Design,Analysis & Fabrication of suspension of all terrain vehicle

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Table of content

SUSPENSION

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1. List of symbols and abbreviation

2. Important Terminologies

3. Overview of Suspension Systems

a. Introduction

b. Types of Suspension systems

i. Dependent Suspension Systems

ii. Independent Suspension System

4. Front Suspension for BAJA vehicle

1. Selection of suspension system

2. Suspension Geometry

3. Simulation of Suspension

5. Design of front suspension components

1. Wishbone Design

1.Geometry

2.Modeling

3.Analysis

2. Front Upright Design

1.Designing Parameters

2.Material Selection

3.Modeling

4.Loading Scenarios

5.Analysis

3. Front Wheel Hub design

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1.Introduction

2.Wheel bearing Selection

3.Designing Parameters & Considerations

4.Modeling

5.Loading Conditions

6.Analysis

4. Stub Axle design


2.Modeling

3.Analysis & Material Selection

6. Rear Suspension for BAJA vehicle

1. Selection of suspension system

2. Suspension Geometry

3. Simulation

7. Design of rear suspension components

1. Trailing Arm & Rear upright design

1.Geometry

2.Designing parameters & Modeling

3.Analysis

2. Rear Wheel Hub design

1.Introduction

2.Wheel bearing Selection


4.Modeling

5.Loading Scenarios

6.Analysis

8. Fabrication of Front Suspension Components

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1. Wishbones

2. Front Uprights

3. Front Wheel Hubs

4. Stub axles

5. Trailing arms

6. Rear uprights

7. Rear wheel hubs

STEERING

9. Introduction to Steering System

10. Types of Steering Gearboxes

11. Design of Steering System

1. Steering geometry

2. Collapsible Steering Assembly

12. Fabrication of Steering System

13. Summary

14. Bibliography

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1. Units Used

Steering Effort Fst N

Braking torque Tb N-m

Radius of brake disk rb M

Cornering Force Fcorner N

2. Important terminologies

1. Upright: the component that holds together the suspension control arms to the tyre

of the vehicle, the upright accommodates the upper and lower ball joints. It allows

for the pivotion of the tyre to perform steering action, the upright only pivots

about its axis, and moves in a vertical path tracing the tyre.

2. Hubs: The hubs are the rotating components that allow for the rotation of the tyre,

they are placed on the stub axle or live axle, usually supported by wheel bearings,

these hubs carry the rim along with the tyre and in most cases the brake rotors.

3. Stub axle: it is an integral part of the upright, it is an extension from the knuckle

that will eventually be used to carry the hub, the stub axle is a replacement for a

live-axle.

4. Mounting tabs: These are plates that are welded on to the frame, with the required

provision for bolting the suspension components such as wishbones and the shock

absorber. They are positioned according to the required suspension geometry.

5. Shock absorbers: The primary damping components used in a vehicle to allow the

dissipation of energy absorbed by the tires during motion, which may be bump

force or any other form of shock loading. Shocks control spring motion, that is,

they slow down and reduce the magnitude of the spring’s oscillation. The process

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is known as damping. In technical terms, a shock controls the frequency and

amplitude of the suspension's oscillation. In layman’s terms, a shock controls

how fast and how much the suspension compresses and rebounds.

6. Wheel rate: is the actual rate of a spring acting at the tire contact patch. This value

is measured in lbs/inch or N/mm.

7. Suspension frequency: refers to the number of oscillations or "cycles" of the

suspension over a fixed time period when a load is applied to the vehicle.

8. Ductility: In materials science, ductility is a solid material's ability to deform

under tensile stress; this is often characterized by the material's ability to be

stretched into a wire.

9. Finite Element Methods: The finite element method (FEM) is a numerical

technique for finding approximate solutions to boundary value problems for

differential equations. It uses vibrational methods to minimize an error function

and produce a stable solution. Analogous to the idea that connecting many tiny

straight lines can approximate a larger circle, FEM encompasses all the methods

for connecting many simple element equations over many small sub domains,

named finite elements, to approximate a more complex equation over a

larger domain.

10. Hardness: Hardness is a measure of how resistant solid matter is to various kinds

of permanent shape change when a force is applied. Hardness is dependent on

ductility, elastic stiffness, plasticity, strain, strength, toughness, and viscosity.

11. Hardness number: A number representing the relative hardness of a mineral,

metal, or other material as determined by any of more than 30 different hardness

tests such as Brinell hardness number, Rockwell hardness number.

12. Endurance Limit: The maximum stress that a material can withstand for an

infinitely large number of fatigue cycles; maximum cyclic stress level a metal can

withstand without fatigue failure. See also fatigue strength.

13. Bore: In terms of machinery bore is a process of enlarging a hole to a precise

diameter with a cutting tool within the hole by rotating either the tool or the work

piece.

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3. Overview of Suspension Systems

3. 1 Introduction

Vehicle dynamics is concerned with two aspects of the behavior of the machine. The first

is isolation and the second is control. It is a study of the behavior of the vehicle in a

dynamic state.

Chart reference: Multi-body systems approach to Vehicle dynamics by Mike Blundell

and Damian harty.

The most important constituents of vehicle dynamics are given below with a brief

explanation.

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Isolation: The process of separating the driver from disturbances occurring as a result of

vehicle operation. It is evident that there are two major types of disturbances, external

which are due to factors outside the vehicle system and mostly beyond the control of the

driver such as aerodynamics, which is weight gain, lift etc, road factors speak about the

road nature, roughness etc, internal factors constitute the vibrations due to engine and

transmission components etc. The behavior of the vehicle in response to road undulations

is referred to as ‘ride’ and could conceivably be grouped with refinement, as the ride of

the vehicle can be tuned to achieve a desired effect.

Control: response of the vehicle to driver demands. The driver continuously varies both

path curvature and speed, subject to the limits of the vehicle capabilities, in order to

follow an arbitrary course. Speed variation is governed by vehicle mass and tractive

power availability at all speeds on different types of terrain.

The suspension of an automobile is a system that consists of mechanical components that

are designed to fit in a preferred geometry, being able to handle the effects of road

irregularities or dynamic characteristics. The primal objectives of the system are

● To be able to locate all the four wheels of the vehicle

● To be able to maintain the required ride height (ground clearance)

● To provide a stable and comfortable ride and handling

● To provide road contact, even in inhospitable terrain.

This thesis is about the designing process and fabrication methods involved for an all-

terrain vehicle that has been designed and manufactured to compete in the BAJA

SAEINDIA 2014 competition.

Designing suspension system for any car requires technical knowledge in several

disciplines such as design of machine members, kinematics and geometry. The geometry

essentially means the board subject of how the unsprung mass of the vehicle is connected

to the sprung mass.

These connections or links dictate the path of motion and also control the forces that are

transmitted between them.

A suspension geometry must be designed to meet the requirements or ideals of the

vehicle to be built, a lot of factors must be taken into consideration such as the ride

height, the travel, the spring rates etc.

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Usually the chassis or the frame of the vehicle needs to be modified or designed

according to the suspension system, hence one setup cannot be used elsewhere. There is

no single best geometry.

Apart from the basic objectives a few vehicle specific suspension objectives have been

made to design the suspension for the all-terrain BAJA vehicle, they are:

● Large suspension travel

● Adjustability, to tune the suspension.

● Good ground clearance.

● Simple and lightweight construction.

Every suspension setup is an assemblage of control arms, shock absorbers, uprights, axles

and tyres which are laid out in a way to do their part, quietly in the background as the

vehicle is put to all types of loads.

The design of this system gets complex in way because, while being restricted (

controlled ) in their motion path by the control arms, the wheel will have camber, caster

and toe change. Therefore sometimes just links or control arms are not sufficient to

provide a good suspension characteristic. In such a scenario various components such as

toe link, camber link, shock absorber mounting become extremely important.

The suspension design has been done in a phase-wise manner to ease up the task and to

provide better flexibility of the results and also to allow for modification of the design if

there may be a need.

Phase 1: Determination of geometry to be used and the type of setup to be used

Considerations made during Phase 1 were:

● Independent nature.

● Comply with rule book track width of 64”.

● Smaller packaging.

● Fabrication limitations.

● Weight reduction.

Phase 2: Determination of spring and damper system.

Considerations made during Phase 2 were:

● Motion Ratio

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● Installation Ratio

● Spring rates

● Damping characteristics

● Weight of the Spring-Damper system

● Initial compression of spring

● Ride frequency

In this regard, various other parameters which are explained in later chapters have also

been taken care off, some of them being, position of roll centre, minimization of scrub

radius, anti-squat, anti-dive. To avoid rollover the vehicles centre of gravity has been put

as low as possible, by doing this we have restricted the movement of the centre of gravity

to an extent.

After establishing the design parameters the team has done different types of market

surveys locally and on the internet to find components that are well suited for the

purpose, the emphasis was on manufacturing most of the components to avoid

outsourcing, although expensive, it would serve all the requirements as well as have clean

engineering ethics rather than modifying an existing setup to suit ours.

Objective of steering system:

● Allow for sharp steering angles.

● Allow for extended suspension travel.

● Provide straight-line stability.

● Minimize bump steer.

● Have good steering return ability

● Be precise and compact.

3. 2 Types of Suspensions

3. 2.1 Dependent suspension system

This type of system normally has an axle which holds both the wheels parallel to each

other and perpendicular to the axle. When certain amount of changes occur in one wheel

due to some external causes such as bump then the same amount of changes occur in the

other wheel in the same manner.

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Dependent systems may be differentiated by the system of linkages used to locate them,

both longitudinally and transversely. Often both functions are combined in a set of

linkages

Types of dependent suspension system are:

1. Solid axle

1. Solid axle over leaf spring

2. Solid axle over coil spring

2. De-Dion

3. Torsion beam suspension

1. SOLID AXLE:

A solid axle also known as beam axle is a dependent type of suspension system in which

a single beam or a solid axle or a shaft is connected to a pair of wheels in a lateral

manner. With a beam axle the camber angle between the wheels is the same no matter

where it is in the travel of the suspension

1. Solid over leaf spring:

A Leaf Spring is a simple form of spring commonly used in heavy commercial vehicles

and four wheeled drive vehicles. The leaf spring is arc-shaped and works by suspending

the chassis of the vehicle to avoid contact with wheels. In this type of suspension system

the drive axle is clamped to the leaf spring and the shock absorber normally bolted

directly to the axle. The end of the leaf springs is attached directly to the chassis, as are

the tops of the shock absorbers. A leaf spring takes the form of a slender arc-shaped

length of spring steel of rectangular cross-section. The center of the arc provides location

for the axle, while tie holes are provided at either end for attaching to the vehicle body.

For very heavy vehicles a leaf spring can be made from several leaves stacked on top of

each other in several layers, often with progressively shorter leaves.

Leaf springs can serve locating and to some extent damping as well as springing

functions. While the interleaf friction provides a damping action, it is not well controlled

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and results in restriction in the motion of the suspension. It can either be attached directly

to the frame at both ends or attached directly at one end, usually the front, with the other

end attached through a shackle (a short swinging arm). The shackle takes up the tendency

of the leaf spring to elongate when compressed and thus makes for softer springiness.

There are several types of leaf springs available based on number of leafs stacked on each

other and the geometry of the spring few of which are mono-leaf spring, elliptical leaf

spring, semi-elliptical leaf spring and more. In the modern implementation of parabolic

leaf spring. The design is characterized by fewer leaves whose thickness varies from

center to ends following a parabolic curve. In this design, inter-leaf friction is unwanted,

and therefore there is only contact between the springs at the ends and at the center where

the axle is connected. The main advantage of parabolic springs is their greater flexibility,

which translates into vehicle ride quality. The basic advantage of this type of setup is that

the leaf spring acts as a linkage for holding the axle in position and thus separate linkage

are not necessary. It makes the construction of the suspension simple and strong. But as

the positioning of the axle is carried out by the leaf springs so it makes it disadvantageous

to use soft springs i.e. a spring with low spring constant. This type of suspension does not

provide good riding comfort. The inter-leaf friction between the leaf springs affects the

riding comfort. Acceleration and braking torque cause wind-up and vibration. Also wind-

up causes rear-end squat and nose-diving. The main drawback with this arrangement is

the lack of lateral location for the axle, meaning it has a lot of side-to-side slop in it.

2. Solid over coil spring:

This is a variation and update on the system described above. The basic idea is the same,

but the leaf springs have been removed in favor of either ‘coil-over-oil’ spring or shock

combos. A coil spring, also known as a helical spring, is a mechanical device, which is

typically used to store energy due to resilience and subsequently release it, to absorb

shock, or to maintain a force between contacting surfaces. They are made of an elastic

material formed into the shape of a helix which returns to its natural length when

unloaded. Because the leaf springs have been removed, the axle now needs to have lateral

support from a pair control arms. The front ends of these are attached to the chassis, the

rear ends to the axle. The variation shown here is more compact than the coil-over-oil

type, and it means we can have smaller or shorter springs. This in turn allows the system

to fit in a smaller area under the car. In addition to cylindrical springs, with which the line

of force moves along the damper axis, lateral forces compensating side load springs are

produced that have a line running diagonally to the spring center line. Use of side load

springs leads to increase in driving comfort, driving safety and optimum use of space due

to extensive compensation of lateral forces. One of the main advantage of coil spring is

that it gives a better road handling and better braking.

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The principal advantage of the beam axle is its simplicity. This simplicity makes it very

space-efficient and relatively cheap to manufacture. Beam axles are also ideal for

carrying heavy or varying loads because they do not ever exhibit any camber change as

the suspension travels. They are nearly universally used in heavy-duty trucks and most

light and medium duty pickup trucks, SUV’s, and vans also use a beam axle, at least in

the rear. The drawbacks are that it does not allow each wheel to move independently in

response to bumps, and the mass of the beam is part of the unsprung weight of the

vehicle, which can further reduce ride quality. Also the cornering ability is typically

worse than other suspension designs because the wheels have zero camber angle gain

during body roll.

2. DE-DION:

A de Dion tube is an automobile suspension technology. It is a sophisticated form of non-

independent suspension and is a considerable improvement. A de-Dion suspension

uses universal joints at both the wheel hubs and differential, and uses a solid tubular

beam to hold the opposite wheels in parallel. Unlike an anti-roll bar, a de Dion tube is not

directly connected to the chassis nor is it intended to flex. In suspension geometry it is

close to the trailing beam suspension seen on many front wheel drive cars, but without

the torsional flexibility of that suspension. With this system, the wheels are

interconnected by a de Dion Tube, which is essentially a laterally-telescoping part of the

suspension designed to allow the wheel track to vary during suspension movement. This

is necessary because the wheels are always kept parallel to each other, and thus

perpendicular to the road surface regardless of what the car body is doing. This setup

means that when the wheels rebound, there is also no camber change which is great for

traction, and that's the first advantage of a de Dion Tube. The second advantage is that it

contributes to reduced unsprung weight in the vehicle because the transfer case /

differential is attached to the chassis of the car rather than the suspension itself.

Naturally, the advantages are equaled by disadvantages, and in the case of de Dion

systems, the disadvantages would seem to win out. First off, it needs two CV joints per

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axle instead of only one. That adds complexity and weight. Well one of the advantages of

not having the differential as part of the suspension is a reduction in weight, so adding

more weight back into the system to compensate for the design is a definite disadvantage.

Second, the brakes are mounted inboard with the calipers attached to the transfer case,

which means to change a brake disc, you need to dismantle the entire suspension system

to get the driveshaft out. (Working on the brake calipers is no walk in the park either.)

Finally, de Dion units can be used with a leaf-spring or coil-spring arrangement. With

coil spring it needs extra lateral location links, such as a Panhard rod, wishbones or

trailing links. Again - more weight and complexity.

de Dion suspension was mostly used from the mid 60's to the late 70's and could be found

on some Rovers. More recently deDion suspension has had somewhat of a renaissance in

the specialist sports car and these all uniformly now use outboard brake setups for ease-

of-use, and a non-telescoping tube, usually with trailing links and an A-bar for lateral

location (rather than a Watts linkage or Panhard rod). Unlike most fully independent

suspension there are no camber changes on axle loading and unloading (or rebound).

Fixing the camber of both wheels at 0° assists in obtaining good traction from wide tires

and also tends to reduce wheel hop under high power operations compared to an

independent suspension. And with this setup the designer has free will to opt for various

shock absorbers. The most disadvantageous point is the manufacturing cost of this setup

which is high when compared to other types of setup.

3. TORSION BEAM SUSPENSION:

A torsion bar suspension, also known as a torsion spring suspension or torsion beam

suspension, is a general term for any vehicle suspension that uses a torsion bar as its main

weight bearing spring. One end of a long metal bar is attached firmly to the vehicle

chassis; the opposite end terminates in a lever, the torsion key, and mounted

perpendicular to the bar that is attached to a suspension arm, a spindle, or the axle.

Vertical motion of the wheel causes the bar to twist around its axis and is resisted by the

bar's torsion resistance. The effective spring rate of the bar is determined by its length,

cross section, shape, material, and manufacturing process. The ride height may be

adjusted by turning the adjuster bolts on the stock torsion key, rotating the stock key too

far can bend the adjusting bolt and more importantly place the shock piston outside its

standard travel. Over-rotating the torsion bars can also cause the suspension to hit the

bump-stop prematurely, causing a harsh ride. Aftermarket forged-metal torsion key kits

use relocked adjuster keys to prevent over-rotation, and shock brackets to keep the piston

travel in the stock range.

One of the main advantages of this setup is that its durable and it allows the user to adjust

the ride height with an ease. And the amount of space occupied by this setup s less when

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compared to coil spring suspension. And one of the main disadvantage of this type of

setup is that it cannot a progressive spring rate like a coil spring suspension setup.

3. 2.2 Independent Suspension System

Unlike dependent system an independent suspension system allows wheels to rise and fall

on their own without affecting the opposite wheel

There are several types of geometry present in this type of suspension system among

which few are:

1. Double wishbone

2. Macpherson strut

3. Semi-Trailing arm

1. DOUBLE WISH BONE:

This type of system is also known as A-arm or wishbone type of suspension system. This

type of suspension system is a design using two wishbone shaped arms placed parallel to

each other in a vertical manner and are locate to the wheel. Each wishbone or arm has

two mounting points to one at the chassis and the other at the joint of the knuckle on top

and bottom. The shock absorbers are mounted to the wishbones to control vertical

movement. Double wishbone designs allows carefully to control the motion of the wheel

throughout suspension travel, controlling parameters such as camber, caster angle

, degree of toe, roll center height, scrub radius, scuff and more.

The upper arm is usually shorter to induce negative camber, and often this arrangement is

titled an "SLA" or short long arm suspension. When the vehicle is in a turn, body roll

results in positive camber gain on the lightly loaded inside wheel, while the heavily

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loaded outer wheel gains negative camber. Each arm is connected to chassis from one

side and to the knuckle on the other.

Double wishbone suspension setup allows the engineer to mount shock absorbers at ease.

It is easy to work out the effect of moving each joint, which helps in easy tuning of

kinematics of the suspension and wheel motion can be optimized. It is also easy to work

out the loads that different parts will be subjected to which allows more optimized

lightweight parts to be designed. They also provide the engineer for easy tuning of

camber gain, unlike the Macpherson strut, which provides negative camber gain only at

the beginning and then reverses into positive camber gain at high jounce amounts.

As disadvantage, it may take more space and is slightly more complex than other systems

like a Macpherson strut. Due to the increased number of components within the

suspension set up it takes much longer to service and is heavier than an equivalent

Macpherson design.

Double wishbones are usually considered to have superior dynamic characteristics as

well as load-handling capabilities, and are still found on higher performance vehicles.

Examples of makes in which double wishbones can be found include Alfa romeo,

Honda and Mercedes-benz., double wishbone suspension, is very common on front

suspensions for medium-to-large cars and is very common on sports cars and racing cars.

It also provides least camber change at bump and rebound condition.

MACPHERSON STRUT:

This type of system is currently the most widely used front suspension system in most of

the cars. Macpherson struts consist of a wishbone or a substantial compression link

stabilized by a secondary link which provides a bottom mounting point for the hub

or axle of the wheel. The system basically comprises of a strut-type spring and shock

absorber combo, which pivots on a ball joint on the single, lower arm. The strut itself is

the load-bearing member in this assembly, with the spring and shock absorber acting as a

member holding the car up. The Macpherson strut required the introduction of unibody

or monocoque construction, because it needs a substantial vertical space and a strong top

mount, which monocoque body can provide, while benefiting them by distributing

stresses

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The biggest advantage of this type of setup is the low manufacturing cost and the

simplicity of design, it is not generally considered to give as good handling and the

quality of ride as double wishbone geometry. Despite these drawbacks, the Macpherson

strut setup is still used on high performance cars.

SEMI-TRAILING ARM:

A semi-trailing arm suspension is an independent rear suspension system

for automobiles where each wheel hub is located by a large arm that pivots at two points.

This type of suspension employs two trailing arms which are pivoted to the car body at

the arm's front edge. The arm is relatively large compare with other suspensions' control

arms because it is in single piece and the upper surface supports the shock absorber. It is

rigidly fixed to the wheel at the other end. Viewed from the top, the line formed by the

two pivots is somewhere between parallel and perpendicular to the car’s longitudinal

axis. It is generally parallel to the ground. Trailing-arm and multilink suspension designs

are much more commonly used for the rear wheels of a vehicle where they can allow for

a flatter floor and more cargo room. Many small, front-wheel drive vehicles feature

a Macpherson strut front suspension and trailing-arm rear axle.

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Advantage of this type of setup us that it has anti-lift effects while braking. A

disadvantage of this suspension setup is that during roll in a turn, the rear wheels attain a

positive camber angle equal to the roll angle of the car body, reducing slightly the rear

axle cornering capacity. Another disadvantage is that it complies to toe-out under lateral

loads. This tends to promote over-steer. To reduce this tendency, trail arm bushings need

to be highly stiff and rigid, which increases ride harshness.

4. Front Suspension for BAJA Vehicle

4. 1 Selection of type of suspension With the premise of introduction of suspension in chapter 2, the design of front

suspension for the BAJA vehicle is described in this chapter.

After reviewing various types of suspension geometries the “Double ‘A’ arm” setup was

preferred as it met many of our criteria which were:

● Adjustability.

● Ability to package in a small space.

● Simpler construction.

● Lightweight yet robust construction.

Main components of the suspension system (front):

● Wishbones or “A” arms

● Uprights or Knuckles

● Hubs

● Stub Axles

● Mounting tabs

● Shock absorbers

A standard double “A” arm or double wishbone geometry consists of two links that are

used to connect the chassis to the upright. The two links namely, upper wishbone and

lower wishbone each of which is provided with two revolute joints at the chassis end and

one rotational joint at the upright.

The design of suspension earlier was done using paper doll-models connected with

threads to verify the motion, but in a more sophisticated way the design has been done

using various computer software’s that provide better accuracy and analysis.

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4. 2 Suspension Geometry

The primary design constraint for designing the front suspension was the track width

limitation which was set to 64” by the BAJA SAE official rulebook, after adding some

human error and tolerances in installation we have decided on an optimum track width of

60” which would be the maximum achievable track width for us basing on our

manufacturing standards without the risk of exceeding the governed track width.

Maximum track width was chosen in order to achieve the highest possible wishbone

lengths which directly affect the travel or the motion in a vertical path of the suspension,

having a large suspension travel was one of the objectives of the design.

Elaborated in later chapters the double wishbone design also offers us a higher stiffness

and load bearing capacity per weight than any other type of suspension geometry.

Geometric planning: The setup started with a set of unknowns and a set of desired values,

important unknown parameters for design of components were:

● Length of wishbones

● Height of uprights

● Kingpin Inclination

● Hub offset

● Spindle offset

● Angles of wishbones with axis at desired ride height

● Wheel offset

These parameters were required to design the suspension components for stress analysis

as well as for dynamic simulation.

Desired suspension characteristics were:

● Low scrub radius

● Ride height

● Height of roll center

● Height of mounting point of steering rack w.r.t lower wishbone.

A brief description is given with regards to support the above mentioned requirement.

● Low scrub radius: to avoid excessive scrubbing and premature wear of

tyres, double wishbone suspensions induce a scrub radius due to its basic design

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from in which its steer axis does not coincide with its centerline axis, unlike

Macpherson strut setup which are inherently zero scrub geometries. The aim is to

minimize scrub as much as possible. The scrub radius if affected by the Kingpin

inclination, hub offset and wheel offset.

● Ride height: a ride height of 250mm was decided to be adequate to clear

most of the obstacles to pass underbody without being obstructed, it also is a good

tradeoff between roll, as greater ride heights, greatly influence the position of

center of gravity with respect to the ground.

● Height of roll center: the roll center position greatly affects the way the

vehicle behaves and also the way in which the forces are transmitted to the chassis

through the suspension. Technically speaking the roll centre must be as low as

possible to the ground as it will facilitate the transfer of weight into body roll

instead of it being transferred to the tyres on their contact patch, therefore to

maintain a low roll center the vehicle has to run a very high spring rate and a stiff

setup, which is not very ideal in the case of an off-roader where the surface of

vehicle operation is not smooth. A roll centre that is very close the chassis will

cause excessive jacking forces and put a lot of stress on the suspension mounting

points, therefore one needs to run a very low spring-rate but the ride is defined by

being “Wobbly” with a lot of movement of the vehicle due to body roll. Therefore

the best position chosen for the roll center for the vehicle was somewhere in

between the road and the chassis biased towards the chassis in the span of ride

height.

● Height of mounting point of steering rack : the position is extremely

important as later discussed ‘bump steer’ is greatly influenced by this factor. The

position is also important to package the steering rack on to the chassis.

4. 3 Simulation Modelling of link geometry is followed by checking for link-clash or component clash

scenarios, such as Full droop, Maximum Bump, Side roll

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Fig. Modelled Link geometry of the vehicle showing various desired as well as calculated

values.

Full Droop scenario, no link clash detected.

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Max.bump scenario, no link clash detected.

Side body roll, no link clash detected.

From the above analysis it is noted that the design satisfies the requirements and does not

have any aberrant effects, the link geometry is also found to be good.

Wishbones or “A” arms: they are the control arms that allow for the travel of the

suspension in a controlled manner. Between the outboard end of the arms is a knuckle

with a spindle and hub. The wishbones are provided with bushing or rod ends, to

constrain their motion and allow motion only in desired path. The bushes also provide a

noiseless operation of the system.

Selection of Shocks:

The shock absorbers used in the 2014 BAJA vehicle are air shocks manufactured from

FOX racing corporation, California, USA. They have been acquired as part of

POWERSPORTS SAE/MINI BAJA program, the shocks provide a lightweight solution

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to the heavy coil-overs add to that the ability to run variable spring rates, ability to tune as

per driver specific requirements equal or surpass our requirement.

Specification of the shocks:

Manufacturer brand and model FOX Racing, Float 2.

Weight 0.8Kg

Length 18.8” ( Extended )

Width 6.2”

Travel 6”

Performance characteristics of the shocks:

Motion ratio : The ratio of the variables d1 and d2 and the angle between the shock

mounting with the vertical as shown in the figure below

image courtesy : Eibach motorsport.

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This ratio is based on the leverage theory of levers, mathematically represented by:

MR = (d1/d2) ∗ ACF

Where, ACF=cosA 9 (angle correction factor)

MR = (261/522) ∗ cos45

MR = 0.35

The value of the motion ratio has been kept close to 0.35, which creates a smoother ride,

while maintaining the required stiffness.

The following table shows the calculated values of frequency and wheel rate.

25

Fig. Suspension frequency for various vehicles, reference: Optimum Kinematics.

The suspension frequency is 1.57Hz which is moderate and as per allowed limits.

Simulation : The calculated values from the earlier designs has been modelled in a 3d

vehicle dynamics software “LOTUS Engineering, Suspension Analyzer - SHARK

module”, this software allows us to accurately model the suspension system and run tests

for scenarios such as 2d bump, 2d steer,3d bump, 3d steer, 3d roll

The figure below shows the front suspension model in LOTUS SHARK.

26

Fid. 3 dimensional bump, note there is no excessive camber change or toe change.

Fig: 3 dimensional steer.

Fig: 3 dimensional roll.

Graph 1:Camber (deg) vs Roll angle ( deg)

27

the vehicle gains 9 degrees of

negative camber at full roll

Graph above shows the camber change (deg) with respect to steer travel.

The following simulations have been done to ensure the suspension behaves in a planned

manner to determine the ride and handling, a lot of testing is required on various terrains,

tuning the suspension becomes of paramount importance to achieve the best required

control and handling of the vehicle.

28

5. Design of Front Suspension Components

5. 1 Design of Control arms

5. 1.2 Modeling

The following dimensions for the modeling of upper and lower control arms were

provided by suspension analyzer. The length of them were based on front nose

dimensions, track width and various other performance significant parameters. The model

ought to reflect the lengths mentioned above.

29

Various shapes for control arms were considered initially. Since the control arm is the

link between tire and body of the vehicle, it need to be stiff and strong to support also

control the tire motion.

There were a lot of variant in design of control arms few are:

30

After studying various design, the finalized design was make it out of a single tube by

bending into a parabola.

31

Reasons for selection of the design

1. Easy to fabricate

2. Consumes less time in production

3. Sophisticated jig not required

4. Less number of welds hence low heat effected areas.

5. Easy correction in design by opening and closing of bend.

The upper wishbone according to geometry was prepared as following

32

The upper stock connected the upright. It will be inserted by stock steering rod ends of a

commercial vehicle.

The model of lower wishbone is different from the mentioned as it needs to mount the

shock absorber. It is done by providing a lateral tube with the mount. Since the design is

for front suspension lower one mounts the shock absorber where as in case of it being

used in rear suspension mounts are given on upper wishbone due to obvious reasons

33

The tubing for control arm was chosen to be AISI 4130 steel with outer diameter 25mm

and 3.5 mm wall thickness. The tabs for shock mount were from 4mm thick mild steel

sheet.

5. 1.3 Analysis

5. 2 Design of Upright

5. 2.1 Design parameters

Before designing of any components there are various parameter that are to be included

in it. Irrespective of other details the main design parameters determine mostly the

performance, adaptability with the environment, mates with the sub-component in an

assembly, space occupancy etc. They are Special consideration and often are the

constrains which are to be met.

The parameters that molded the design of the upright were:

1. Include castor angle of 6o along the vertical axes of upright.

2. Project the brake caliper mounts at one side of upright.

3. Provide sufficient thickness to brake caliper mounts to endure sudden torque from

the disk rotors.

4. Check alignment of the brake caliper mount on both of the uprights i.e. Left and

Right uprights. Since the brake caliper doesn’t have plane of symmetry along its

center, brake mounts will have different spacial arrangements along the side of

both uprights.

5. Steering arm mount be on the opposite side that of the brake caliper mount of the

respective uprights.

6. Twin steering arm mount be provided to facilitate double shear for steering arm

bolts.

7. Dual bolt holes will be provided to counteract the moment in the steering arm.

8. Have sufficient fillet radius throughout the design to minimize notch sensitivity.

9. Length of upright will be taken as per the suspension design that gave the

optimum results.

10. Upper and lower wishbone mounting will be dependent on the kingpin

inclination obtained from the suspension geometry

34

11. Bore be provided to accommodate the stub axle.

12. Press fitting tolerance to be provided in the central bore diameter to press fit the

stub axle.

13. Sufficient wall thickness to make the component rigid, unsusceptible to external

moment.

14. Design optimization will be done after the component is analyzed for various

loading cases to relief weight.

5. 2 .2 Material Selection

As it was mentioned above in design parameters, weight consideration was the main

objective that dictated the choice of material for the suspension components. Also as

mentioned earlier the preference of low unsprung mass in the vehicular system, it was

necessary to opt the material which could bear the forces induced during the motion as

well be light.

Market survey revealed the local availability of the following suitable material candidate

for the component.

1. Cast Ductile Iron

2. Titanium Alloy Ti 6Al-4V

3. Aluminum Alloy AA 6351 T-6

4. Alloy Steel AISI 4140

Specific Strength comparison

It is the strength to weight ratio of the material. Strength here can be tensile or yield

strength. A higher ratio dictates the material has appreciable strength compared to its

weight.

35

The above graphs shows the usage of Titanium alloy to be most apt choice for the

component. Nevertheless other factors also have be considered before material selection

is to be made.

Cost per unit yield strength

Comparison also has to be made with the cost of the material. The cost comparison (C) is

made by following equation;

𝐶 = 𝑐𝑚 ∗𝜌

𝜎

Whereas,

Cm ꞊ Cost of the material per unit mass (Rs/Kg)

𝜌 ꞊ Density of the material (Kg / m3)

σ ꞊ Yield strength of the material (N/m2)

77674.6

186816.4

101851

58598.7

0

20000

40000

60000

80000

100000

120000

140000

160000

180000

200000

Ductile Iron Ti6Al4V AA 6351 T6 AISI 4140

SPEC

IFIC

STR

ENG

HT

N M

/ K

G

36

Ductility

Defined as the percentage elongation under stress before the material ruptures. This

property is important as it gives a clear warning before the metal breaks down.

The above graph shows the cost value for unit strength of the material. It can be seen that

Titanium alloy have the highest value for its strength whereas ductile alloy have the least.

Interpretations from above three detrimental property graphs;

1. Titanium and aluminum seems to be viable candidates where the strength to

weigh ratios are to be considered.

0

2000

4000

6000

8000

10000

12000

14000

16000

18000


SPEC

IFIC

CO

ST R

S K

G /

M N

0

5

10

15

20

25

30


ELO

NG

AT

ION

%

37

2. Ductile cast iron is the cheapest material which is to be considered, also fact that

most of the commercial automobile are fit with ductile cast iron uprights.

3. Titanium alloys have the highest specific cost among all

Considering all the above, it has been decided AA 6351 T-6 seems to be the most viable

material for the component. The strength on weight ratio is sufficient to meet our

standards. Choosing aluminum alloy also seems to be a most economical selection.

Since Aluminum alloy have been chosen, fatigue characteristics of haven to be taken into

account.

Components subjected to fluctuating loading fail at much lower loads than their service

loads due to fatigue. When not considered in the design stage it can lead to catastrophic

failures in their service life. Factors causing a fatigue loading may be many which are:

1. Large fluctuation of Loads i.e. Stress amplitude of high magnitude

2. Sufficient large number of cycles of stress applied

Additional factors which may exaggerate fatigue failure are:

1. Corrosion

2. Residual stress

3. Stress concentration

4. Temperature

5. Surface finish

6. Stress range

7. Use of welding

The mechanism of fatigue failure can be simple put into two stages i.e. crack initiation

and crack propagation to the point of static failure. This crack once formed begins to

grow in each cycle of loading. Growth is also accelerated by higher amplitude of loading

often characterized by multiple cracks initiated. Final catastrophic failure occur when a

crack has grown to a significant length such that the next application of load results in

static failure due to reduced area in that region. Fatigue cracks may start at various places

such as at concentration of plastic strains, extrusion or intrusions of the surface, grain

boundaries, internal voids and surface scratches.

Design methodology for any aluminum component used in car was to:

1) Load be highly approximated to actual condition which is being acted upon the

component

38

2) Optimize the component so that the life of the component be under 105 cycles

3) Check for maximum stress under the stated endurance range instead of checking

for yield criteria.

5. 2.3 Modeling

The uprights geometry determines the dynamic characteristics of the vehicle. Irrespective

of the model, geometry is kept same in all the models. The models prepared were first

based on manufacture feasibility, economic consideration and analytical sustainability.

Various models were initially prepared for the same.

The following are the various models prepared and the reasons for disapproval of the

designs.

39

Design-1 : Saruman

The first of the designs was initially desired to be made up of Steel plates. The plates were given

the thickness of 10mm. The plates were countered cut with the required measurements and

profile. Joining process was chiefly welding i.e. TIG (Tungsten Inert Gas) welding.

Adaptability of the Design :

o Cheap material cost o Easy Manufacturing

o Less event completion time

o Provision of steering stopper

Disapproval of Design

o Weight of the component- 2.5 Kgs.

Since high thickness of plates were being used the weight of the component was considerably

40

high. Also use of 10mm thick sheet was indispensable as any reduction of thickness was

detrimental.

o Low fatigue strength

Since the plate were joined by welding process the fatigue strength of the component was

considerably effected. This is because uncontrolled cooling and heating rates produced

around the weld zone, also called as heat effected zones. This microstructural variation

produces areas of high hardness whereby reducing strength.

o Steering upright geometry not accurate

Since manual fabrication processes were being employed, the dimensional accuracies were

uncertain.

Design-2 : Gollum

Mild Steel pipe of outer diameter 65mm and thickness of 2.5 mm was found to be adequate to

press fit wheel bearing. The profiled Mild steel rods following the geometry was welded to the

mother rim.

Adaptability of design

41

o Simple design

o Easy manufacturing

o Easy procurement of the materials required

Disapproval of design

o Intricate couping requirement.

o Welding reduces the fatigue strength and improper hardness in the sample.

o Strength not uniform due to alternate heating and cooling rates.

o Brake mounting arms are weak.

Brake mounting arms due to relative length from the parent pipe, became long and prone to

failure

o Weight of the component

As mild steel tubes were used the weight of the component was measured up to 1.2 Kgs.

o Poor Strength

Finite Element analysis found it to be weak at weld zones. Hence prone to failure

42

Design-3 : Gandalf

Stock Aluminum alloy 6351 T-6 would have been used. Manufacturing process being CNC

milling.


o Geometry of the upright was captured better o CNC milling process was fast and accurate

o High torsional rigidity.


o Mounting points for upper and lower wishbone was too small whereby increasing stress in those areas.

o Increasing the upper and lower wishbone mount areas increased the weight of the

component considerably.

43

o Weight of the component

It weighs 2.5 Kgs comparable to its steel counterpart.

Design-4 : Bilbo

It was then decided to provide stub-axle on the upright itself. The aluminum Upright was then

modeled to be CNC milled.


o Fast and accurate manufacturing

o Holes and slots provided to relief weight. It weighs only 760 gm. The least weight until now.

o Geometry captured accurately and effectively

o Bearing size is considerably reduced, whereby reducing weight.

44

o Bearing lock could be provided on built in stub axle


o Many relief holes increased notch sensitivity of the component o Built in stub axle cannot have wheel bolt as threading on aluminum shears off easily.

o Required hub offset cannot be changed.

o In case of failure of stub axle whole upright have to be changed

o Longer stub axle created immense bending strength on upright o Fillets and notches are requirement on the stub axle to be able to hold other components

well. These fillets and notch sensitivity added up stress on bending stress, creating more

stress it can hold o Built in stub axle was failing in large loading conditions.

o Longer steering arm also made it more susceptible to shear from the upright

o Also built in steering arm dictates changing the upright when steering arm fails, the

failure which is very common in the testing.

Rejection of all the other designs was valid as even though they all met with the

requirement of designing parameters but failed at various other important suppositions.

New design was required to have the modeling advantages of all other designs at the

same time to rectify all the drawbacks of the previous designs.

The new and improved design was modeled in the following steps.

Step 1:

The basic mold was created around the upright geometry. This captures almost all

the features which were important to suspension design.

45

Step 2 :

The next step was to differential the upper and lower wishbone mounting

positions. An arbitrary thickness was given which will be verified once the model

is added to testing module. Care was also taken care to be able to provide

sufficient room for accommodation of castle nuts.

Step 3 :

The brake mount position was carefully positioned so as when brake caliper in

engagement would reach the end of the brake disk, for it to hold it firmly when

under braking effect. The longer brake mount was given a parabolic profile to

have a uniform strength also to reduce material.

After calculated width of material was left for the center hole on the upright which

would hold the stub axle.

The extended length of the steering arm was reduced to decrease the added

moment at the end of the upright.

Step 4:

After bail calculation considering the material, diameter and step dimensions for

the stub axle. the central bore for the same was provided on upright

46

Step 5:

It was seen as an added strength to steering arm by placing its bolt under double

shear. Also to counteract the effect of couple number of holes for its bolts were

increased from 1 to 2.

Step 6:

The castor was given 5 degrees, according to which the position of upper and

lower mounting points were determined. The hole diameter was kept equal to the

rod-end diameter.

47

The model was thus prepared with rectifying all the drawbacks obtained from the

previous failures of various models, also every model was an improvement over the other

until the final design (Hobbit) was considered to be selected.

The various dimension were kept variable in view of further analysis. Finite element

analysis will dictates all the dimension of the component.

5. 2.4 Loading scenarios

Contemplation of actual force distribution on a single suspension component during

motion of car can be quite complicated. To ease this complication and various

48

complicated vector forces experienced by the component, it is usually split down into

various individual scenarios. These scenarios take up the maximum force derived from

calculation and constrains the model appropriately to represent the actual event.

The following scenarios of loading will be considered for the analysis

1. Steering effort

2. Brake mounting force

3. Remote loading

4. Cornering Force

It is to be noted that weight of the component was considered during analysis.

1. Steering Effort Calculation,

As steering effort was calculated in section-12, Steering effort (Fst) is taken as 1.5g

𝐹𝑠𝑡 = 4400

The tie rod being designed to take purely axial load, whereas such is not the case with

steering arm. As the wheel travels the direction of forces changes. To accommodate this

effect the direction of steering force wasn’t taken parallel to ground but at an angle 45o

with the horizontal.

Then components along X and Y comes out to be equal i.e.

𝐹𝑠𝑡 𝑥 = 𝐹𝑠𝑡 𝑦 = 3100

2. Braking Mount force calculation,

49

The following free body diagram shows the position of brake caliper mounting pints with

respect to center of the disk. This is because the brake mounting position is not

symmetrical with the center of the caliper. And hence the force experienced by the each

mount will be different.

Braking Mount Force (Fbm)

𝐹𝑏𝑚 =𝑇𝑏

𝑟𝑏

By sum of forces and applying moment at a point we get,

Force on larger brake mount ꞊ 3200 N

Force on smaller brake mount ꞊ 2300 N

3. Remote loading

Due to rim and hub offset, the bumps force is not a direct loading but an eccentric

loading. The values of the mentioned dimension have already been calculated.

Fshock have already been calculated (6. c.vi) ꞊ 22,000 N

4. Cornering force

As per calculation the cornering force was taken to be 1.5g

𝐹𝑐𝑜𝑟𝑛𝑒𝑟 = 5000

50

5. 2.5 Analysis

An upright is the crucial component as every force experienced by the car is passed on

through it. Also analyzing every effect of this complex ever changing moment and forces

becomes complicated. A worst case scenario’s is therefore recommended where forces

are scaled and restrains are applied in view of real time. The team performed various

finite element computational analysis to match approximately with the actual conditions

that will be experienced by the component so as to avoid failure in real-time.

There are many different types of analysis that must be completed to ensure that the part

in question is able to withstand the applied loads. In addition, there are other factors that

must be included in each analysis to ensure that the analysis itself is correct. In order to

analyze the part correctly, the restraints must be an accurate representation of the real

world scenario and the loads must be calculated for different loading scenarios.

Finally, the mesh must be as homogenous as possible. This would include minimizing the

difference in aspect ratio between elements, as well as maximize element mapping

quality. We must ensure that all of the meshes we use for the different components in our

assembly are set up to be compatible with one another.

The main objective behind analysis was to check the maximum stress induced, predict the

life of component and establish a suitable factor of safety in design.

Since most of suspension component used were aluminum it must be noted forth that it

doesn’t exhibit a fixed fatigue limit unlike major steel categories.

All of the analysis was done in student package of Solid works 2013 Simulation module.

The meshing package utilizes the tetrahedral mesh on the component. To give better

accuracy the mesh size was made finer until the results became stagnant. Any more

decrease in mesh size would just waste computer resource without marginally increasing

any solution accuracy.

51

Modal analysis was also done on the component to determine how the system behaves in

its displacement dynamic response. It was done to check how different frequencies

naturally excite the system to a degree where resonance fluctuates through the component

resulting in dropping of life expectancy of the system. For this study no external load is

applied to the component, while it is known that external loads do not affect the natural

frequency. The component was however restrained.

In all the analysis self-weight of the component was considered.

Scenario Bump loading

Loading Stub-axle bore at an

offset of 50mm

Constrains Upper and lower

wishbone mounting

Force 22,000 N

Maximum Stress 109 MPa

Maximum Deflection 0.053 mm

Factor of safety 2.42

Scenario Brake caliper mount

loading

Loading Caliper holes

Constrains Upper and lower

wishbone mounting

Force 3200, 2300 N

respectively



52


Scenario Cornering of vehicle

Loading Upper and lower

wishbone mounting

Constrains Stub-axle bore

Force 5000 N




Scenario Steering of vehicle

Loading Steering arm holes

Constrains

Upper and lower

wishbone mounting

points

Force 3000 N




The above plots of the results for various cases that might be experienced the component

during its operation life cycle

53

Being the suspension component of an automobile, it is under constant non uniform

loading condition. Also the loading condition vary drastically from terrain. In order to

simplify the loading scenarios fully reversed loading i.e. R (Stress Ratio) ꞊ -1 was

preferred over other stress ratios ranging from -∞ to ∞. The tabular values of Sa (Stress

amplitude) with the corresponding life cycles [1] was fed into Solid works Fatigue

Analyzer. The corresponding S-N diagram was obtained from the data.

Solid works doesn’t have built in S-N graphs for all the materials that are present in its

directory, hence it was needed to plug them manually into the software.

Following is the obtained fatigue data that has to be fed into the system to perform

fatigue analysis.

Table 1: Fatigue data distribution for Aluminum 6061 T6

No of Cycles Stress Amplitude

(Zero Mean Stress)

N/m2

10.000000 482000000.000000

70.000000 482000000.000000

100.000000 420000000.000000

200.000000 325000000.000000

500.000000 241000000.000000

1000.000000 198000000.000000

2000.000000 168000000.000000

5000.000000 142000000.000000

7000.000000 135000000.000000

10000.000000 120000000.000000

54

20000.000000 99000000.000000

50000.000000 80000000.000000

100000.000000 71000000.000000

200000.000000 64000000.000000

500000.000000 58000000.000000

1000000.000000 55000000.000000

2000000.000000 53000000.000000

5000000.000000 51000000.000000

10000000.000000 50000000.000000

20000000.000000 49880000.000000

50000000.000000 49280000.000000

100000000.000000 48990000.000000

200000000.000000 48700000.000000

500000000.000000 48500000.000000

1000000000.000000 48400000.000000

55

Figure 1: S-N curve for zero mean stress of Aluminum 6061 T6

The fatigue analysis could be done for individual scenarios however for a combined force

situation it seems logical to optimize. For this bump force, heavy braking, steering pull

were applied to component.

56

It can be seen the life of the component is well defined for more than 95% area having

1E9 cycles. Some of the areas shows a reduced life but that can be ignored potentially.

All attempts were made to restrict the maximum von misses stress under 100 MPa, after

studying the fatigue properties of aluminum. Also not compromising on weight gave the

best combination of strength, weight and fatigue characteristics for the component.

5. 3 Design of Front Hub

5. 3.1 Introduction

A wheel hub is a mounting position for wheel of the vehicle, it houses the wheel bearing

as well as supports the lugs and brake disk. It can either transmit power or be just rolling.

Its function is basically to keep the wheel spinning freely on the bearing while keeping it

attached to the vehicle. Designing a hub is very crucial as it alones is the interface

between the wheels and rest of the vehicle.

Lug bolts are usually integral to hub, hence these are also called as locking bolts.

Spacers are also used to fit between the hub and the brake disks. This is done to

accommodate different brake calipers as to avoid the scraping between them and the

calipers.

Usually in hubs in commercial vehicle are made up of alloy steels or cast iron. But for a

mini Baja vehicle it is advantageous to look at alternative material to make it light

weight.

57

5. 3.2 Bearing selection

Selection of suitable bearing for a particular purpose is immensely important in view of

the load it is meant to take at given rotational speed giving a certain life.

Max gear ratio of the transmission,

7.6

Max rpm of the engine crank,

3800 𝑟𝑝𝑚

The maximum rotational speed attained by the tire is,

500 𝑟𝑝𝑚

Assuming the life of the bearing to be designed for is,

1000 ℎ𝑜𝑢𝑟𝑠

Loading ratio (From data book)

𝐶

𝑃 = 3.11

Axial Force is assumed to be during cornering, which is taken as 1.5g (Pr)

2500 𝑁

58

Radial Force is the drop weight of the car (Pa)

4000 𝑁

the ratio of axial to radial,

𝑃𝑎

𝑃𝑟 = 0.45

for the corresponding ratio, the value of equivalent load

𝑃 = 𝑆(𝑋𝑃𝑟 + 𝑌𝑃𝑎)

𝑃 = 4000 𝑁

The dynamic load carrying capacity was found out to be,

12995 𝑁

For the given life and load rating the bearing number SKR 6007 was chosen for the front

wheels.

The given are dimensions of the bearing

5. 3.3 Designing Parameters & Considerations

Following are the Parameters which guided the design of Front Hubs,

59

1. Pitch Circle Diameter of Lug Bolts on rim

Since stock Maruti rims were chosen to be on the vehicle on all four wheels. In order that

the hub to sit on rim, the pitch circle diameter of the rim had to match with the designed

hubs.

Pitch Circle Diameter of Lug bolts on hub ꞊ 144 mm

2. No of Lug bolts and their size

Since the rim had 4 equi-spaced lug bolts holes. With the hole size of 12.5 mm diameter.

The holes to match with rim had to be provided on hub.

4 Lug Bolt holes with diameter ꞊ 12 mm ø

3. Bearing Provision

Since as mentioned earlier the bearing selected was SKF 6007. The bore on the hub with

sufficient tolerance is to be provided for the bearing to sit inside the hub.

The bearing bore on hub ꞊ 62 −0.03 0

4. Bearing Seat

For able to lock the bearing in the hub on one end it should have bearing seat. The

thickness of seat must be enough to withstand axial force while cornering.

5. Common Holes for Rim and Brake disk mounting

Common holes for both rim and brake will minimize the number of holes from 8 to 4 on

the hub.

6. Bearing Lock provision

On other end of the hub, the bearing will be locked by an internal circlip. A groove of

2mm is to be provided to facilitate the circlip.

5. 3.4 Modeling

A literature survey was undertaken before modelling of the component began. Initially as

a reference the stock Maruti hub was taken. The model was loaded in the Solid-works.

Initially it was decided to use it but due to its weight idea was soon dropped. Also

component was analyzed to develop a lighter and stronger equivalent in aluminum alloy.

60

The details in the model were removed to simplify the analysis. The brake mounting

holes were also removed to see the effect of analysis. Since the goal of the analysis will

be to target potent areas of weight reduction and geometry changes to suit the Baja

vehicle. Such method reduces the chances of failure of design as the adopted design is

commercially utilized. Also as mentioned in design parameter stub axle will be

eliminated in the front hubs to accommodate hub bearing which in turn will hold the axle.

An arbitrary force of any value have been loaded on the stub axle of the hub in context.

The constrains were the lug holes. The plot shows the stress distribution close to stub axle

extending towards lug holes. It was also evident that rest of the area experienced a

relatively less induced stress.

A plot of design insight reveals the observation to a scaled level.

61

The study of stock hub gave us the insights to new design of the hubs that would be

precisely follow its design guidelines.

Following is the timeline of the model of the component.

The blank is initially made and stepped to

host the wheel bearing. Provision is also

given to lock the bearing in the other

direction by a bearing seat.

The Pitch circle diameter equivalent to rim is

kept on the hub.

62

A single chuck of material is removed to

observe the effect.

The pattern is then repeat around to obtain

optimized wheel hub.

The grove for the circlip is provided to lock

the bearing in opposite direction.

Hence modelling is complete.

5. 3.5 Loading scenarios

Following are the various loading that were applied to hubs for analysis. This of these

forces so obtained will help the prediction of actual results.

1. Drop test

63

This test try to replicate stress produced when the vehicle falls from a certain height. To

obtain the impact load on the hubs,

Assuming the vehicle falls from the height (h),

ℎ = 1 𝑚

The impact velocity (v) of the vehicle is,

From kinematic relation for rectilinear motion

𝑣 = √(2 ∗ 𝑔 ∗ ℎ) − 𝑢2

whereas,

g ꞊ acceleration due to gravity (9.8 m/s2 )

u ꞊ Initial velocity of the vehicle before the fall ( 0 m/s)

Substituting the above value,

𝑣 = 4.42 𝑚/𝑠

And now for calculating the impact force (F)

From Work-Energy Principle,

Change in Kinetic Energy of the object ꞊ Work done on the object

1

2 𝑚 ( 𝑣2 − 𝑢2) = 𝐹 ∗ 𝑑

whereas,

d ꞊ the compression of the shock springs ( 0.15 m)

Thus,

𝐹𝑠ℎ𝑜𝑐𝑘 ꞊ 22,800 𝑁

2. Braking Torque Test

This test analysis the effect of panic braking on wheel hub. Since brake disk and hub are

directly connected, while under braking, brake disks induces opposite torque on the hub

to halt the vehicle.

Assuming the vehicle comes to a complete halt from 55 Km/h in a distance within 6m on

an off-road track.

Initial velocity (u) ꞊ 15.2 m/s

64

Braking distance (d) ꞊ 6m

Deceleration (ad),

𝑣2 = 𝑢2 + 2 ∗ 𝑎𝑑 ∗ 𝑑

whereas,

v ꞊ final velocity

thus,

𝑎𝑑꞊ 11.7 𝑚/𝑠2

Assuming weight distribution is 50:50, Force on Front wheels (Ff),

𝐹𝑓 ꞊ 2047.5 𝑁

Braking torque on front wheels (Tf) is,

𝑇𝑓 ꞊ 𝐹𝑓 ∗ 𝑅𝑡

whereas,

Rt ꞊ Radius of Tire (0.3048 m)

Thus,

𝑇𝑓 ꞊ 624 𝑁𝑚

3. Cornering & Skidding

Slip angle changes at the turn of the vehicle, sometimes amateur driver may fail

understand its significance and it results in vehicle skidding in turns.

It has been taken as 1.5g force for skidding whereas 1g for cornering force. As speed of

the vehicle is restricted the assumed values holds good.

Cornering Force ꞊ 3500 N

Skidding Force ꞊ 5200 N

4. Rim fracture

In the event of a rim failure, i.e shear of rim across the lug bolts or failure from flange, it

produces an eccentric loading at the hub due to its rim offset.

To take this into account the self-weight of the vehicle will be taken in consideration.

65

5. 3.6 Analysis

Scenario Drop Test

Loading Central bore

Constrains 4 Lug holes

Force 11,000 N




Scenario Braking Test

Loading Central bore

Constrains 4 lug holes

Torque 700 Nm



Factor of safety

Scenario Cornering & skidding

Loading Skidding-Bearing seat

66

Cornering-Central bore


Force Skidding-5200 N

Cornering-3500 N


Maximum Deflection 0.049


Scenario Rim fracture

Loading 3 spokes

Constrains Central bore

Force 4000 N




67

To optimize the component even further it was then added into Solid works Optimization.

The variable that was kept in the study was the thickness of the blank of the hub.

Following results were found,

Component

name Units Current Initial Optimal Scenario1 Scenario2

thickness mm 8 8 4 4 8

Stress1 N/mm^2

(MPa) 85.493 85.493 172.74 172.74 85.493

Mass1 Kg 0.151107 0.151107 0.115572 0.115572 0.151107

It can be seen that reducing the thickness from 8 mm to 4 mm would increase the max

stress in the component to 172.74 MPa which is fairly under the limit. The weight of the

component could also be reduced by 25%. But the optimization wasn’t carried out in the

fabrication due to reduced fatigue effects.

Compo

nent

name

Units Curren

t Initial

Scenari

o1

Scenari

o2

Scenari

o3

Scenari

o5

Scenari

o6

bearing

seat

thickne

ss

mm 4.5 4.5 2.25 4.5 2.25 2.25 4.5

seat

thick mm 5 5 2.5 2.5 5 7.5 7.5

68

Compo

nent

name

Units Curren

t Initial

Scenari

o1

Scenari

o2

Scenari

o3

Scenari

o5

Scenari

o6

Stress N/mm^

2 (MPa) 131.28 131.28 150.3 128.02 155.78 160.57 135.51

Mass kg 0.15110

7

0.15110

7

0.11901

1

0.13734

5

0.12954

1 0.14007 0.16487

The optimization was however carried to determine the optimum thickness of bearing

seat and thickness center bearing tube. Based on above optimization it was seen to reduce

the weight by 10%, the thickness of the bearing seat thickness as reduced to 2.5 mm from

5mm.

And hence after optimization the results were sent for fatigue optimizations. To check the

life of the component, a combined study was prepared to sum up all the forces in

different scenarios. This being done gives us the combined fatigue analysis also saves the

time for each individual iterations.

Optimizes component

Weight ꞊ 132 grams

Seat thickness reduced by ꞊ 50%

69

Since it was assumed the vehicle to be running 1000 hours at constant speed of 60km/h.

The predicted cycles each component has to run is calculated to be 2.8 E8 cycles.

The above fatigue plot shows various portion of the component actually close to required

value of cycle it was prescribed for. Hence the fatigue analysis were in good agreement

with the fatigue design parameters.

5. 4 Design of Front Stub Axle

5. 4.1 Designing Parameters & Considerations

Since the primary objective of stub axle was to host hub and upright, it needed to be

extremely rigid to sustain heavy loading. Since it is a dead axle in case of front

suspension, it experience the bending moment to a large extend.

Following are the design parameters that are to be met in modelling for the component to

be effective in function,

1. It need to have steps to ensure the bearing locks on one side.

2. Threaded tail to accommodate wheel nut.

3. One sided flange to fit into upright’s one end, hence locking its relative translator

motion in a direction.

4. Respective step should shoulder the selected hub bearing.

5. Provide hole for cotter pin on the threaded end.

Consideration

70

As other components of the system have been modeled, stub axle design is considerably

just adjusting the length between upright and hub. Hub offset is again dictated by the

suspension geometry. Also it must be noted that in suspension software’s position and

space constrain of brake disk is notably ignored. The whole assembly must have to be

verified for dynamic interference before the final hub offset can be decided upon. As

such the hub offset was changed to 80mm from previously 50mm to accommodate

braking component and avoid their interference with rest of the suspension components.

The nut for locking of bearing have to selected in such a way that it hold the inner

stationary ring of the same. If the dimension aren’t met then use of appropriate washers

have be put in place to satisfy the condition.

The position of hole for cotter pin have to be in such that the nut is properly locking the

bearing. Affirmation only has to be made when the components are physically assembled.

5. 4.2 Modeling

Since the modeling dimension were pre-defined by the upright, hub bearing and hub

offset. Also in considerations with the detail design parameters the modelled was

prepared initially to satisfy all those.

When compared with other components of the suspension the stressed volume of the

component was just 39 %. Stating that the low stressed region to be carrying 60% of

weight. This high mass has to be compromised. Any steps taken to reduce the weight

have to be followed by Finite elements analysis.

71

Potent steps to reduce the mass of the component would be to bore the component. Also

threaded portion can be reduced on diameter.

5. 4.3 Analysis & Material Selection

The model was then loaded to FEM solver COSMOL in solid works simulations. Since

the selection of material is indeterminate to stresses induced in the component, an

arbitrary material value was specified to observe the stress levels.

To enact the actual conditions the following constrains were in place.

Constrains Contact point of upright

Loading Force-1

Longitudinal 5200 N

Tire skid force

Force-2

Lateral 11000 N

Impact force

Efficient meshing of the component is very critical to obtain accurate results. Initially by

using default meshing conditions it was found that most of the elements aspect ratio were

more than 1. Effect of more than 1 aspect ratio includes the stiffness matrix to be unstable

whereby reducing effectiveness of the mesh.

72

As obtained above the aspect ratio for most of the elements have been close to 1. This is

done by reducing the global size of the element. Also areas of low thickness were given

mesh control to reduce further the element size by using curvature based mesh.

Analysis was run and the maximum stress obtained was 310 MPa. Appropriate material

was then selected. It was decided to select AISI 4340 was selected with yield strength

700 MPa.

Design optimization had to be done to reduce the obtained mass of the component which

when came out to be 700gms.

A central bore was made initially of 12mm throughout the component. It was put under

Design study to obtain at optimum diameter thereby limiting stress and decreasing the

weight of the component.

73

After several iterations the software’s shows the serious of results.

Design Study Setup

Design Variables

Name Type Values Units

Diameter of

bore

Range with Step Min:12 Max:20 Step:1 mm

Constraints

Sensor name Condition Bounds Units

Stress1 is less than Max:600 N/mm^2

(MPa)

74

Goals

Name Goal

Properties Weight

(

grams)

Mass of component Minimize Mass 700

11 of 11 scenarios ran successfully.

Compon

ent name Units Current Initial Optimal

Scenario

1

Scenario

6

Scenario

2

Diameter

of bore mm 12 12 20 12 17 13

Stress

N/mm

^2

(MPa)

311.2

8

311.2

8

310.5

6

311.2

8 300.5

337.7

1

Mass 603.6

6

603.6

6

425.1

37

603.6

6

502.5

44

586.2

26

Scenari

o3

Scenari

o4

Scenari

o8

Scenari

o9

Scenari

o8

Scenari

o9

Scenari

o5

14 15 19 20 19 20 16

296.1 303.2

5 330

310.5

6 330

310.5

6

303.9

2

567.3

98

547.1

74

452.3

34

425.1

37

452.3

34

425.1

37

525.5

56

Hence the weight of the component was decreased by 40 % for the same strength. And hence the

75

component was prepared.

The above process is different from other design processes as due to many design constrains and

parameters, most of the design was already judges, leaving very less flexibility to designer, hence this

approach of design was selected to arrive at final model.

6. Rear Suspension for BAJA Vehicle

6. 2 Suspension geometry

The rear suspension of the BAJA vehicle is different from the front suspension in the

following ways:

● Must be designed to bear greater loads due to rear placement of engine

● Must not allow great camber changes

76

● Allow the live-axle to be fixed without link-clash

● Must not allow toe changes

● Must not allow axle-plunge out.

● Must be independent.

With the above premise, the types of suspensions were reviewed again, following which

the decision came down to the selection between double wishbone suspension and a

trailing arm design. The double wishbone suspension although very adjustable and

lightweight would not be a match here as the rear track width limited to 58” which meant

that the packaging space will not be enough, the absence of chassis support members to

mount the arms led us to the design of trailing arm.

Trailing arm suspension provides all the required camber and toe control.

Components used in the rear suspension:

● Trailing arm

● Trailing arm mount

● Rear Upright

● Rear hub

6. 3 Simulation

The shock absorber calculations are done in a way that is similar to the front suspension,

they only differ in way of running higher spring rates to accommodate in order to

maintain the sprung mass.

77

Fig: Modeled rear suspension also shown are drive shafts.

The rear suspension has been spooled to a zero droop rear design, to maintain traction of

the rear wheels at all times as they are the wheels that provide power to the vehicle. The

rear suspension also has limited travel in order to prevent axle plunge out, the travel is

restricted to 6” which was determined after intensive testing.

Fig.3 dimensional bump.

Fig. 3 dimensional body roll.

78

Fig. Camber change in bump and rebound.

Fig. Spring travel in bump and rebound.

79

Fig. Full car model in LOTUS SHARK.

Fig. 3d bump analysis of the full car model.

80

Fig. 3d steer analysis of full car model.

Studying the positions of the front and rear roll centers, shows the vehicle will be stable

as there is no excessive displacement of the said roll center.

Therefore after the simulation process, the design of the combination suspension of front

and rear is deemed fit for being used in the BAJA vehicle, the body roll and steer

characteristics have shown that the suspension will be capable of doing the required job

even on the inhospitable terrain. Provided with the zero droop rear design, the vehicle

will not encounter any mechanical problems such as axle plunge-out in case of high

speed cornering or large potholes.

7. Design of Rear Suspension Components

7. 1 Design of Trailing arm & Rear Upright

7. 1.1 Designing Parameters & Modeling

The trailing arm which chiefly locates the rear wheels is important in design point of

view as any changes reflect the power transmission ability and many other suspension

characteristics. Therefore its design requires careful trial and error. As in our case it’s not

just a tube but others connected to it.

Some of the consideration that are to be bore before modeling phase begins are:

81

1. Length of trailing arm

It be such that it projects the c.v shafts in perpendicular to travel of the vehicle. As any

deviation from it cause power loss

2. Should accommodate the brake caliper

3. Should have provision in end of it to pass through c.v joint. Also it should be such

that even at any articulation it shouldn’t be interference to trailing arm.

4. At appropriate motion ratio, it should have capacity to hold the air shock firmly

also allowing relative rotation

5. Should not deflect either ways when cornering. Ideally should not allow any

chamber change

6. Should not deform in heavy loading. The diameter and tube be considerably

adjusted to loading.

7. A single link should be able to locate the wheels without need of any other links

like camber control link etc.

82

7. 1.2 Modelling & Analysis

With all the consideration in mind the following trailing arm design was modelled.

Cup of the trailing arm was created

keeping in mind the clearance for cv joint

Cut away for articulation of cv shaft

Arm was then made to suit the length.

83

Holes for upright and plate for brakes was

then made accordingly.

Shock mounting was prepared giving

optimum motion ratio

Metal bush was then attached

Thus was completed.

Before the modelling manual calculation was need to arrive at the design values for the

tube that will be taken henceforth the design continued.

It was also decided to use alloy steel for the whole setup.

84

Analysis was don’t basically on two scenarios as they will be subjected to bending and

side loading while cornering. Both for which results are plotted below.

FIGURE 2 SHOWING MAXIMUM BENDING STRESS

85

FIGURE 3 SIDE SKIDDING ANALYSIS

7. 2 Design of Rear Hub

7. 2.1 Introduction

Rear wheel hub unlike front hub has an extra responsibility of transfer of torque. The stub

axle is live and splined. The chief component in power train system for transmission of

speed at a range of angle called as Constant Velocity (C.V) joint is directly engaged into

rear hub. This requires the splines to be formed into hub that matches that of the

respective C.V joint.

Many vehicles uses hubs made of Cast iron, alloy steels. They are either made by casting

or are forged to the required shape. Latte being producing strong components. Cast iron is

usually preferred because of its cheap cost and worthy compressive strength.

Many of the setup which uses the disk brake as their breaking setup has spacers that gaps

brake disk and hub, this enable the space to host brake caliper in between. The setup is

such that it makes servicing of brake disk easy and without removing whole of the

assembly.

7. 2.3 Design parameter & Considerations

Following are the Parameters which guided the design of Front Hubs,

86

1. Pitch Circle Diameter of Lug Bolts on rim

Since stock Maruti rims were chosen to be on the vehicle on all four wheels. In order that

the hub to sit on rim, the pitch circle diameter of the rim had to match with the designed

hubs.

Pitch Circle Diameter of Lug bolts on hub ꞊ 144 mm

2. No of Lug bolts and their size

Since the rim had 4 equi-spaced lug bolts holes. With the hole size of 12.5 mm diameter.

The holes to match with rim had to be provided on hub.

4 Lug Bolt holes with diameter ꞊ 12 mm ø

3. Bearing Provision

Since as mentioned earlier the bearing selected was SKF 6009. The outer diameters of

stub-axle on the hub with sufficient tolerance is to be provided for the bearing to sit on

the stub

The shaft diameter on hub ꞊ 45 0 0.03

4. Common Holes for Rim and Brake disk mounting

Common holes for both rim and brake will minimize the number of holes from 8 to 4 on

the hub.

5. Bearing Lock provision

The groves on the shaft should be provided to shoulder the two external circlips. This

facilitates the locking of bearing on the hub.

6. The internal diameter of the stub-axle had be such that it accommodates the C.V

joint

7. The length of stub-axle be more or equal to length of splines on C.V joint.

7. 2.4 Modeling

As the study carried out in Chapter 6. C.i, the profile of hub was taken to be same, which

relieves the material unaffected by the force.

The modelling was carried in following steps

87

The disk of the hub was modeled. The shaft is

part of the model.

As torque is being transferred the sudden

change in diameter acts as stress risers. To

avoid this fillet is provided of considerable

radius such that transition from lower to high

diameter is smooth

The in effected area is profiled.

The holes for lug bolts is then provided

depending upon the PCD. The central bore in

stub axle is provided which house C.V joint.

88

7. 2.5 Loading Condition

The loading conditions applied in here that differs from front is that torque is being

transferred through it.

Following are various loading conditions are were subjected to rear hub under certain

constrains to observe and analysis the functionality of the component.

1. Sudden Torque transfer

Since the maximum torque is obtained in 1st gear.

1𝑠𝑡 𝑔𝑒𝑎𝑟 𝑟𝑎𝑡𝑖𝑜 ꞊ 31.418

𝑀𝑎𝑥𝑖𝑚𝑢𝑚 𝑒𝑛𝑔𝑖𝑛𝑒 𝑡𝑜𝑟𝑞𝑢𝑒 = 19.67 𝑁𝑚

Sudden torque transfer, (Kt ꞊ loading factor ꞊ 1.6),

𝑇𝑟 = 1.6 ∗ 31.418 ∗ 19.67

𝑇𝑟 = 975 𝑁𝑚

2. Drop Test

Refer 5. c. v

Impact force on single hub ꞊ 11,000 N

3. Remote Loading

Self-weight of the vehicle ꞊ 3000 N

This force will be applied on two opposite side spokes of hub. At a distance away from

the center of hub by 200mm.i.e the rim offset of the wheel.

4. Braking reverse torque

Refer 5. c.v

Torque being applied is 600 Nm

5. Single Bending

A force of 2000 N is being applied only one a single spoke of the hub.

7. 2.6 Analysis

89

After the loading conditions were thought about and verified the model was proceeded

towards the analysis phase to check for strength and fatigue strength.

Scenario Torsion Test

Loading Bore of Stub axle


Force 975 Nm




Scenario Drop Test

Loading

Half Bearing

Contact on stub

axle


Force 11,000 N

90


Maximum

Deflection 0.081 mm


Scenario Remote loading

Loading Two lug holes of

opposite side

Constrains 2 opposite lug holes

Force 3000 N




Scenario Single Bending

Loading One of the spoke of

hub

Constrains Rest of the three lug

holes

Force 2000 N

91




Scenario Fatigue Test

Loading

Torque 500

Nm

Static force

3000N

Constrains Lug holes

Life of the component 1 e 9 cycles

Loading ratio

R ꞊ -1

Fully reversed

loading

8. Fabrication of Suspension Components

8. 1 Control arms/Wishbone:

Steps involved in the fabrication:

1. Manufacturing of Bullets

2. Manufacturing of wishbone

92

MANUFACTURING OF BULLET:

The purpose of manufacturing of bullet is that the size of the spherical rod end is 16mm

and the internal diameter of the chromoly tube is 18mm. so the bullets are manufactured

to act as a medium of attachment between the chromoly tube and the spherical rod end.

Materials required:

1. MS solid bar

2. Spherical bearing

PROCESS PLANNING:

Step 1:

Cutting: In this process MS solid rod of outer diameter 20mm and length of 15mm is cut

from the parent metal tube.

Step 2:

Step turning: In this process the ms solid rod goes through the process of turning with the

help of lathe machine where the one side of the rod is step turned to the length of 18.4mm

and with the outer diameter of 18mm.

Step 3:

Boring: In the process of boring the other side of the bullet is bored with the depth and

diameter of bore decided as per the length and diameter of the spherical rod end, so that it

can be exactly matched into it.

Step 4:

Tapping: In the process of tapping (i.e producing internal threading) the bored side of the

bullet is provided with internal threading with the help of tapping tool for the spherical

rod end to screw into the bullet.

The image below shows bullet after full assembly.

93

MANUFACTURING OF WISHBONE:

Material required:

1. Hollow chromoly tubes

PROCESS PLANNING:

Step 1:

Cutting: In this process two hollow chromoly tubes of length 496mm and 592mm is cut

from the parent metal tube for the future manufacturing process.

Step 2:

Bending: In this process the hollow chromoly tubes which are cut from the parent metal

tube is sent for the process of bending where a hydraulic bending machine bends the

hollow tube into a u-shaped tube also known as wishbone.

Step 3:

Welding: In the process of welding, first the bullets are assembled onto the u-shaped tube

and welded to it with the help of tungsten inert gas welding machine which completes the

manufacturing process of a wishbone.

94

The above shown figure shows the U-shaped tube after the process of bending.

8. 2 Front Uprights

Material Required

AA 6351 T-6 Aluminum alloy

Drawing of the component

96

Process planning

1. Procurement of stock

The block of aluminum was cut into 100 x 220 x 50mm

2. CAM

Writing computer cnc programming. Two cycles were made firstly roughing cut and later

smooth cut. NX cam was used as CAM software.

97

3. Positioning on machine

Various tolerance were checked which are flatness, parallelness etc.

4. CNC

98

After the codes were verified they were dumped into vertical CNC milling machine.

5. Quality control

After the component was manufactured it was verified with the drawing of the

same.

6. Final Component

99

8. 3 Front Wheel Hubs

Material Required



101

Process planning


The cicular block of aluminum was cut into 150 x 30 mm

2. CAM



102



4. CNC


103

5. Quality control

After the component was manufactured it was verified with the drawing of the same.

6. Final Component

104

8. 4 Stub axle

Materials Required

AISI 4340 rod


106

Process Planning

1. Stock was obtained 55 x 90 mm

2. Lathe setting

The center of the stock and lathe was positioned. Also the rotation of the component

was stabilized.

3. Lathe operation

The step was obtained as per the dimensions

4. Thread cutting

After the Tread per inch of the M27 were measured with pitch guage. treat cutting with

the configuration on lathe was done

5. Quality control


6. Final Component

8. 5 Trailing arms:

Materials required:

1. MS hollow rod of outer diameter 35mm and inner diameter 25mm(Rhino rod)

2. MS plate

3. High tensile rated bolts and nuts of size M12

4. MS hollow pipe of outer diameter 110mm with thickness 4mm(bangle pipe)

5. Hollw ms tube of OD 20mm and ID 16mm(bushing tube)

Drwaing of the component

108

PROCESS PLANNING:

Step 1:

Cutting: In this process firstly, the rhino rod is cut from parent metal as per the desired

length. Secondly, the bushing tube is cut from its parent metal for the length of 2.5in with

the help of cutting machine. And finally a circular disk is cut from the MS plate with the

help of gas cutting machine.

Step 2:

Grinding: In the grinding process the rhino rod is profiled into a curve shape on one side

which matches onto the surface of the bangle pipe and on the other side the rhino rod is

profile grinded into a curve shape which matches onto the bushing tube.

Step 3:

Drilling: In the process of drilling, holes are drilled onto the circular MS disk as per the

rear knuckle with the help of radial drilling machine using drilling bit size as M12.

Step 4:

Welding: In the process firstly all the parts are arranged and assembled in an order and

the welding is done with the help of tungsten inert gas welding machine where first the

bangle pipe is welded onto the circular disk which is already drilled with holes of size

M12 and later the profile grinded parts of the rhino rod are welded onto its specific

match.

Hence, completing the assembly and manufacturing process of a trail arm. Figure shown

below is the sketch of the trail arm after the completion of the above mentioned process.

8. 6 Rear uprights

Material Required



110

Process Planning

1. Stock was obtained 55 x 90 mm

The circular block of aluminum was cut into 150 x 30 mm

2. Lathe setting

The center of the stock and lathe was positioned. Also the rotation of the component

was stabilized.

3. Lathe operation

Turning and facing operations were done

4. Boring

Internal boring was done also the bearing seat was made.

5. Quality control


6. Final Component

8. 7 Rear wheel hubs

Material Required



112

Process planning


The circular block of aluminum was cut into 150 x 30 mm

2. CAM





4. CNC


5. Quality control


113

6. Final Component

9. Introduction to steering system

In a steering gear, the part that is like the bolt is called the worm. The Worm is secured to

the lower end of a shaft with the steering wheel on the opposite end so that the worm and

steering wheel turn together. The steering gear part that is like the section of a nut is

called the sector and its shaft is called the pitman arm shaft. The pitman arm is splined to

the pitman arm shaft. The steering gear worm (bolt) and the sector (nut section) are

machined so that there is very little lash or clearance between their threads in the

midposition. However, as the worm is turned to steer the vehicle either to the right or the

left, the amount of lash increases. This makes up for the unequal wear that occurs in

normal use.

Vehicles are operated in the straight ahead position most of the time, so most of the wear

is in the center of the steering gear worm. It requires 2 1/2 to 3 1/2 turns of the steering

114

wheel and worm to move the pitman arm shaft through its entire allowable movement, an

arc of about 70. That pivots the front wheels from a hard turn in one direction to a hard

turn in the opposite direction. The steering wheel has to be turned farther because of the

mechanical advantage gained by the worm and sector.

Most steering gears are designed so that they provide more mechanical advantage in

the midposition than when turned to the extreme right or left, so they are said to have a

"variable" ratio. Many different kinds of steering gears are used, but they all work in

about the same manner.

10. Types of Steering Gearboxes

There are several types of steering gears available among which few are

1. Worm and sector

2. Worm and roller

3. Cam and lever

4. Recirculating ball

5. Rack and pinion

1. WORM AND SECTOR:

This type of steering gear looks a lot like our bolt and nut, but the sector of this type

looks like a gear instead of a nut. The teeth of the sector are machined in an arc, or curve,

so that they actually look like a section of a gear. As the steering wheel and worm turn,

the worm pivots the sector and pitman arm shaft. The sector pivots through an arc of

70 because it is stopped at each extreme when it touches the steering gear housing. The

worm is assembled between bearings, and some means is provided to adjust the bearings

to control worm end play. The pitman arm shaft is fitted into the steering gear housing on

bearings (generally the bushing type, but roller type bearings are sometimes used). A lash

adjustment screw is also provided so that the sector can be moved closer to, or farther

away from, the worm gear to control the backlash between the sector and worm threads

or teeth.

115

The worm and sector steering gear is very simple in construction. This makes it cheap to

build and easy to maintain. A disadvantage is that it has a lot of friction because of the

sliding action between the worm and sector gear teeth.

2. WORM AND ROLLER:

The worm and roller steering gear is much like the worm and sector, but the sliding

friction is changed to rolling friction so that less effort is required to turn the steering

wheel. This is made possible by machining the sector teeth on a roller. Friction is reduced

even more by mounting the roller on bearings in a saddle at the inner end of the pitman

arm shaft. The worm has an hourglass shape, smaller in the center than at the ends. The

hourglass shape makes the roller stay in better contact with the worm teeth at the ends of

the worm.

3. CAM AND LEVER:

116

In the cam and lever steering gear, the worm is known as a cam. The inner end of the

pitman arm shaft has a lever that contains a tapered stud. The stud engages in the cam so

that the lever is moved back and forth when the cam is turned back and forth. When the

tapered stud is fixed in the lever so that it can't rotate, there is sliding friction between it

and the cam. Therefore, on some vehicles with this type of steering gear, the stud is

mounted in bearings so that it rolls in the cam groove (threads) instead of sliding. Some

large trucks use a cam and twin lever steering gear. This is nothing more than a cam and

lever gear with two tapered studs instead of one. The studs may be fixed in the lever, or

they may be mounted on bearings.

4. RECIRCULATING BALL:

The recirculating ball mechanism which is still found on trucks and utility vehicles. This

is a variation on the older worm and sector design where the steering column turns a large

screw the worm gear which meshes with a sector of a gear, causing it to rotate about its

axis. As the worm gear is turned an arm attached to the axis of the sector moves

the Pitman arm which is connected to the steering linkage and thus steers the wheels. The

recirculating ball version of this apparatus reduces the considerable friction by placing

large ball bearings between the teeth of the worm and those of the screw; at either end of

the apparatus the balls exit from between the two pieces into a channel internal to the box

which connects them with the other end of the apparatus, thus they are recirculated. The

recirculating ball mechanism has a much greater mechanical advantage and hence it is

found on larger, heavier vehicles.

117

5. RACK AND PINION:

In the rack and pinion steering system, the steering gear shaft has a pinion gear on the end

that meshes with a long rack. The rack is connected to the steering arms by tie rods,

which are adjustable to maintain proper toe angle. As the steering wheel is rotated, the

pinion gear on the end of the steering shaft rotates. The pinion moves the rack left and

right to operate the steering linkage. Rack and pinion gears are used on small passenger

vehicles where a high degree of precision steering is required. Their use on larger

vehicles is limited.

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Rack and pinion combinations are often used as part of a simple linear actuator where the

rotation of a shaft powered by hand or by a motor is converted to linear motion. The rack

carries the full load of the actuator directly and so the driving pinion is usually small, so

that the gear ratio reduces the torque required. Rack gears have a higher ratio, thus

require a greater driving torque, than screw actuators.

11. Design of Steering System

11.1 Steering Geometry

The diagram below outlines the important geometry in determining the motions of the

steer wheels in a vehicle that uses Ackerman steering geometry. Ackerman is an

interesting problem because it is dynamic. That is to say that we have two components

moving together – the left and right steering knuckles, but the relationship between their

motions changes as we move them.

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The problem is that we know the distances and are trying to find angle B. We need the

inverse function ARCTAN. Rearranging, we get:

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Plugging the number into a calculator or Excel, or looking up in a table,

ARCTAN (.45) = 24.46°

So, the Ackerman Angle is 24.46 degrees. We can use this to find the length of the tie

rod

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To find out, you have to pick an Ackerman Arm Radius. Choose 6” as an assumption.

We recall that the SIN of an angle is the ratio between the side opposite the angle and the

hypotenuse.

SIN 24.6º = Y/6”

6” * SIN 24.6º = Y

You can look SIN 14.036 up in a table, or punch it up on a calculator, giving you:

6” * .41 = Y

2.46” = Y

Length of tie rod expressed mathematically,

LT = DKC – 2*RAA*SIN Ackerman Angle

Where:

LT is the length of the tie rod

DKC is the distance between kingpins center to center

RAA is the radius of the Ackerman Arm

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LT = 55” – 2*6”*SIN 24.6º

LT = 36” – 2*6”*.41

Therefore,

LT = 50.004”

So, for a car configured as this one is, the tie rod needs to be 31.08” from the center of

one rod end to the center of the other.

Kin

g Pi

n C

ente

r to

Cen

ter

Dis

tan

ce

2

Ackerman

Angle

A B

C

Wheelbase

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So, for a car configured as this one is, the tie rod needs to be 31.08” from the center of

one rod end to the center of the other.

The static values have been set. Let’s contemplate a turn. Let us consider a line drawn

diagonally from point D to B. This creates three angles that add together to give the

angle of the wheel that pivots at point D. We’ll call the first angle K, the second angle γ

(pronounced gamma), and the third angle is of course, the Ackerman angle.

Now we can set to work on determining each. About angle k, we can determine it

because for any steer angle, we know the positions of the ends of the diagonal line. If we

assigned point A the coordinate of 0, 0 then point D would have the coordinates Kingpin

Center to Center Distance, 0. In the case specifically point D’s coordinates would be 55,

0. Point B’s coordinates are found by trigonometric operations. We can calculate it’s

locations with the following formulae:

Point B’s X coordinate = RAA * COS (AA + SAL)

Point B’s Y coordinate = RAA * SIN (AA + SAL)

Where:

RAA is the Ackerman Arm Radius

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AA is the Ackerman Angle

SAL is the steering angle of the left wheel. Zero degrees is straight ahead.

Positive values are a left turn, negative values are a right turn.

Plugging in our numbers for a 20º left turn…

Point B’s X coordinate = 6” * COS (24.6º + 20º)

Point B’s Y coordinate = 6” * SIN (24.6º + 20º)

Point B’s X coordinate = 4.27

Point B’s Y coordinate = 4.21

So, the coordinates of Point B at a 20º left turn are 4.21, 4.21. We can project straight to

the left of point B and straight up from point A to create a new point called point E.

Because we projected straight left and straight up, the angle at E is by definition 90º.

Also, because point E falls on segment AD, we can calculate distance DE with the

formula:

DE = AD – AE

Plugging in our numbers

DE = 55” – 4.21”

Crunching the numbers…

DE = 50.79”

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Now that we know EB and ED, we can find the length of BD because it is a hypotenuse

of the triangle formed. Using Pythagorean Theorem:

BD = (EB2 + (DE)2

Plugging our numbers in…

BD = (50.79”)2 + (4.27”)2

BD = 50.96”

Furthermore, because we know the sides of the triangle we can determine angle k in the

following manner:

TAN k = EB/ED

Of course, we’re trying to find k, so let’s get that by itself by taking the ARCTAN

ARCTAN (EB/ED) = k

Plugging in our numbers…

ARCTAN (4.27”/50.79”) = k

4.80º = k

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So now that we know angle k and the Ackerman angle, we need to find γ (pronounced

gamma). Note that γ is included in triangle BDC. Let’s think about what we know about

this triangle. We know that side DC is the length of the Ackerman arm, which we chose

to be 6”. We know that side CB is the length of the tie rod, which we calculated earlier to

be 31.08”. Finally, we know the distance BD, which we determined using Pythagorean

Theorem to be 50.61”. So we have a triangle and we know the lengths of each of the

three sides.

COS γ = A2 + B2 – C2

2AB

Modifying the relationship, computing for γ we use the ARC COS function to determine

the same, use the values obtained earlier we have

ARCCOS A2 + B2 – C2 = γ

2AB

ARCCOS (50.61)2 + (6)2 – (50.004)2 = γ

2(50.61) (6)

80.77° = γ

Now if we add up angle k, γ and the Ackerman angle, we’ll have the tire’s steer angle

from the line that connects the two kingpins. To get the steer angle, we have to subtract

90°. The formula is:

Steer Angle = k + γ + Ackerman Angle - 90°

Plugging our numbers in…

Steer Angle = 4.80º + 80.77° + 24.6° - 90°

Steer Angle = 20.17°

This engineering calculation shows that the steer angles are 24.6° and 20.17°. The wheels

do not turn parallel, therefore satisfying the Ackerman condition. Additionally,

experience on real cars on the alignment rack indicates that these numbers are reasonable.

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11.2 Collapsible Steering column:

Need: In an event of a frontal collision, due to the forward placement of the steering rack

to accommodate the reverse Ackerman geometry, the driver of the vehicle faces a risk in

a way due to the displacement of the steering rack or gearbox, the whole column

assembly will move into the cabin space, may cause injury to the driver due to

impalement. To prevent this from happening and to increase the safety of the vehicle the

team has developed a collapsible assembly.

Objective:

● To design a cost effective solution to address the issue

● To develop a reliable assembly

Overview: various assemblies exist for achieving a collapsible mechanism that is

actuated by impact, some of the mechanisms that were studied by the team were, using

steel sleds, using tolerance rings, using pneumatic assemblies, but these mechanism do

not satisfy the required parameters, hence the team has gone to taking the use of a shear

pin assembly.

Assembly: The setup carries two mild steel ( AISI 1018 ) rods which are manufactured

to provide a simple telescopic assembly. The length of the collapsing distance is set

to 140mm which would be enough to serve for an impact of more than 17000N, the

length of the column is very flexible. This assembly is press fitted with a nylon

shear pin whose diameter is calculated as per required shear load. The operation of the

column is to provide the required rotational movement to the steering gearbox, to provide

the steering function. In case of an front crash, the nylon shear pin shears at the edges

allowing the inner rod of the assembly to completely collapse within itself.

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12. Fabrication:

The assembly consists of the following components:

● Pigeon block bearing, 30mm dia.

● Outer rod, 30mm Od, 23mm Id.

● Inner rod, 22mm Solid.

● Nylon pin, 18mm Dia.

Process plan:

● Turning of outer rod from 35mm to 30mm.

● Boring of outer rod from 25mm to 23mm.

● Turning of inner rod from 25mm to 22mm.

● Turning of Nylon pin from 20mm to 18mm.

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● Drilling on Outer and Inner rod, 17.5mm.

● Press fitting of Outer rod in bearing.

● Step turning of inner rod to 16mm from 22mm.

● Press fitting of inner rod in universal coupling.

13. Conclusion

With the advent of automobiles, the importance of suspension and steering systems can

never be underestimated; this fact is recognized by the designers of automobiles around

the world. The suspension and steering systems play an integral role in allowing the

vehicle to perform its basic functions. The suspension and steering system go all the way

ahead to ensure the vehicle’s safety during operation, on our part we made an intensive

approach in studying the system in a way to understand and design.

As a part of the design we have process we have worked to achieve our objectives that

were required to excel in the competition. We have made a tedious effort to reduce the

unsprung weight which led us to design most of the assemblies on our own, which while

being a gargantuan task was an experience of learning like no other, the division of

phases, project planning has lead us to understand the atmosphere in a tight work

environment.

The design process was first divided into design of links backed by kinematics, design of

steering system backed by theory of steering by Ackerman, the process proceeded in

design of components that incorporated the geometries obtained by the previous

calculations. The design of components had to be approached in a way that incorporated

machine design, stress-strain analysis and design of life of a component, the component

also involved simulation, both static and dynamic, involved destructive testing, selection

of bearings, material selection and failure analysis. The fabrication has taught us the

world of machining, where workmanship, the skill to operate the machines, CNC part

programming is learnt and applied.

The final manufactured components which needed heat treatment were heat treated

accordingly before assembly. The assembly process included the tolerance and types of

fits, use of allen or hex bolts, torque wrech etc.

To conclude the project, we feel we have achieved our objectives, fabricated innovative

one of a kind ultra light weight components that have been installed in our one of a kind

competition spec vehicle that had made its presence felt in the BAJA SAE 2014

competition at the NATIONAL AUTOMOBILE TESTING AND RESEARCH

FACILITY, INDORE.

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The vehicle was tested on various grounds and strictly scrutinized by experts from

automobile industries. The vehicle was put to a grueling four hour endurance race, where

it was to be driven in very inhospitable terrain testing each and every one of our

components that were installed in the vehicle.

We can proudly state that the vehicle has successfully completed the endurance race

achieving a position of all INDIA 21, among 120 other student built vehicles without a

single component failure.

14. BIBLOGRAPHY

Bajasaeindia.com. (2014). [online] Retrieved from: http://bajasaeindia.com

[Accessed: 2 Apr 2013].

Blundell, M. & Harty, D. (2004). Multibody systems approach to vehicle dynamics.

Oxford: Elsevier Butterworth-Heinemann.

Ellis, J. R. (1969). Vehicle dynamics. London: Business Books.

Khurmi, R. S. & Gupta, J. K. (2003). Theory of machines. New Delhi: S. Chand &

Co.

Khurmi, R. S. & Gupta, J. K. (1996). A text book of machine design. New Delhi:

Eurasia.

Lingaiah, K. (2002). Machine design data handbook. New York: Mcgraw-Hill.

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Milliken, W. F. & Milliken, D. L. (1995). Race car vehicle dynamics. Warrendale,

PA, U.S.A.: SAE International.

Pirate4x4.com. (2014). Pirate4x4.Com - The largest off roading website in the

world.. [online] Retrieved from: http://pirate4x4.com [Accessed: 2 Jun 2013].

Shigley, J. E. (1956). Machine design. New York: Mcgraw-Hill.

Singer F. L. (1994). Engineering mechanics statics and dynamics. New Delhi: Indus.

Smith, C. (1978). Tune to win. Fallbrook, CA: Aero Publishers.

Smith, C. (1984). Engineer to win. Osceola, Wis.: Motor books International.

B. J. Lazan and A. A. Blatberwiek, “Fatigue Properties of Aluminum Alloys at

Various Direct Stress Ratios”, WADC technical report 52-307 part 1, University of

Minnesota, December, 1953

The presence of chromium bearing segregated doesn’t reduce the fatigue strength.-

Douglas Aircraft Company

G.T.Yahr, “Fatigue Design Curves for 6061-T6 Aluminum”, Oak Ridge National

Laboratory, 1993

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PSG College “ Design Data Book” Kalakhathir, Coimbatore,2010

Design,Analysis & Fabrication of suspension of all terrain vehicle

Engineering

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