Design,Analysis & Fabrication of suspension of all terrain vehicle

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Table of content
SUSPENSION
Page
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.

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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 = 𝑐 𝑚 ∗
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
SPECIFICSTRENGHTNM/KG

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
SPECIFICCOSTRSKG/MN
0
5
10
15
20
25
30
ELONGATION%

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 ∗ 𝑎 𝑑 ∗ 𝑑
whereas,
v ꞊ final velocity
thus,
𝑎 𝑑꞊ 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.

118
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.

119
The problem is that we know the distances and are trying to find angle B. We need the
inverse function ARCTAN. Rearranging, we get:

120
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

121
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

122
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.
KingPinCentertoCenterDistance
2
Ackerman
Angle
A
B
C
Wheelbase

Design,Analysis & Fabrication of suspension of all terrain vehicle

Design,Analysis & Fabrication of suspension of all terrain vehicle

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