)

Since the roll gradient is within 1.0 to 1.8 deg/g for low downforce vehicles, we do not need

the Anti-Roll Bar to be installed.

4.3.2.4 ADVANTAGES OF NOT HAVING ARB

It reduces the unsprung mass of the vehicle. The suspension geometry becomes comparatively easier to design.

No need to provide complex provisions to house ARB in the chassis.

4.3.3 SHOCK ABSORBER AND SPRING SELECTION

Shock absorbers are used to absorb the vibrations, shocks that get generated from the

imperfections on the road and prevent it from getting transmitted to the vehicle chassis and the

vehicle occupants.

4.3.3.1 AIR-SHOCKS SELECTION

Air-Shocks acts both as the suspension spring and shock absorber. Air-Shocks were

chosen over other type of springs as it provides better suspension travel and characteristics

than any other systems. It also weighs so less that reduces the unsprung mass by large

margins.

4.3.3.2 ADVANTAGES OF AIR-SHOCKS

Reduces un sprung mass

For high performance of the vehicle the unsprung mass must be within 12-18% of

sprung mass


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Good suspension travel

Negative spring prevents the bucking effect.

Progressive air spring-prevents the roll of the vehicle.

Figure 4.6 Comparison of various springs

The figure 4.6 compares the stiffness of the air-spring with other coil-over springs. It

also shows how the stiffness of the progressive air-spring increases whereas it is almost alinear one for the coil-over shocks.

4.3.3.3 DAMPING RATIO

It is the ratio of the damping force of the spring to the damping force of the car.

= C/Ccr

Where C = Damping force (N)

= damping ratio

And Ccr =


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An undamped system will tend to eternally vibrate at its natural frequency. As the

damping ratio is increased from zero, the oscillation trails off as the system approaches a

steady state value. Eventually, critical damping is reached- the fastest response time without

overshoot. Beyond critical damping, the system is slow responding. But the amount of

damping does not change the steady state value- it only changes the amount of time to get

there and the overshoot. Figure 4.7 shows the frequency response of the spring-damper system

at different frequencies.

Figure 4.7 Spring damper system

4.3.3.4 RIDE AND SINGLE WHEEL BUMP DAMPING

Choosing a damping ratio is a tradeoff between response time and overshoot.

Passenger vehicles generally use a damping ratio of approximately 0.25 for maximizing ride

comfort. In racecars, 0.65 to 0.70 is a good baseline; this provides much better body control

than a passenger car (less overshoot), and faster response than critical damping. Some cars end

up running damping ratios in ride greater than 1 which shows that there is a compensation for


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a lack of damping in roll and pitch, as the dampers that control ride motion usually also

control the roll and pitch motion.

4.3.3.5 TRANSMISSIBILITY

The transmissibility (TR) is the ratio between output and input amplitude. The input

amplitude is the height of the speed bump, with output amplitude being vertical movement of

the body. The response of the suspension system (the car sitting on the suspension) is dictated

by the frequency and amplitude of the input.

Rearranging the equation above gives a method to calculate vertical body movement

from input disturbance amplitude and the transmissibility, which we can calculate from the

mass, spring rate, and damping ratio.

Output Amplitude = TR * Input amplitude

Figure 4.8 Transmissibility ratio


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Input is a displacement of the wheel caused by the bump. The input is the wheel

movement. The distance the mass of the car will move up and down is the output amplitude.

The time it takes for the wheel to complete the up-down cycle is the frequency divided by two.

As we increase the speed, we increase the frequency- and for sprung mass systems the

transmissibility changes with frequency.

The figure 4.8 helps one to choose the damping ratio, as it also gives the

transmissibility of the spring at a frequency upto 30 Hz.

4.3.3.6 FORCE VS VELOCITY CURVE

Force Vs velocity curve (figure 4.8 )is used as a medium to compare the force required in the

shocks to damp the vibration at a particular velocity and to select the type of Air-Shocks.

Initial Slope =

= Damping ratio in ride

= Ride frequency (Hz)

= Sprung mass supported by damper (kg)

Figure 4.9 Force Vs Velocity Curve

0

200

400

600

800

1000

1200

0 20 40 60

F

o

r

c

e

(

l

b

s)

Velocity (in/s)

Force Vs velocity Curve

Float R

Obtained Force VS

velocity curve


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4.3.3.7 FOX FLOAT-R AIR SHOCKS

FOX FLOAT-R AIR SHOCKS was selected, as the force Vs velocity curve of the

Float-R air shocks matches with Force Vs Velocity curve required for our ATV.

Figure 4.10 FOX Float R

The figure 4.9 depicts the external and the internal components of the Air-shocks

system. One need not confuse the presence of a spring in the shocks, it is the negative spring

that eliminates the effect of bucking.

FOX-FLOAT R- Air-Shox Feature-Rebound damping controls the rate at which the

shock returns after it has been compressed. The proper rebound setting is a personal

preference, and changes with rider weight, riding style and conditions. A rule of thumb is that

rebound should be as fast aspossible without kicking back and driving the bars into the riders

hands. The rebound knob (see figure 4.10) is located on the lower air sleeve body cap.

4.3.4 SUSPENSION GEOMETRY ITERATION IN LOTUS SUSPENSION

ANALYZER

Lotus Suspension Analyzer is used to iterate the suspension geometry. The software is

used to set the desired suspension kinematics such as camber, castor, toe in and toe out, scrub

radius, king pin axis inclination and it gives the overall view of the suspension(figure 4.11)


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Figure 4.11 Isometric view of suspension system

4.3.4.1 SETTING UP OF DESIRED SUSPENSION KINEMATICS

In the lotus suspension analyzer the co-ordinates are given for the various suspension

links. Then the co-ordinate points are moved iteratively to obtain the desired suspension

kinematics like camber gain, toe change, etc.

Figure 4.12 Bump Vs Camber gain

The figure 4.12 shows the front and the rear camber change with respect to the vehicle

bump. The front camber gain is more acute than the rear camber gain to generate more traction

force at the front than the rear.

-5

-4

-3

-2

-1

0

1

-100 0 100 200

Camber

Bump

Front Camber

Rear Camber


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Figure 4.13 Bump Vs Toe Change

The figure 4.13 shows that there is very little toe change in the front,which shows the

bump steer problem has been eliminated. Some degree of toe change is assigned to the rear for

easy cornering of the rear.

Figure 4.14 Bump Vs Roll angle

The figure 4.14 shows the roll angle with respect to the bump and it can be seen that

the roll-angle is minimum because of the higher suspension stiffness.

-2

-1.5

-1

-0.5

0

0.5

1

1.5

-100 0 100 200

Toechan

ge

Bump

Front toe

Rear Toe

-15

-10

-5

0

5

10

15

-100 0 100 200RollAng

le

Bump

Roll angle

Series2


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The figure 4.15 shows the behavior the suspension during the bump. It can be seen that

the castor angle is maintained around 4.5 degree to increase the straight-line ability. It can also

be seen that there is no bump steer.

Figure 4.15 Bump Vs Suspension characteristics

Figure 4.16 Rear suspension system

-6

-5

-4

-3

-2

-1

0

1

2

3

4

5

-100 0 100 200

Bump Travel

Bump Vs Suspension Characteristics

Front camber change

Rear camber change

Bump steer(toe

change)

castor angle


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The kinematic linkages of the multi link trailing arm rear suspension system is shown

in the figure 4.16 and the figure 4.17 shows the roll centre movement of the rear suspension

system

Figure 4.17 Roll centre Movement

4.3.5 3-D DESIGN OF THE SUSPENSION ARMS

The suspension trailing arm was designed in such a way that it becomes easy to

fabricate the suspension arms through welding. The 3-d design of the trailing arm is shown in

the figures 4.18 and 4.19

Figure 4.18 3-D Trailing arm design


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Figure 4.18 Top View of trailing arm design

4.3.6 MATERIAL SELECTION

Materials selection is another important point to cover. There is a huge range of

different properties that can be achieved with the use of engineering materials, from

composites to many different alloys. We focused in three major goals when we try to find the

best material that meet our needs, such goals are: the material deformation with several

impacts, the cost of the material, and manufacturability. In the second place but not less

important, we consider the weight of components. The material which best match these

requisites was the alloy steel AISI 1026.

Table 4.1 Chemical Properties of AISI 1026

ELEMENT WEIGHT %

Carbon 0.22-0.28

Manganese 0.60-0.90

Phosphorus 0.04 (max)Sulphur 0.05 (max)
http://www.efunda.com/materials/elements/element_info.cfm?Element_ID=Chttp://www.efunda.com/materials/elements/element_info.cfm?Element_ID=Phttp://www.efunda.com/materials/elements/element_info.cfm?Element_ID=Shttp://www.efunda.com/materials/elements/element_info.cfm?Element_ID=Shttp://www.efunda.com/materials/elements/element_info.cfm?Element_ID=Phttp://www.efunda.com/materials/elements/element_info.cfm?Element_ID=C


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4.3.6.1 PROPERTIES CHART

Properties

Density (1000 kg/m ) 7.858

Poissons Ratio 0.27-0.30

Elastic Modulus (GPa) 190-210

Tensile Strength (Mpa) 490

Yield Strength (Mpa) 415

Elongation (%) 15

Reduction in Area (%) 40

Hardness (HB) 143

Table 4.2 Properties Of AISI

MATERIALS MODULUS

OF

ELASTICIT

Y (GPA)

YIELD

STRENGT

H

(MPA)

TENSILE

STRENGT

H

(MPA)

PERCENT

ELONGAT

ION

DENSITY

(G/CME3)

Alloy ti-6AL-

4V

114 830 900 14 4.43

Tungsten

(commercial)

400 760 960 2 19.3

Steel alloyA36

207 220-250 400-500 23 7.85

AISI 4130

CHROMOLY

210 360.6 560.6 28.2 8

Stainless alloy

304

193 205 515 40 8

AISI 1026 210 415 490 40 7.85

Table 4.3 Comparison Of Materials

AISI 1026 has Average properties that match a wide range of operating scenarios,

unlike it fails in extreme loaded conditions, those extraordinary cases were calculated
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assuming top speed impacts, neglecting any driver maneuvers and without applying braking.

Since the elongation range is considerably wide, we have found out that AISI 1026 can be an

excellent option. A big range in the plastic region will give us a warning when the components

must be replaced, and it will also help us to avoid a sudden fracture. Also, the yield strength is

high enough to withstand several scenarios with considerable load range without being

deformed.

Table 4.4 Suspension Specifications

Front suspension Rear Suspension

Type Double SLA

wishbone

suspension

Trailing arm with

camber links

Shock Absorber Fox Float R

EVOL

Fox Float R EVOL

Shock travel 6 inch 6 inch

Shocks stiffness 26.10 N/mm 37.320 N/mm

Motion Ratio 1.83 1.83

Material SAE 1026 SAE 1026

Roll centre height 234mm 274mm

Lateral load

transfer

distribution

35% 65%

Roll rate

distribution

37% 63%

Camber -1.5degree -1degree

Castor 5 degree -

Bump Steer Nil Nil


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5. FABRICATION

The arms after welding with one other was annexed with the chassis and the knuckle

wheel through HEIM joints and fasteners.

5.1 HEIM JOINTS

A rod end bearing shown in figure 5.1, also known as a heim joint or rose joint, is a

mechanical articulating joint. Such joints are used on the ends of control rods, steering

links, tie rods, or anywhere a precision articulating joint is required. A ball swivel with an

opening through which a bolt or other attaching hardware may pass is pressed into a circular

casing with a threaded shaft attached. The threaded portion may be either male or female.

Figure 5.1 Heim Joints

5.1.1 ADVANTAGES OF HEIM JOINTS

Heim joint provides good articulation than other types of joints.

They offer easy mounting and adjustment of linkages and control rods.


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5.2 TRAILING ARM TO THE CHASSIS

The clamp was welded to the chassis of the vehicle.

A single M 14 heim joint was used to join the trailing arm with the chassis of the

vehicle.

A bush with internal threading of M 14 and 1.25 pitch was machined and was welded

to the extended portion of the trailing arm.

The shank of the heim joint was made to mate with the bush of the trailing arm. A

stainless steel M14 bolt was used to fasten the heim joint of the trailing arm with the

chassis.

5.3 TRAILING ARM AND THE CAMBER LINK

The rear extended portion of the trailing arm was connected to the camber link with a

M12 Heim joint.

At first a M12 nylon nut was TIG welded to the rear extended portion of the trailing

arm and to the end of camber links.

The threaded shank of the heim joint was fastened with the camber link and a M12 bolt

was used to fasten it with the welded nylon lock nut of the extended trailing arm.

A thread sealant was used to prevent the thread of heim joint threaded shank from

coming out of the lock nut.

5.4 CAMBER LINKS AND THE ENGINE BAY NOSE

A single M 12 heim joint was used to join the camber link with the engine bay nose

box.

A M12 nylon nut was TIG welded to the camber link end.

The threaded shank of the heim joint was fastened with the camber link and a M12 bolt

was used to fasten it with the engine bay nose box.

The four camber links, two on each side was formed with the same method


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5.5 KNUCKLES AND THE TRAILING ARM

The suspension arm is connected to the hub and then to the wheels through knuckles.

The knuckle was first decided to be made of aluminum casting, but the decision was

later reverted owing to the difficulties in joining it with the trailing arm.

The knuckles was then designed in such a way that it can be welded with the trailing

arm directly

The knuckle was designed with flanges on either side that has semi-circular contour as

shown in figure 5.2, to weld the knuckle with the trailing arm.

Figure 5.2 Rear Knuckle

5.6 MOUNTING OF THE SHOCKS

Mounting of the shocks is a very important process in the fabrication of the suspension

system. Shocks position on the control arms with respect to the distance from the wheel axis

determines the motion ratio. So having calculated the motion ratio at first, prompt care must

be taken for mounting of the shocks on the suspension arms.


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The shocks absorb all the forces that act on the wheel and thus bending forces are

induced on the arms on which the shocks are mounted. Therefore while mounting the shocks

on the arms; the distance of the shocks pivot from the wheel must be taken into account. If the

distance between the shocks pivot and the wheel increases, the bending moment that acts on

the arm increases and so the distance must be kept minimal without disturbing the desired

motion ratio.

The mounting of the shocks on the arm influences the following,

The shocks travel

The motion ratio of the suspension

The stiffness of the suspension

Bending moment that acts on the arms

5.6.1 SHOCKS MOUNTING

The shocks clamps (shown in figure 5.3) are welded to the trailing arm and to the

REAR ROLL HOOP of the vehicle.

Figure 5.3 Shocks Clamp

The wheels are connected to the hub and the knuckle of the trailing arm.


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The shocks are pressurized to 45 psi in the main chamber in order to provide the

required suspension stiffness and a pressure of 150psi is being maintained in the

EVOL chamber.

The wheels are skied off and kept in the full droop position

The shock is then mounted to the trailing arm clamp

The wheels are then lifted to the normal running position using a jack, so that the

upper pivot of the shocks can be clamped to the RRH of the vehicle.

Figure 5.4 Shocks Mounted

Figure 5.4 shows the vehicle being levitated to mount the shocks. It can be clearly seen

that there is no interference between the engine bay and the suspension system which again is

the major advantage of the multi-link trailing arm suspension system


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6. ANALYSIS

Total weight of the vehicle = 3237.3 N (330 kg)

Force on each wheel =

=

= 809.325 N

Trailing arm length = 748 mm

Trailing arm angle = 16.732degrees

Upper camber link length = 414

Lower camber link length = 410mm

Upper camber link angle =17.85

Lower camber link angle = 19.376

6.1 AT STATIC VEHICLE CONDITION

The angle of trailing arm with respect to horizontal axis = 16.732 deg

eight on each wheel * rear track width distance = Weight of Trailing Arm (cosine of length of

trailing arm)

809.325* 0.635 =WTrailing arm * (0.748 * cosine (90-16.732))

Force along horizontal direction, FX = Fcos73.268

FX = RX = 668.1650N

Force along vertical direction, FY =Fsin73.268

FY = RY = 2289.167 N

WTrailing arm = 2390.33N


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6.2 REAR IMPACT

Figure 6.1 Rear Impact

The figure 6.1 shows a vehicle at a speed of 16.67 m/s hitting our ATV which is at

rest, the force will be acting on the camber links. So the rear impact test is essentially a test for

the camber links.

From Newtons second law,

Force = mass * acceleration

FA=MA*aA, FB=MB* aB

By re-arranging the equations, we get FA=

MA=MB = 330 kg

VResultant= VA+VB (+ sign since the vehicle moves towards each other but VA=0)

VResultant= 0+16.67=16.67m/s

Time period of impact, t = 0.5 seconds.

Substituting all the values in the equation,

F=

(4 camber are there on a whole)

F=5501.1N (force acts on the camber links)

Force on each camber link= 5501.1/4 = 1375.275N


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Figure 6.2 Stress on camber links due to rear impact

Figure 6.3 Deformation on camber links due to rear impact

The figure 6.2 and the figure 6.3 show the stress and the deformation of the camber

links when the load acts at it because of the rear impact


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6.3 SIDE IMPACT

The figure 6.4 shows a vehicle at a speed of 16.67 m/s hitting the ATV on the side.

The force will be acting on the trailing arm. Most of the forces here will be withstood by the

tire itself and only the remaining will get transmitted to the trailing arm.

Figure 6.4 Rear Impact

Force on vehicle 1 and 2is given by

FromNewtons second law,

Force = mass * acceleration

FA=MA*aA, FB=MB* aB

By re-arranging the equations, we get FA=

MA=MB = 330 kg

VResultant= VA+VB (+ sign since the vehicle moves towards each other but VA=0)

VResultant= 0+16.67=16.67m/s

Time period of impact, t = 0.5 seconds.

Substituting all the values in the equation,

F =


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F=5501.1N

(2 links make a trailing arm)

Figure 6.4 Stress on trailing arms due to side impact

Figure 6.5 Deformation on trailing arms due to side impact

The figure 6.4 and the figure 6.5 show the stress and the deformation of the trailing

arm links when the load acts at it because of the side impact.

Force on each arm= 5501.1/2 = 2750.05N


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6.4 FALLING

Consider the vehicle falls from a height of 0.5m at a speed of 25 km/hr

ByNewtons second law,

Force(impact) = mass * acceleration

Fimpact =

Mass of the vehicle =330 kg

Initial velocity of the vehicle, Vi = 25*(5/18) = 6.94 m/s

Final velocity of the vehicle, Vf=

Vf=

Final velocity of the vehicle, Vf= 3.1m/s

Fimpact =

Shock absorber equation

From equation of motion ,

V=U+2as

U=0

V=3.13/s

Kinetic Energy =mv=1585.65N

F= KE/s=1585.65/0.5F = Ry = 3171.3N

Shock absorber reaction =8227.42-3171.3

=

Fimpact = 8227.42N

Shock absorber reaction =5056.12N


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Maximum moment:

M = 0.748*3171.3+5056.12*0.396 = 4374.35 Nm

Figure 6.6 Deformation on trailing arms due to fall

Figure 6.7 Stress on trailing arms due to fall

The figure 6.6 and the figure 6.7 show the deformation and stress of the trailing arm

links when the load acts at it because of the fall. The trailing arm to chassis point was fixed

and the loads are applied at the knuckle joining point at the trailing arms.


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6.5 BUMP

In this case we will assume that the vehicle cruises at 16.67m/s and sees a semi-

circular bump of radius 1m. Passing through this bump causes a resultant centripetal force,which then magnifies our normal force.

Centripetal force, Fc = Fnormal force- Self Weight

Fc = mv/r

Mass of the vehicle =330 kg

Velocity of vehicle = 6.94 m/s

Radius of bump = 1m

Fc = 330*0.6*6.94/1 = 9426.78 N

Fnormal force = Fc + W

= 9426.78 + 3237.8

Fnormal force = 12264 N

Resolving the components,

FX = RX =Fcos73.23

FX = RX =3585.67 N

FY = RY = Fsin73.23

FY = RY = 11700N

Shock absorber reaction

Kinetic Energy =mv=1585.65N

F= KE/s=1585.65/0.5

F = Ry = 3171.3N

Shock absorber reaction =11700-3171.3

=8528.8N


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57

M=6662*0.7105+23118*0.3552=12945Nm

Figure 6.8 Stress on trailing arms due to bump loads

Figure 6.9 Deformation on trailing arms due to bump loads

The figure 6.8 and the figure 6.8 show the stress and the deformation of the trailing

arm links when the load acts at it because of the bump. The trailing arm to chassis point was

fixed and the loads are applied at the knuckle joining point at the trailing arms.


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7. CONCLUSION

In this project, a multi link rear trailing arm suspension system is designed, analyzed

and fabricated in an All Terrain Vehicle, to increase the ride handling and the performance

characteristics of the vehicle.

There is no fixed procedure or set of universal constant values to hang around in the

design of the suspension system. Suspension system design involves meaningful assumptions

backed by literature survey, being taken at every stage. Its the system of meaningful

assumptions with trade-off between the various advantages that each suspension kinematic

parameters has to provide.

The Multi Link trailing arm was selected because it had several distinct advantages

like easy packaging and servicing of the engine bay components in an All Terrain Vehicle.

Also it has excellent load withstanding capacity than any other independent suspension

system. The project also involved the addition of camber links to the pure trailing arm system

as a move to get rid of zero camber adjustment drawbacks, since negative camber gain during

a bump and cornering is very much essential for an off road race car to maintain the required

contact patch between the tire and the road, thereby providing the enough traction force.

The multi-link trailing arm suspension system is not only easy to fabricate but also is

an excellent option for the Rear suspension system, especially for the rear-engine cars and All

Terrain Vehicles. In the rear-engine cars and ATVs, the use of other suspension systems at

the rear will make the servicing of the engine and the drive train components extremely

difficult and every time the shocks and the other linkages has to be dismantled to get access to

the engine bay. But with the advent of the multi-link trailing arm suspension system, no

linkages and the shocks has to be disturbed for the access to the engine bay, by this way the

servicing becomes lot easier. Apart from servicing and packaging ease, it provides highvertical wheel travel rate; the prime requirement of any suspension system. Thus the multi-link

trailing arm suspension system edges all the other suspension system types with its extreme

Documents

Final Report1 2