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Analysis and Improvement of the

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Dr. S. Neelakrishnan. et al. Int. Journal of Engineering Research and Application www.ijera.com
ISSN : 2248-9622, Vol. 7, Issue 5, ( Part -4) May 2017, pp.18-25

RESEARCH ARTICLE OPEN ACCESS

Analysis and Improvement of the Steering Characteristics of an

ATV.
1.
Dr. S. Neelakrishnan, 2. Kowshik T, 3. Krishnakumar G,
4
. Bharathi Mohan M P
1, 2, 3,
4 (Department of Automobile Engineering, PSG College Of Technology, Coimbatore-04

ABSTRACT
The main objective is to analyze the steering characteristics of an ATV (All Terrain vehicles) in order to improve
the maneuverability of the vehicle. It is found that in our last year’s BAJA buggy the vehicle oversteerd more
than we expected, when inputs are given in the steering wheel while cornering at sharp turns which tends to
move the vehicle out of the track. We focus to design an effective steering system with a reduced steering ratio
of 3:1. This helps the driver to maneuver the vehicle with ease.

I. INTRODUCTION II. DESIGNED PARTS OF STEERING

A Steering gear box helps to steer the front SYSTEM:
wheels of the vehicles which includes commercial
vehicles, passenger vehicles, an ATV (All Terrain
vehicles), etc. There are various types of steering
gear boxes among which rack and pinion gear is
considered to be more precise as it consists of fewer
parts easier to control, respond to the inputs given
and less backlash. The rack is housed in a casing.
The casing is supported on the frame near
its ends. The ends of the rack are connected to the
track rod with the help of a ball joint or Heim joint. Fig 1: Parts of steering system.
The pinion shaft is carried in the bearings housed in
a casing. The pinion meshes with the rack. Its main 1. Steering wheel 2. Steering column
objective is to convert the rotational motion of the 3. Steering gearbox 4. Tie rod
steering wheel to the linear motion of the rack. 5. Steering Upright.
While the driver turns the steering wheel the column
rotates the pinion gear which in turn meshes with the
Design Methodology for Steering System:
rack, then the linear motion is transferred to the
drive wheels via the track rod.
The steering ratio is the main criteria for
designing the steering system. It is nothing but the
number of degrees of rotation of the steering wheel
to the number of degrees of rotation of road wheels.
The steering ratio varies based on the applications,
which in changes the diameter of the steering wheel
accordingly. Generally, the steering ratio for
passenger cars is 14:1. In case of heavy vehicles it is
20:1. The number of degrees of rotation of steering
wheel increases with increase in steering ratio. Rack
and pinion steering is quickly becoming the most
common type of steering on cars, small trucks and
SUVs.

Fig 2: Design methodology flow chart.

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Dr. S. Neelakrishnan. et al. Int. Journal of Engineering Research and Application www.ijera.com
ISSN : 2248-9622, Vol. 7, Issue 5, ( Part -4) May 2017, pp.18-25

The main objective of this project is to

design a steering system based on the flow chart
shown in fig 1. The design process starts with the
road wheel angles, Ackerman geometry, the
minimum lock to lock, steering effort, steering arm
length, Upright and Tie rod location.
Ackermann steering plays an important in a
steering system as it enables the vehicle to steer the
vehicle without slippage of the tire. Secondly, the
minimum lock to lock of the steering wheel is kept
by optimizing the steering ratio of 3:1 in order to Fig 3: CAD modelling of rack.
have better maneuverability for a BAJA buggy. 2.2 Pinion:
Then followed by steering effort and wheel Pinion gear is designed with 23 teeth with a travel of
alignment parameters are considered. Wheel 270 degree from lock to lock.
alignment parameters include camber, caster,
steering axis inclination, Toe-in and Toe-out. The
angles of all the wheel alignment parameters are
specified based on the requirement of BAJA buggy.
Hence, on the whole the dynamic behavior and
stability of the vehicle is improved by studying the
steering characteristics.

Design Process, Analysis And Simulation:

The complete CAD modelling is done in
solid works software, including Rack and pinion
gear box, steering upright, Tie rod, steering column
and steering wheel. It is important to carry out Fig 4: CAD Modelling of pinion.
analysis for all the custom made parts. Ansys
software is used for analysis to check the component 2.3 Rack and pinion gearbox casing:
strength, fatigue life which includes components Casing is designed to cover the rack and
factor of safety, equivalent stress and total pinion assembly. Nylon bushings are provided at
deformation to ensure safety design standards. both the ends of casing for rack sliding. Ball Bearing
After carrying out analysis, simulation is is placed at the center to enhance the rotatory motion
carried out with the help of Lotus shark software by of the pinion. Backlash adjustment is provided at the
generating the coordinates from the CAD assembly. bottom side of the casing to avoid the free play.
The dynamic behavior of wheel alignment
parameters, bump steer and Ackerman percentage is
checked with respect to steer travel. The graphs were
generated after simulation.

Design Caculations:
1.1.1 Rack and pinion gear Box:
Parameter Rack Pinion
No. of Teeth 22 23
Length 356mm 52mm
Diameter 20mm 34mm
Pressure Angle 20⁰ 20⁰
Fig 5: CAD modelling of rack and pinion Gear box
Travel 4.25” 270 deg casing.
Table 1: Rack and Pinion Specifications
The casing also incorporates the mounting
2. Steering Components Cad Modelling: holes of 8mm diameter of 4holes for rigid mounting.
2.1 Rack: The casing is supposed to made-up of the
Cad modelling is done for rack with 4.25” Aluminium materials for weight reduction.
rack travel. It consists of 22 teeth. Heim joints are
fixed at both the ends of the rack to connect the tie
rod. The total length of rack travel is 14”.

www.ijera.com DOI: 10.9790/9622-0705041825 19 | P a g e

Dr. S. Neelakrishnan. et al. Int. Journal of Engineering Research and Application www.ijera.com
ISSN : 2248-9622, Vol. 7, Issue 5, ( Part -4) May 2017, pp.18-25

2.4 Rack and pinion assembly: = - 700*45*sin (10) *sin (36) -710*45*sin (10)
*sin (45) +700*45*sin (10) *cos (36) -710*45*sin
(10) *cos (45)
= -6646Nmm

LATERL MOMENT:
ML = (Fyl + Fyr)*r*tan Ѵ
Fyl = µFzl = 0.45*700 = 315N
Fyr = µFzr = 0.45*710 = 319N
Fig 6: Rack and pinion assembly. = (315+319) *285*tan (10)
= 31860Nmm
2.5 Final assembly of Steering gear box:
TRACTIVE MOMENT:
Mt = (Fxl – Fxr) *d
= (710-700)*45 = 450Nmm
EQUIVALENT TORQUE:
Equivalent torque= 28963Nmm
STEERING EFFORT:
Fig 7: Steering gear box assembly. Steeringeffort=EquivalentTorque/(steeringwheel
radius*S. R)
= 28963/(150*3)
III. CALCULATIONS
= 65N.
3.1. 1. Specifications:
Wheel base 1561mm 3.1.3. PINION DESIGN AND CALCULATION:
Front track 1350 mm Pinion Torque = Equivalent Torque/Steering ratio
Rear track 1244mm = 28963/3
Turning circle radius(g) 1900mm = 9654Nmm
Lock to lock angle(Steering 360 degrees Ft = Pinion Torque/(PCD/2)
wheel)
Diameter of the steering wheel 300 mm
= 9654/23
Rack travel 107.95mm = 419N.
Tie rod length 320 mm
Steering effort 65N Bending Stress σb = Ft*q.k*q.e / b*m
Steering ratio 3:1 = 419*1*2.7/ 3.5*2
Caster angle 10 deg
Kingpin inclination 10 deg
= 162 N/mm2 .
Table 2: Specifications
Design bending stress (Tension):
3.1.2. Steering Effort:
VERTICAL MOMENT:
Mv = - (Fzl+Fzr)dsin γsinб + (Fzl-Fzr)dsin Ѵcos б x = (t/m – pi/2)/(2*tanα)
Dynamic load = FAW+L. T Where, t = thickness of teeth.
= (210*0.45*9.81) + m = module.
(w*a*h/g*L) α = pressure angle.
= (927)+(283) x = (3.5/2 – 3.14/2)/(2*tan(20))
= 1410 N = 0.25
Fzr = 710N , Fzl = 700N = 1.3
LEFT TURN: = 0.35
Mv = - Fzl d sinγ sinбi - Fzr d sinγ sinбo +
Fzl d sinγ cosбi - Fzr d sinγ cosбo
= 3211.005 + 1200
= -700*45*sin(10)*sin(45)- = 4411.005 kgf/
710*45*sin(10)*sin(36) [ = 1696 kgf/
+700*45*sin(10)*cos(4)-
710*45*sin(10)*cos(36) = 170 N/
= -7749.5Nmm <[ ]
RIGHT TURN: So, the material is safe.
Mv = - Fzl d sinγ sinбo - Fzr d sinγ sinбi +
Pressure angle α= 20 deg
Fzl d sinγ cosбo - Fzr d sinγ cosбi
Module m = 2
Number of teeth Z1 = 23

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Dr. S. Neelakrishnan. et al. Int. Journal of Engineering Research and Application www.ijera.com
ISSN : 2248-9622, Vol. 7, Issue 5, ( Part -4) May 2017, pp.18-25

Pitch circle dia(d1) = Z1*m axle mounting point from its center. The main
= 23*2 objective is to reduce the static inclination in the
= 46mm. front side of the vehicle. This reduces the roll center
Tip circle dia = d1+2m height. The rear side roll center height is higher than
that of the front. While cornering the vehicle, due to
= 50mm
the elastic and geometric load transfer in the front
Base circle dia = d1cosα and rear side of the vehicle tends to increase the slip
= 46*cos(20) angle in the rear. Hence over steer is achieved.
= 43.22mm
Tooth thickness on Pitch circle = ∏m/2 4.1.1 Analysis by Ansys:
= 3.14mm. 1. Steering Upright
Step 1: Generating mesh for the upright:
3.1.4. STEER CONDITION:
(Wf/Cxf) -(Wr/Cxr) = K (K=understeer
gradient)
Wf = 882.9N Wr = 1569. 6N
Cxf = 155. 75N/deg Cxr =
267N/deg =
(882.9/155.75) - (1569.6/267).
= - 0.21 (negative).

Since Wf < Wr (oversteer condition).

In this case, the lateral acceleration at the CG
causes the slip angle on the rear wheel to increase Fig 8: Meshing of upright
than the front wheels. The outward drift at the rear
of the vehicle turns the front wheel inward, thus Step 2: Fixed support is given at two points:
diminishing the radius of turn.

3.1.5. CRITICAL SPEED:

In oversteer case, a critical speed will exist above
which the vehicle will be directionally unstable.
V=
= 65m/s.
Long wheelbase vehicles will have a higher
critical velocity than the shorter wheelbase vehicles.
3.1.6. YAW VELOCITY GAIN:
Yaw velocity r, is the rate of rotation in the heading
angle Fig 9: fixed support is given
r = 57. 3V/R
Where V-forward speed. Step 3: Bearing load of 4500N is applied at the
R-radius of turn. steering arm:
r = (57.3*65*5)/(18*1765.3)
= 0.586deg/Sec.
Yaw velocity gain is defined as the ratio of steady
state yaw velocity to the steer angle.

3.1.7. CURVATURE RESPONSE GAIN:

CR= (1/L) /(1+K*v2/Lg)
= (1/1651)/(1+0.21*65^2/1651*9.81)
= 0. 0121deg/Sec

IV. UPRIGHT DESIGN AND ANALYSIS:

4.1. Steering Upright:
Fig 10: bearing load is applied at steering arm
The Steering Upright is designed in such a
way that Provision is given for mounting the control
arms, Tie rod, Brake calipers and space for housing
the wheel assembly with the help of Stub axle. The
important feature in this upright is offsetting of stub

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Dr. S. Neelakrishnan. et al. Int. Journal of Engineering Research and Application www.ijera.com
ISSN : 2248-9622, Vol. 7, Issue 5, ( Part -4) May 2017, pp.18-25

Step 4: Bearing load of 18000N is applied to the

spindle mount point considering one wheel
landing:

Fig 15: safety factor

2. Analysis for rack and pinion

Step 1: Remote point is applied at the surface of the
Fig 11: bearing load applied at spindle mounting pinion
point.

Step 5: A moment of 140 Nm is applied at the

caliper mounting point:

Fig16: Applying remote point at the rack and pinion

Step 2: Meshing of Rack and pinion:

Fig 12: load applied at caliper mounting point.

Step 6: The total deformation is found out by

solving the upright:

Fig17: meshing of rack and pinion

Step 3: Remote displacement is applied:
Fig 13: total deformation of knuckle

Step 7: Equivalent stress is found out:

Fig18: Remote displacement is applied at the pinion

surface
Fig 14: Equivalent stress
Step 8: Finally safety factor is checked:

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Dr. S. Neelakrishnan. et al. Int. Journal of Engineering Research and Application www.ijera.com
ISSN : 2248-9622, Vol. 7, Issue 5, ( Part -4) May 2017, pp.18-25

Step 4: No separation is applied along the surface Step 7: Equivalent stress

of the rack

Fig19: No separation is applied along the rack Fig22: Equivalent stress of rack and pinion
surface
Step 8: Factor of safety
Step 5: Frictionless support is applied at the rack
surface:

Fig23: Factor of safety

V. LOTUS SHARK ANALYSIS:

Fig20: Frictionless support is given at the rack
surface
Step 6: Force is applied along the z-axis to the
rack:

.
Fig 16: Steering geometry

Fig 7 shows the geometry of steering with

double wishbone with lower damper and simulated
for 300mm arms travel for better camber, toe and
caster angle variations
Fig21: Force applied along the z-axis to the rack

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Dr. S. Neelakrishnan. et al. Int. Journal of Engineering Research and Application www.ijera.com
ISSN : 2248-9622, Vol. 7, Issue 5, ( Part -4) May 2017, pp.18-25

5.1.1 Lotus Shark Points-Front:

Double wishbone Damper to Lower Wishbone -
Static values.
Points. X(mm) Y(mm) Z(mm)
Lower wishbone -1003.4 -124.4 -549.4
front pivot
Lower wishbone -759.6 -124.8 -588.9
rear pivot
Lower wishbone -844.2 -552.8 -651.7
outerball joint
Upper wishbone -985.9 -145.8 -459.7
front pivot
Fig 19: Bump steer vs toe angle
Upper wishbone -748 -145.85 -498.2
1.
rear pivot
The first graph represents the steer travel vs
Upperwishbone -828.5 -533.8 -554.8 toe angle. The inner wheel turns at an angle of 45deg
outerball joint and outer wheel turns at angle of 35deg during full-
Damper wishbone -837.8 -329.7 -586.5 lock.The high steering angle of 45deg is used to aid
end in the sharp turns.It is also made sure that the tie rod
Damper body end -916.7 -584.5 -594.5 does not hit the upright,suspension arms and the
Outer track rod -954.3 -134.4 -503.2 wheel rim during these manuers.It is also ensured
ball joint that brake hose does not bend beyond certain limit
Inner track rod -773.8 -217.8 -190.1 thus brake hose having only very minimal damage.
ball joint 2. 45% Ackermann is obtained during full lock of
Upper spring -837.9 -329.7 -586.5 steering wheel.A configuration in between parallel
pivot point steering configuration and pure Ackermann
Lower spring -889.5 -551.1 -662.9 configuration is used as there is considerable
pivot point difeerence in the slip angle of the inner and the outer
Wheel spindle -889.5 -646.5 -662.9 wheels. The percentage of Ackermann and the
point amount of bump steer were also determined for
Table 3: Lotus shark point. different rack travels.
Bump steer will be low when the suspesion
3.1.8. Graphs after simulation: arms and the tie rod are parallel.In our case tie is not
exactly parallel to the suspension arms.The outer tie
rod end is made to move in arc with inner tie rod end
as its Locus.Analysis were made using Lotus
software for minimal change in toe when the vehicle
hits the bump.The graph shows that the angle is 3deg
for 150mm of bump travel.

VI. CONCLUSION
The main motivation behind this project is
to rectify the faults experienced and to enhance the
steering characteristics of a BAJA buggy used for
Fig 17: Steer travel Vs Toe angle engineering competitions.
In order to enhance the steering
characteristics, a Steering gear box and steering
upright is designed. The designed parts are analysed
and finally simulation is carried out.
The values obtained after the design process i.e, such
as analysis and simulation is desierd to run the
BAJA buggy efficiently.

REFERENCES
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Dr. S. Neelakrishnan. et al. Int. Journal of Engineering Research and Application www.ijera.com
ISSN : 2248-9622, Vol. 7, Issue 5, ( Part -4) May 2017, pp.18-25

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